Design and fabrication of gearbox with inboard braking of an all terrain vehicle

DESIGN AND FABRICATION OF GEARBOX WITH
INBOARD BRAKING OF AN ALL TERRAIN VEHICLE
A major project report submitted to
Jawaharlal Nehru Technological University Hyderabad
in partial fulfillment of the requirements for the award of the degree of
BACHELOR OF TECHNOLOGY
in
MECHANICAL ENGINEERING
Submitted By
MOHAMAD ABDUL : 14WJ1A03H2
Under the Guidance of
Mr. V.ANANDA MOHAN
Associate Professor
DEPARTMENT OF MECHANICAL ENGINEERING
GURU NANAK INSTITUTIONS TECHNICAL CAMPUS
(Affiliated to JNTU, Hyderabad, Approved by AICTE, New Delhi)
Ibrahimpatnam, Ranga Reddy District -501506
Telangana, India.
2017-2018

CERTIFICATE
This is to certify that the major project entitled “DESIGN AND
FABRICATION OF GEARBBOX WITH INBOARD BRAKING OF AN ALL
TERRAIN VEHICLE” is being submitted by Mr. MOHAMAD ABDUL
(14WJ1A03H2) in partial fulfilment for the award of the Degree of Bachelor of
Technology in Mechanical Engineering to the Jawaharlal Nehru Technological
University Hyderabad is a record of bonafide work carried out by them under my
guidance and supervision.
The results embodied in this Project report have not been submitted to any
other University or Institute for the award of any Degree or Diploma
Internal Guide
Mr. V. ANANDA MOHAN
Associate Professor
External Examiner Head of the Department
(MECHANICAL ENGINEERING)

DECLARATION
I declare that this Project report titled “DESIGN AND FABRICATION OF
GEARBBOX WITH INBOARD BRAKING OF AN ALL TERRAIN
VEHICLE” submitted in partial fulfilment for the award of the Degree of Bachelor
of Technology in Mechanical Engineering to the Jawaharlal Nehru
Technological University Hyderabad is a record of original work carried out us
under the guidance of Mr. V.ANANDA MOHAN, Associate Professor,
Department of Mechanical Engineering, and has not formed the basis for the award
of any other degree or diploma, in this or any other Institution or University. In
keeping with the ethical practice in reporting scientific information, due
acknowledgements have been made whenever the findings of others have been cited.
MOHAMAD ABDUL
(14WJ1A03H2)

ACKNOWLEDGEMENT
I wish to express our sincere thanks to Dr. H.S. SAINI, Managing Director,
Guru Nanak Institutions and Dr. M. RAMALINGA REDDY, Director, Guru Nanak
Institutions Technical Campus, School of Engineering and Technology, for providing
us with all the necessary facilities and their support.
I wish to express our sincere thanks to Dr. G. SANKARANARAYANAN,
Professor & Dean, Department of Mechanical Engineering for his valuable
suggestions in the project.
I place on record our sincere thanks to Dr. M. HARINATHA REDDY and
Dr. A. RAJ KUMAR, Professors and Head of the Department, Mechanical
Engineering for their whole-hearted co-operation, providing excellent lab facility,
constant encouragement and unfailing inspiration.
I would like to say sincere thanks to Dr. S. NAGAKALYAN, Professor,
Department of Mechanical Engineering for Co-ordinating Projects
I especially thank our internal guide Mr. V. ANANDA MOHAN, Associate
Professor, Department of Mechanical Engineering for the suggestions and constant
guidance in every stage of the project. we also like to thank all of our lecturers
helping us in every possible way.
On a more personal note we thank our BELOVED PARENTS and
FRIENDS for their moral support during the course of our project.
MOHAMAD ABDUL : (14WJ1A03H2)

ABSTRACT
We design, build and test an off-road race vehicle each year. During this season, a
customised gearbox with inboard braking is typically designed as a reduction after the
Continuously Variable Transmission.
A preliminary design was first prepared keeping in the mind, the guidelines and rules
issued by SAE. Indian standards for driver space and ergonomic preference
transmission type was selected and designed according to requirements of the driver.
Hence after ensuring safety, the design was finalised and fabricate it.
Moreover, the axle braking is to slow down or stop the vehicle safely and effectively
by converting kinetic energy into heat. It is the one of the main safety of a vehicle.
The goal of the project is to develop a lightweight, compact gear reduction that will
increase the efficiency and durability of the vehicle. And also, in addition we have
provided hydraulic braking for the drive shaft (i.e. known as inboard braking) for
which the calliper is mounted to the gearbox itself.
The pedal must actuate the master cylinder without any usage of cables and should be
capable of locking four wheels in static condition and dynamically on paved and
unpaved surfaces.
While designing a brake system for a vehicle that can produce adequate braking force
to meet competition regulations while being as light as possible. A budget, timeline,
proof of design, fabrication and testing will also incorporate in the report.

vi
TABLE OF CONTENTS
DESCRIPTION PAGE NO
CERTIFICATE ii
DECLARATION iii
ACKNOWLEDGEMENT iv
ABSTRACT v
LIST OF FIGURES x
LIST OF TABLES xii
ABBREVIATIONS/NOMENCLATURE xiii
1. INTRODUCTION 1
1.1. Overview of the project 1
1.2. Objective 2
1.3. Limitations 4
2. LITERATURE SURVEY 6
2.1. Terms used in gears 6
2.2. Introduction to design of gears 8
2.2.1 Gear Failure 8
2.2.2 Gear noise and vibration 9
2.2.3 Geometrical modification of gears 10
2.2.4 Establishment of gear test rig 12
2.3. Introduction to design of shafts 14
2.3.1 Material used for shafts 15
2.3.2 Manufacturing of shafts 15
2.3.3 Types of shafts 15
2.3.4 Standard sizes of transmission shafts 16
2.3.5 Stresses in shafts 16

vii
2.3.6 Maximum permissible working stresses for
transmission shafts 16
2.3.7 Design of shafts 17
2.4 Introduction to braking system 17
2.4.1 Functions of braking system 17
2.4.2 Working of brakes 18
2.4.3 Types of braking system 18
2.4.3.1 Introduction to hydraulic system 19
2.4.3.2 Disc Brake 21
2.4.4 Inboard Braking 22
2.4.4.1 Rotors 22
2.4.4.2 Calipers 24
2.4.4.3 Master cylinder 24
3 DESIGN CONSIDERATIONS OF POWERTRAIN 26
3.1 Introduction to powertrain 26
3.2 Transmission 27
3.2.1 Goal 28
3.2.2 Analysis of CVT system 29
3.2.2.1 CVT set-up 29
3.2.2.2 Calculations 32
3.3 Modelling of gears 34
3.4 Modal Analysis 36
3.4.1 Structural static analysis 36
3.4.2 Loads in a static analysis 36
3.4.3 Linear Vs Non linear static analysis 37
3.5 Overview of steps in a static analysis 37
3.6 Static analysis 38
3.6.1 FEA Results of gears 45
3.7 Technical Data 46
3.7.1 Input shaft dimensions 46
3.7.2 Output spline data 47
3.8 Gear box casing 47

viii
3.9 Rendered images 49
4 DESIGN CONSIDERATIONS OF BRAKING
SYSTEM 50
4.1 Brake Torque 50
4.2 Clamping force 50
4.3 Coefficient of friction 50
4.4 Thermal capacity 51
5 FABRICATION OF GEARBOX 58
5.1 Classification of manufacturing process of gears 58
5.2 Methods of forming gears 59
5.2.1 Roll forming 59
5.2.2 Stamping 59
5.2.3 Powder metallurgy 60
5.2.4 Extrusion 60
5.3 Gear generating process 60
5.3.1 Gear hobbing 60
5.3.2 Types of hobbing 60
5.3.2.1 Arial hobbing 60
5.3.2.2 Radial hobbing 61
5.3.2.3 Tangential hobbing 61
5.4 Gear shaping process 61
5.4.1 Rack type cutter 61
5.4.2 Pinion type cutter 62
5.4.3 Advantages 62
5.4.4 Disadvantages 62
5.5 Gear cutting by milling 62
5.5.1 Disc type cutter 62
5.5.2 End milling cutter 63
5.5.3 Advantages 63
5.6 Bevel gear generating 63
5.6.1 Straight bevel gear generator 64
5.6.2 Spiral bevel gear generator 64
5.6.3 Gleason method 64

ix
5.7 Gear finishing process 65
5.7.1 Gear shaving 65
5.7.2 Gear grindings 65
5.7.3 Shot blasting 66
5.7.4 Phosphate coating 66
5.8 Gear Planning 66
5.8.1 The sunderland process 66
5.8.2 The maag process 67
5.8.3 Principle of gear planning 67
5.8.4 Photographs taken while fabrication 67
5.9 Installation to the vehicle 69
6 RESULT & DISCUSSION 70
7 CONCLUSION 72
REFERENCES 73
APPENDIX 75

x
LIST OF FIGURES
FIGURE TITLE PAGE NUMBER
1.1 Mini Baja race of ESI 2018 competition 1
1.2 Picture of transmission system in Drag race 2
2.1 Normal and modified gear profile models 10
2.2 Selective reinforced gear with spline hub 11
2.3 Position of circular fillet 12
2.4 Simple construction of gear test rig 13
2.5 Test rig for investigating gear noise emission 14
2.6 Simple braking system 18
2.7 Hydraulic system 19
2.8 Pascal’s Law 19
2.9 Construction of hydraulic braking system 20
2.10 Working of hydraulic braking system 21
2.11 Disc brake 21
2.12 Integrated brakes 22
2.13 Inboard brakes 22
2.14 Brake rotor 23
2.15 Brake rotor FEA 23
2.16 Pulsar 220F caliper 24
2.17 Working of Tandem master cylinder 25
2.18 Master cylinder with reservoir 25
3.1 Torque Vs Engine RPM 26
3.2 Power Vs Engine RPM 27
3.3 Free body diagram of the vehicle 28
3.4 Basic concept of CVT Drive train system 30
3.5 3-D Drawing of CVT Drive train system 31
3.6 Simple depiction of the layout of the rear of
the frame and prospected optimal placement
of the engine
31
3.7 Motor Torque Curve 32
3.8 First stage reduction 35
3.9 Second stage reduction 35

xi
3.10 Showing the model pinion gear 39
3.11 Meshing of pinion gear 39
3.12 Loading scenarios on pinion 40
3.13 Equivalent stress on pinion gear 1 40
3.14 Total deformation 41
3.15 Intermediate gear 41
3.16 Meshing of intermediate and output gears 42
3.17 Loading scenarios 42
3.18 Output gear 43
3.19 Intermediate gear 43
3.20 Output gear 44
3.21 Input shaft and gear 44
3.22 Analysis on intermediate gears 45
3.23 Analysis on output gear 45
3.24 Input shaft 46
3.25 Output shaft(Rzeppa Joint) 47
3.26 Side view of the gear box 48
3.27 Rear view of the gear box 48
3.28 Rendered image-I 49
3.29 Rendered image –II 49
5.1 Raw material 67
5.2 CNC milling 68
5.3 Pinion 68
5.4 Gearbox casing under machining 68
5.5 Gear box installed to the vehicle 70

xii
LIST OF TABLES
TABLE TITLE PAGE NUMBER
2.1 Mechanical properties of steels used for
shafts
15
3.1 This table displays our numerical data as
it relates to our assumptions and the
equations.
33
3.2 This table displays our numerical data as
it relates to our assumptions and the
equations with a slight change in ratio.
33
3.3 Overall gear features 46
3.4 Output shaft data 47

xiii
ABBREVATIONS
SAE Society of Automotive Engineers
ESI Enduro Student India
OHV Over Head Valve
OEM Other Externally Manufacturer
TE Transmission Error
ANSI/AGMA American Gear Manufacturer’s Association
FEA Finite Element Analysis
ASME American Society of Mechanical Engineers
B&S Briggs & Stratton
RR Rolling Resistance
AR Air Resistance
GR Grade Resistance
TTE Total Tractive Effort
CVT Continuous Variable Transmission
ANSYS Simulation software
IGES Initial Graphics Exchange Specification
DOT Department of Transport
MC Master Cylinder
ISO Indian Standards Organization
ⱷ Pressure angle
Pc Circular pitch
D Diameter
T No. of teeth
Pd Diametral pitch
m Module
dB Decibel
σt Permissible tensile stress
σu Ultimate tensile strength
σel Elastic limit in tension
τ Permissible shear stress
μ Coefficient of friction

xiv
fr Coefficient of rolling resistance
Rcvt CVT ratio
Ncvt Efficiency of CVT
Rr Reduction ratio
TBr Braking torque
FF Frictional force on front tires
Ff Vertical force on front tires
We weight of car
Xcg distance from front axle to car’s centre of gravity
I Wheelbase
Ycg height above ground of car’s centre of gravity
CF clamping force
PM maximum hydraulic pressure
AT total effective area of caliper piston
TBd Brake torque developed
µL coefficient of friction between brake pads and rotors
Re effective rotor radius
Fp force on master cylinder piston
AP area of the MC piston
Dp Diameter of the MC piston
K kinetic energy
TR temperature rise
Wb weight of all rotors
Kb kinetic energy before stop
Ka kinetic energy after stop
Kc kinetic energy change

xv
NOMENCLATURE
Temperature ºC, ºF
Angle º
Force Newton(N), lbs
Speed Mph. kmph
Torque lb-ft, N-m
Power HP, KW
Weight lbs, Kgs
Diameter Inches, in
Length m, cm, mm
Efficiency %
Pressure psi, Pa
Energy ft.lbs
Volume ml

1
CHAPTER 1
INTRODUCTION
1.1 Overview of the project
Baja SAE is an inter collegiate design competition run by the Society of
Automotive Engineers (SAE). Teams of students from universities all over the world
design and build small off-road cars. The cars all have engines of the same
specifications. As of 2018 the engine has been an unmodified Briggs &
Stratton Model 19 Vanguard engine single-cylinder with a displacement of 305cc and
power output of approximately 10 bhp (7.5 kW).
The goal in Baja SAE racing is to design, build and race off-road vehicles that
can withstand the harshest elements of rough terrain. The vehicles used in Baja SAE
racing are often similar in appearance to dune buggies. Before 2007, the events were
called "Mini Baja."
Figure 1.1 Mini Baja race of 2018 ESI Competition
Each year as many as 141 Baja cars are entered in the Baja SAE events across
the US and around the world where events are held including India, China, Brazil,
South Africa and Korea. In India, this event is run by SAE India. In China, this event
is run by SAE China starting in 2015. All cars must adhere to SAE's rules, and pass
SAE's technical inspection and judging; a car may not race until all safety inspections
are passed. Small engine manufacturer Briggs & Stratton sponsors Baja SAE teams
by providing the SAE sanctioned engine free of charge, at a replacement rate of one
engine for every two years in competition.

2
Figure 1.2 Picture of transmission system in drag race
There are multiple dynamic events, usually four per event, as well as a single
four-hour endurance race. The dynamic events include hill climbs, sled pulls,
maneuverability events, rock crawls, and suspension & traction events. Previously the
cars had to be able to float and propel itself on water under its own power. This was
changed from the 2012 competitions onward due to safety concerns.
Static events, such as written reports, presentations and design evaluations are
provided by participating teams. This is when the teams are judged on ergonomics,
functionality, and producibility of their cars; ensuring that the final placement of the
team does not rest solely on the vehicle's performance but rather on a combination of
static and dynamic events. Required reports detail the engineering and design process
that was used in developing each system of the team’s vehicle, supported with sound
engineering principles.
Also, a cost report that provides all the background information necessary to
verify the vehicle’s actual cost is used to rate the most economically feasible for
production. These reports are submitted weeks in advance of each event, where the
presentations and design evaluations are given on site in the presence of SAE design
judges.
1.2 Objective
Baja SAE is an intercollegiate engineering design competition for undergraduate
and graduate engineering students. The object of the competition is to simulate real-

3
world engineering design projects and their related challenges. Each team is
competing to have its design accepted for manufacture by a fictitious firm. The
students must function as a team to design, build, test, promote and compete a vehicle
within the limits of the rules, also to generate financial support for their project and
manage their educational priorities.
Each team's goal is to design and build a prototype of a rugged, single seat, off-
road recreational vehicle intended for sale to the non-professional weekend off-road
enthusiast. The vehicle must be safe, easily transported, easily maintained and fun to
drive. It should be able to negotiate rough terrain without damage. As of 2010, the
SAE Baja Series consisted of three competitions, though in the past there have been
as many as seven sanctioned events. A Baja SAE competition event consists of three
to four days.
Several factors contribute to make a winning buggy. First and foremost the
buggy has to meet strict specifications of the rule book. The philosophy hasn’t
changed since the event’s birth back 1976 – the teams still need to build a simple all
terrain vehicles for recreational purpose that is aesthetically and ergonomically sound
while still being a fun and durable machine in the real world conditions. The
evaluation process for BAJA SAE INDIA is a twofold process and students have to
clear the virtual preliminary round before they start manufacturing their buggies for
the main event.
Our primary goal of the project is to develop a lightweight, compact gear
reduction that will increase the efficiency and durability of the vehicle. And also, in
addition we have provided hydraulic braking for the drive shaft (i.e. known as inboard
braking) for which the caliper is mounted to the gearbox itself.
The pedal must actuate the master cylinder without any usage of cables and
should be capable of locking four wheels in static condition and dynamically on
paved and unpaved surfaces.
While designing a brake system for a vehicle that can produce adequate braking
force to meet competition regulations while being as light as possible. A budget,
timeline, proof of design, fabrication and testing will also look in the report.

4
1.3 Limitations of the project
Since the vehicle was being made for a competition, there are rules and
regulations to be followed. There are no other limitations on various aspects of the
vehicle which will be discussed further in the coming topics:
Limitations are as follows:
1.3.1 Selection of engine [1]
Rule B2.4 Engine Requirement and Restrictions (NEW)
To provide a uniform basis for the performance events, all vehicles must use the same
engine: a stock four cycles, air cooled, Briggs & Stratton OHV Intake Model.
The following Briggs & Stratton engine is the only acceptable engine for the 2018
Baja SAE India competition:
No Exceptions
Baja Acceptable Engine
19L232-0054-G1
No other engine models will be accepted. No engine models from previous
competition years will be accepted.
1.3.2 Braking system
Rules are as follows:
B10.1 Foot Brake
The vehicle must have hydraulic braking system that acts on all wheels and is
operated by a single foot pedal.
The pedal must directly actuate the master cylinder through a rigid link (i.e., cables
are not allowed). The brake system must be capable of locking ALL FOUR wheels,
both in a static condition as well as from speed on pavement and on unpaved
surfaces. [1]
B10.2 Independent Brake Circuits
The braking system must be segregated into at least two (2) independent hydraulic
circuits such that in case of a leak or failure at any point in the system, effective
braking power shall be maintained on at least two wheels.
Each hydraulic circuit must have its own fluid reserve either through separate
reservoirs or by the use of a dammed, OEM-style reservoir.

5
B10.3 Brake(s) Location
The brake(s) on the driven axle must operate through the final drive. Inboard braking
through universal joints is permitted.
Braking on a jackshaft through an intermediate reduction stage is prohibited
B10.4 Cutting Brakes
Hand or feet operated “cutting brakes” are permitted provided the section (B10.1) on
“foot brakes” is also satisfied. A primary brake must be able to lock all four wheels
with a single foot. If using two separate pedals to lock 2 wheels apiece; the pedals
must be close enough to use one foot to lock all four wheels. No brake, including
cutting brakes, may operate without lighting the brake light.
B10.5 Brake Lines
All brake lines must be securely mounted and not fall below any portion of the
vehicle (frame, swing arm, A arms, etc.). Wheel ends should be connected with
flexible pipe only. Ensure they do not rub on any sharp edges.

6
CHAPTER 2
LITERATURE SURVEY
Gears are toothed wheels that provide for the transfer of rotary motion from one shaft
to another (or rotary-to-linear motion in the case of racks), as well as speed increase
or reduction from one shaft to another.
2.1 Terms used in gears
The following terms, which will be mostly used, are as follows:[3]
1. Pitch circle. It is an imaginary circle which by pure rolling action, would give the
same motion as the actual gear.
2. Pitch circle diameter. It is the diameter of the pitch circle. The size of the gear is
usually specified by the pitch circle diameter. It is also called as pitch diameter.
3. Pitch point. It is the common point of contact between two pitch circles.
4. Pitch surface. It is the surface of these rolling discs which the meshing gears have
replaced at the pitch circle.
5. Pressure angle or Angle of obliquity. It is the angle between the common normal to
two gear teeth at the point of contact and the common tangent at the pitch point. It is
denoted by ⱷ . The standard pressure angles are 14 ½̊ and 20 ̊ .
6. Addendum. It is the radial distance of a tooth from the pitch circle to the top of the
tooth.
7. Dedendum. It is the radial distance of a tooth from the pitch circle to the bottom of
the tooth.
8. Addendum circle. It is the circle drawn through the top of the teeth and is
concentric with the pitch circle.
9. Dedendum circle. It is the circle drawn through the bottom of the teeth and is
concentric with the pitch circle.
10. Circular pitch. It is the distance measured on the circumference of the pitch circle
from a point of one tooth to the corresponding point on the next tooth. It is usually
denoted by pc. Mathematically,
Circular pitch, pc = πD/T
D= Diameter of the pitch circle, and
T= Number of teeth on the wheel.

7
11. Diametral pitch. It is the ratio of number of teeth to the pitch circle diameter in
millimetres. It denoted by pd. Mathematically,
Diametral pitch, Pd = T/D
= π/Pc
Where T = Number of teeth
D = Pitch circle diameter
12. Module. It is the ratio of the pitch circle diameter in millimetres to the number of
teeth. It is usually denoted by m. Mathematically,
Module, m = D / T
Note: The recommended series of modules in Indian Standard are 1, 1.25, 1.5, 2, 2.5,
3, 4, 5, 6, 8, 10, 12, 16, 20, 25, 32, 40 and 50. The modules 1.125, 1.375, 1.75, 2.25,
2.75, 3.5, 4.5, 5.5, 7, 9, 11, 14, 18, 22, 28, 36 and 45 are of second choice.
13. Clearance. It is the radial distance from the top of the tooth to the bottom of the
tooth, in a meshing gear. A circle passing through the top of the meshing gear is
known as clearance circle.
14. Total depth. It is the radial distance between the addendum and the dedendum
circle of a gear. It is equal to the sum of the addendum and dedendum.
15. Working depth. It is radial distance from the addendum circle to the clearance
circle. It is equal to the sum of the addendum of the two meshing gears.
16. Tooth thickness. It is the width of the tooth measured along the pitch circle.
17. Tooth space. It is the width of space between the two adjacent teeth measured
along the pitch circle.
18. Backlash. It is the difference between the tooth space and the tooth thickness, as
measured on the pitch circle.
19. Face of the tooth. It is surface of the tooth above the pitch surface.
20. Top land. It is the surface of the top of the tooth.
21. Flank of the tooth. It is the surface of the tooth below the pitch surface.
22. Face width. It is the width of the gear tooth measured parallel to its axis.
23. Profile. It is the curve formed by the face and flank of the tooth.
24. Fillet radius. It is the radius that connects the root circle to the profile of the tooth.
25. Path of contact. It is the path traced by the point of contact of two teeth from the
beginning to the end of engagement.
26. Length of the path of contact. It is the length of the common normal cut-off by the
addendum circles of the wheel and pinion.

8
27. Arc of contact. It is the path traced by a point on the pitch circle from the
beginning to the end of engagement of a given pair of teeth. The arc of contact
consists of two parts, i.e.
(a) Arc of approach. It is the portion of the path of contact from the beginning of the
engagement to the pitch point.
(b) Arc of recess. It is the portion of the path of contact from the pitch point to the
end of the engagement of a pair of teeth.
The literature review is carried out on the above said areas to understand the failures
and rectifications of gear manufacturing.
2.2 INTRODUCTION TO DESIGN OF GEARS
The objectives of designing a gear consisting of the following fields:[11]
i. Gear failure
ii. Gear noise and vibration
iii. Geometrical modification of gears
iv. Establishment of gear test rig
2.2.1 GEAR FAILURE
A special attention needs to be taken for reducing gear failure that occurs
during normal working cycles. The mode of failure indicates that changes are needed
in lubricant choice and gear design. Gear tooth failure modes can be classified into
two major categories, first one is lubricant-related failure and another one is non-
lubricant-related failure. Failure modes related to lubricants are Hertzian fatigue
commonly called as pitting and micro pitting, adhesive and abrasive wear, scuffing,
etc.
Normally, wear occurs under low pitch line velocity and thin film thickness;
basically wear is the constant removal of metallic particles from the tooth flank .
Intensity of wear directly depends on lubrication, working condition, surface finish,
profile geometry and also material characteristics. Scoring and scuffing occurs due to
sudden weld between two contacting teeth surfaces. When this occurs at low
temperature, the phenomenon is called scoring and if it occurs at high speed, it is
called scuffing. Scoring failure initiates with small craters that progress through entire

9
active tooth flanks. Scuffing occurs due to overheating which results in a lubricant
film rupture and creates localized surface adhesion. Preceding to this phenomenon of
failure, there will be an increase in vibration and noise. Pitting is a fatigue failure
mostly found near spur gear pitch line. This type of failure usually originates from
surface cracks. When the crack is grown to some extent, they separate a piece of
material, thus a pit is formed. When several pits merge together, a large pit referred as
„spall‟ is formed. The main cause for micro pitting is operating conditions like load,
temperature, speed, lubricant film thickness, lubricant additives, and material
properties. To extend pitting life, contact stress must be kept low and lubricant film
thickness must be high. Micro- pitting occurs mainly due to lubricant failure, it
originates from micro cracking below the pitch line.
2.2.2 GEAR NOISE AND VIBRATION TESTING
Noise is one of the main recognized environmental problems which affect the
“quality of life”. Government is consequently increasing pressure for introducing
legislation to reduce noise from automobiles, machines etc. From the investigation
conducted by Zhao and Reinhart (1999), it is inferred that noise is generated by the
forcing functions used to drive the engine. Basic forcing functions of the engine are
impact of gears and bearings, cylinder pressure, piston slap, valve clearance. These
forces vibrate the structure of the engine and radiate noise. In transmission system,
gears are one of the main contributors for noise. Alternating torque results in
excitation of gears which is then transmitted to the engine mountings, panels, struts
etc. resulting in radiation of noise. Transmission Error (TE) is an important excitation
mechanism in gears which results in noise and vibration. The definition of Welbourn
(1979) states that, “Transmission error is the difference between the actual position of
the output gear and the position it would occupy if the gear drive were perfectly
conjugate”. This can also be expressed as a displacement at pitch line. Barthod et al
experimentally describes the rattle phenomenon in a gearbox. In general, transmission
errors are caused by deflections, geometrical modifications and geometrical errors.
Examples for deflections are gear teeth bending, contact deformation, deflection of
gear blank, shaft deflection etc. Examples of geometrical modifications are lead and
profile crowning, tip and root relief, etc. Examples of geometric errors are involute
form and alignment deviation, lead deviations, run-out, pitch error, etc. Mackaldener
(2001) classifies noise generation process in to three parts, namely excitation,

10
transmission, and radiation. Excitation is the origin of noise due to gear mesh in
which vibrations are created. Transmission is the process of vibration transmitted via
the gears, shaft, bearing and to the housing. Radiation of noise occurs when the
housing vibrates, which leads to pressure variations in the surrounding air.
2.2.3 GEOMETRICAL MODIFICATION OF GEARS
Fracture at the gear tooth root due to bending stress is one of the major causes
for gear failure. Parabolic beam model was used for stress concentration correction to
reduce this type of failure (ANSI/AGMA 2001-B88 1988). The shape of the tooth,
fillet geometry, highest point of full load contact and number of teeth are some of the
factors which affect bending strength. To increase payload, power density and
performance of a system the weight of the drive system can be reduced. Designers are
continuously working on ways to reduce drive system weight. (Lewicki et al 1999;
Crippa et al 1998) investigated web and rim thickness sizing of plastic gear for high
loading condition. Lewicki& Roberto (1997) investigated the rim thickness effect on
life of crack propagation. From the result, it is observed that while the rim thickness
decreased, the compressive cyclic stress in the tooth fillet region increased. This slows
down the crack growth and increased the crack propagation life. High quality gears
are often with tooth profile modification (tip relief) to reduce transmission error and
gear noise.
Figure 2.1 Normal and modified gear profile models

11
Faydor et al (2000) have proposed an asymmetric spur gear drive which has
the combination of an involute gear and double crowned pinion that stabilizes the
bearing contact and reduces the magnitude of transmission error. The result of FEA
also confirms the decrease in bending and contact stress. Huseyin (2009) has modified
the tooth width along the single mesh area as shown in Figure 2.1. By this
modification, constant hertz surface pressure was maintained along the profile. As a
result of homogeneous load distribution, negative effects caused by hertz pressure like
pitting, wear etc, were reduced. [18]
Figure 2.2 Selective reinforced gear with spline hub
Senthilvelan&Gnanamoorthy (2006) have developed a selective short carbon
fibre reinforced Nylon 66 spur gears (Figure 2.2). This method is developed to mould
the polymer spur gears with high strength fiber reinforcement in the extremely
stressed area. They reinforced with circular and spline hub and concluded that spline
hub exhibits superior fatigue life.[19]

12
Figure 2.3 Position of circular fillet
Shanmugasundaram&Muthusamy (2011) have introduced circular root filet
instead of standard trochoidal root filet in the gear (Figure 2.3). This novel method is
proposed to prevent the tooth failure in the spur gear. The study reveals that the
proposed design (circular root filet) exhibits higher bending strength than the standard
trochoidal root filet design.
2.2.4 ESTABLISHMENT OF GEAR TEST RIG
Power transmitted by gear boxes are fluctuating strongly in many of the
applications. For example, in automobiles, based on the condition of driving, torque
and speed varies. In machining operations, based on the material to be machined,
torque and speed varies. These evidences show that test rigs are needed for testing
such gear boxes (Roylance et al 1994;Madhavan 1991; Fessett 1975a; Krntz et al
2001).
Performance of gears depends on parameters like its design, material,
manufacturing and working environment. Developing a mathematical model to
predict the life of a gear will be difficult because of interaction between these
parameters (Damodar 1972; Kato 1994; Fessett 1975b).
Therefore a separate test rig is needed for predicting the life of the gears. Even
though many impressive progresses have been made in the field of simulation and
analysis in the past decades, experiments are necessary in many fields to conduct
investigation on the mechanical components (Athanassios 2009).

13
Gears are one of such components that needed to be tested experimentally.
Noise emission from a vehicle gearbox and gear fault was detected using a
microphone (Essam et al 2011). Sometimes, noise from the gearbox becomes a
dominating one and this creates a bad impression over the gear quality.
To overcome this, noise from the gear to be reduced to 10 - 15 dB when
compared with other noise sources like engine noise. Therefore, gears should be
tested for noise under controlled environment (Hellinger et al 1997;Campell et al
1997). Other than that, when the gear pair exceeds its load carrying capacity, different
modes of failure will occur like micro-pitting, pitting, tooth breakage, scuffing,
excessive wear, etc. other parameters like the gear’s dynamic behaviour and its
efficiency are also to be investigated experimentally.[19]
Figure 2.4 Simple construction of gear test rig
Therefore, a test rig which allows testing the gears under controlled
environment is needed. A simple construction of a test rig is shown in Figure 2.4,
which consists of a prime mover (motor), loading device (brake, dynamometer, pump,
etc.) and the test gear box which is to be installed between the prime mover and
loading device.

14
Figure 2.5 Test rig for investigating gear noise emission
2.3 Introduction to design of shafts:
A shaft is a rotating machine element which is used to transmit power from
one place to another. The power is delivered to the shaft by some tangential force and
the resultant torque (or twisting moment) set up within the shaft permits the power to
be transferred to various machines linked up to the shaft. In order to transfer the
power from one shaft to another, the various members such as pulleys, gears etc., are
mounted on it. These members along with the forces exerted upon them causes the
shaft to bending. In other words, we may say that a shaft is used for the transmission
of torque and bending moment. The various members are mounted on the shaft by
means of keys or splines.[3]
Notes: 1. The shafts are usually cylindrical, but may be square or cross-shaped in
section. They are solid in cross-section but sometimes hollow shafts are also used.
2. An axle, though similar in shape to the shaft, is a stationary machine element and is
used for the transmission of bending moment only. It simply acts as a support for
some rotating body such as hoisting drum, a car wheel or a rope sheave.

15
3. A spindle is a short shaft that imparts motion either to a cutting tool (e.g. drill press
spindles) or to a work piece (e.g. lathe spindles)
2.3.1 Material Used for Shafts:
The material used for shafts should have the following properties:
1. It should have high strength.
2. It should have good machinability.
3. It should have low notch sensitivity factor.
4. It should have good heat treatment properties.
5. It should have high wear resistant properties.
The material used for ordinary shafts is carbon steel of grades 40 C 8, 45 C 8, 50 C 4
and 50 C 12.
The mechanical properties of these grades of carbon steel are given in the following
table.
Table 2.1 Mechanical properties of steels used for shafts
When a shaft of high strength is required, then an alloy steel such as nickel, nickel-
chromium or chrome-vanadium steel is used.
2.3.2 Manufacturing of Shafts:
Shafts are generally manufactured by hot rolling and finished to size by cold
drawing or turning and grinding. The cold rolled shafts are stronger than hot rolled
shafts but with higher residual stresses. The residual stresses may cause distortion of
the shaft when it is machined, especially when slots or keyways are cut. Shafts of
larger diameter are usually forged and turned to size in a lathe.
2.3.3 Types of Shafts:
The following two types of shafts are important from the subject point of view:

16
1. Transmission shafts. These shafts transmit power between the source and the
machines absorbing power. The counter shafts, line shafts, overhead shafts and all
factory shafts are transmission shafts. Since these shafts carry machine parts such as
pulleys, gears etc., therefore they are subjected to bending in addition to twisting.
2. Machine shafts. These shafts form an integral part of the machine itself. The crank
shaft is an example of machine shaft.
2.3.4 Standard Sizes of Transmission Shafts:
The standard sizes of transmission shafts are: 25 mm to 60 mm with 5 mm
steps; 60 mm to 110 mm with 10 mm steps; 110 mm to 140 mm with 15 mm steps;
and 140 mm to 500 mm with 20 mm steps. The standard lengths of the shafts are 5 m,
6 m and 7 m.
2.3.5 Stresses in Shafts:
The following stresses are induced in the shafts:[3]
1. Shear stresses due to the transmission of torque (i.e. due to torsional load).
2. Bending stresses (tensile or compressive) due to the forces acting upon machine
elements like gears, pulleys etc. as well as due to the weight of the shaft itself.
3. Stresses due to combined torsional and bending loads.
2.3.6 Maximum Permissible Working Stresses for Transmission
Shafts:
According to American Society of Mechanical Engineers (ASME) code for
the design of transmission shafts, the maximum permissible working stresses in
tension or compression may be taken as
(a) 112 MPa for shafts without allowance for keyways.
(b) 84 MPa for shafts with allowance for keyways.
For shafts purchased under definite physical specifications, the permissible
tensile stress (σt) may be taken as 60 per cent of the elastic limit in tension (σel), but
not more than 36 per cent of the ultimate tensile strength (σu). In other words, the
permissible tensile stress, σt = 0.6 σel or 0.36 σu, whichever is less.[3]
The maximum permissible shear stress may be taken as
(a) 56 MPa for shafts without allowance for key ways.

17
(b) 42 MPa for shafts with allowance for keyways.
For shafts purchased under definite physical specifications, the permissible shear
stress (τ) may be taken as 30 per cent of the elastic limit in tension (σel) but not more
than 18 per cent of the ultimate tensile strength (σu). In other words, the permissible
shear stress, τ = 0.3 σel or 0.18 σu, whichever is less.
2.3.7 Design of Shafts:
The shafts may be designed on the basis of
1. Strength, and 2. Rigidity and stiffness.
In designing shafts on the basis of strength, the following cases may be considered:
(a) Shafts subjected to twisting moment or torque only,
(b) Shafts subjected to bending moment only,
(c) Shafts subjected to combined twisting and bending moments, and
(d) Shafts subjected to axial loads in addition to combined torsional and bending
loads.
2.4 INTRODUCTION TO BRAKING SYSTEM:
Braking is a mechanism used for slowing, stopping & controlling the vehicle.
Braking operation is based on kinetic energy of vehicle is to converting into heat,
which dissipated into atmosphere. While driving the vehicle; torque of the engine
produces the tractive effort due to periphery of driving vehicle. When the brakes are
applied it produces negative tractive effort on wheel. While, this help to slow down a
vehicle.
2.4.1 FUNCTIONS OF BRAKING SYSTEM:
➢ To stop the vehicle safely in shortest possible distance in case of emergency.
➢ To control the vehicle when it is descending along the hills.
➢ To keep the vehicle in desired position after bringing in at rest.

18
Figure 2.6 Simple Braking System
2.4.2 WORKING OF BRAKES:
A common misconception about brakes is that brakes squeeze against a drum or
disc and pressure of the squeezing action slows the vehicle down. Actually brakes use
friction of brake shoes and drums to convert kinetic energy developed by the vehicle
into heat energy. When we apply brakes, the pads that press against the brake rotor
convert kinetic energy into thermal energy via friction.
2.4.3 TYPES OF BRAKING SYSTEMS:
1. Hydraulic braking system
2. Anti braking system
3. Pneumatic braking system
4. Disc brake system
5. Mechanical brake system

19
2.4.3.1 HYDRAULIC BRAKING SYSTEM
Hydraulic is the use of a liquid under pressure to transfer Force or motion, or
to increase an applied force. The pressure on a liquid is called hydraulic pressure. And
the brakes which are operated by means of hydraulic pressure are called Hydraulic
brakes. These brakes are based on the principle of Pascal’s law.
Figure 2.7 Hydraulic Braking System
PASCAL’S LAW:
The pressure exerted anywhere in a mass of confined liquid is transmitted
undiminished in all direction throughout the liquid and applied in hydraulic lifts and
hydraulic brakes.
Figure 2.8 Pascal’s Law

20
CONSTRUCTION OF HYDRAULIC BRAKING SYSTEM:
Figure 2.9 Construction of Hydraulic Braking System
Hydraulic braking system is mainly confined with “brake fluid” this fluid consist
of Alcohol, castor oil & glycerin. Hydraulic braking system has following
components.
• Master cylinder
• Brake pedal
• Wheel cylinder
• Brake drum
• Retracting spring
• Brake shoe etc.
WORKING SYSTEM:
The brake pedal is connected to the master cylinder by means of piston for
application of brake driver presses the brake pedal, which moves the master cylinder.
In master cylinder pressure is instantly transferred to all four wheels. The brakes shoe
moves against the brake drum to apply brakes. When driver releases the brake pedal,
the master cylinder piston returns to its original position due to return springs,
dropping fluid pressure.[6]

21
Figure 2.10 Working of Hydraulic Braking System
2.4.3.2 DISC BRAKE:
A Disc Brake is a type of brake that uses calipers to squeeze pairs of pads
against a disc or “rotor” to create friction. This action retards the rotation of a shaft,
such as a vehicle axle, either to reduce its rotational speed or to hold it stationary.
Figure 2.11 Disc Brake

22
2.4.4 INBOARD BRAKING
An inboard braking system is an automobile technology wherein the
disc brakes are mounted on the chassis of the vehicle, rather than directly on the
wheel hubs.
The second thing we looked at in the braking system was if we were going to
incorporate an inboard braking system (with the brake disc on the rear drive axle)
in the rear or an outboard (with the brake discs at the wheels). Because the 2014
car had a solid axle output shaft, they used a single inboard rear brake where the
brake rotor was directly connected to the output shaft. The inboard system would
require more braking force to be generated from a single rotor but would utilize
less part. We decided to use an inboard rear braking system. Ease of maintenance
was a main focus for the drive train team this year, so designing a rotor coupler
and caliper mount so that all bolts were easily accessible was a top priority and a
significant improvement over last year’s design. It was decided to have the caliper
bolt to a bracket which was in turn bolted directly to the case. This design should
give easier access to the caliper and to the rotor in the event that they ever need to
be quickly changed out.
Figure 2.12 Integrated brakes Figure 2.13 Inboard brakes
2.4.4.1 ROTORS
The brake rotors are a part of the system that can be enhanced to gain
performance as well as limit weight. There are different types of rotors that

23
include slotted, drilled, and plain rotors. Plain rotors are not used in most racing
applications so we ruled that one out which left us with two options; slotted and
drilled. We originally decided slotted due to the capability of flinging mud out of
the rotor while the wheel is rotating. Upon further research, dilled rotors were
decided due to the cost effectiveness and the braking performance was little to no
difference. The front two wheels will require a rotor each and the rear only
requires one rotor due to the inboard rear braking system. We decided to
manufacture our own because we needed a custom fit to mount onto the inboard
hub design. [10]
The rear only requires one rotor due to the integrated braking system. It was
manufactured in house due to constraints of the drive terrain housing; it is made
out of steel as well. FEA was conducted on the rear rotor to insure that the
mounting holes would not shear off due to stress it would see during braking at
full force.
Figure 2.14 Brake Rotor
Figure 2.15 Brake Rotor FEA

24
2.4.4.2 Calipers
The caliper was the next item we looked at. The caliper transfers braking
fluid to force the brake pads to compress against the rotor, thus slowing the vehicle
down. We decided to utilize Pulsar 220F Fixed calipers because we already had
one spare Pulsar 220F Fixed caliper; this would help reduce the cost of the system
because only two more are required. The pads that fit onto this caliper are of a
sintered steel material. This material works best when it is activated against a steel
rotor and not a stainless-steel rotor (our rotor material is steel). These calipers are
also made from aluminum which will continue the trend to cut weight from the
system.
Figure 2.16 Pulsar 220F Caliper
2.4.4.3 Master Cylinder
The master cylinders play a large role in the design phase. The master
cylinder will have to be able to transfer the right amount of pressure to the brake
caliper pistons. The master cylinders that were chosen are a Tandem master
cylinder with a 5/8 inch bore. Through our calculations we found that this will
provide more than enough pressure when given a 150 N driver input force when
paired with a 4:1 pedal ratio (lever arm ratio for master cylinder input force). It was
determined through research on the internet that the average driver applies 150 N
to the brake pedal. The pedal ratio was decided based on the pedal configuration
and verified by the average driver foot size. The design includes an inboard rear

25
braking system to assist in achieving our cost and weight goal.
Figure 2.17 Working Of Tandem Master Cylinder
Figure 2.18 Master Cylinder with reservoir

26
CHAPTER 3
DESIGN CONSIDERATIONS OF POWERTRAIN
3.1 INTRODUCTION TO POWERTRAIN
The powertrain acts as a power source for the vehicle and its main purpose is
to provide the driving torque at the wheels. This applied torque at the driven wheel
causes vehicle to movie. The power source should be chosen such that it should be
able to provide high torque at low rpm and peak power at high rpm. For such dynamic
requirement, a stock four cycles, air cooled engine serve the purpose. According to
the rules of SAE, we considered Briggs & Stratton (B&S) 10 HP OHV intek engine as
a prime mover.
The following Briggs and Stratton engine is the only acceptable engine for
2018 Baja SAE India competition: 19L232-0054-G1
This engine develops maximum torque of 18.66 N-m at 2600 rpm and peak power of
9.14 HP at 3800 rpm.[1]
Figure 3.1 Torque v/s Engine rpm

27
Figure 3.2 Power v/s Engine rpm
3.2 Transmission
The main objective of transmission is to provide the desired torque and speed.
The desired torque means, the torque required to pull the driving wheel against the
road condition, which includes rolling resistance (RR), aerodynamic resistance/air
resistance (AR) and grade resistance (GR). To choose the transmission capability for
producing enough torque to move the vehicle, it is necessary to determine the total
tractive effort (TTE) required for the vehicle. Below we have generated a gear ratio
from our given constraints with reasonable assumptions,
Assumptions and variables:
Wheel Diameter : 23 inches
Total Weight : 232 Kgs
Slope : 30º
Reduction ratio : 10.4
Efficiency CVT : 88% (from manufacturer of CVtech Cvt)

28
Coefficient of friction(μ) : 1.0 (concrete)
: 0.65 (dry road)
Coefficient of rolling resistance (fr) : 0.014 (concrete)
: 0.05 (dry road)
In this segment all the further calculations will be done according t the above data.
3.2.1 GOAL:
A) Torque:
In the inclined of steep situation of the track, the required torque
should be produced as we assumed the incline to be 30 degree(max) with inspection
of previous year’s course. In order to complete the incline, the force on the wheels
will need to be greater than the component of force of gravity along the incline.
Figure 3.3 Free body diagram of the vehicle
G1 = G sin 30 = 116 N
Force per wheel = G1/2 = 58 N
Torque per wheel = (G1/2) x (D/2) = 174 N-m
Total torque = 348 N-m
So, we can assume that the minimum torque that needs to be transferred to the wheels
is 348 N-m.

29
B) Speed :
We considered 4.2 sec to finish 100 foot course as per the top team result in
previous year’s competition
Distance = maximum velocity x time/2
Max velocity = distance x 2 / time = 100 foot x 2 / 4.2sec =52.25 kmph
Therefore 60 kmph is the goal for the max speed to obtain.
3.2.2 Analysis of CVT system:
The analysis of Continuous Variable Transmission provides the gear ratios that
would be required to obtain the goals introduced above. Though the analysis need
some assumptions such as wheel diameter and total weight should be chosen
appropriately which are mentioned above.
3.2.2.1 Continuous Variable Transmission Set-Up
The CVT has initial gear ratios of 0.45: 1 (high) and 3.1: 1 (low). This
however was not ideal for the goals that have been established. Thus the group had to
consider a secondary reduction or in this case two. As stated above in the assumptions
and variables, the total reduction ratio should be 10.4: 1. For the volume provided to
us by the chassis department, which is approximately 6.3 cubic feet, our team put
together this sample layout of the reduction system as seen in Figure 3.4. In figure 3.5
we depict how the engine, CVT and reduction system sat with in the frame. As you
can see, because of the odd shape of the rear, to optimize the space, the engine should
be mounted approximately 17 inches above the bottom of the frame. This can be
visualized in Figure 3.6. This will allow for ample space to implement the reduction
system and eventually our braking system. [7]
The reduction contains 4 sprockets with different teeth
N1 = 18 N3 = 18
N2 = 48 N4 = 70
Sprocket 1&2 is the first stage with 2.67:1 ratio, sprocket 3&4 is the second stage
with 3.89:1 ratio. The total reduction ratio is 10.4:1.

30
Figure 3.4 Basic concepts of CVT drive train systems

31
Figure 3.5 3-D drawing of CVT Drivetrain system
Figure 3.6 Simple depiction of the layout of the rear of the frame and the
prospected optimal placement of the engine

32
3.2.2.2 Calculations
Figure 3.7 Motor Torque Curve
From the graph above we obtain the RPM and torque output from the engine.
Then we calculated the following with our assumptions:[8]
➢ CVT ratio = 3- (2.5*{rpm-800})/2800 for 800<rpm<3600
➢ Total ratio= Rcvt * Rr* Ncvt = Rcvt * 10.4 * 0.88
➢ Torque on the wheel = Torque output * Total ratio * Ncvt
➢ Speed = {(D*RPM*π)/(Total ratio*10.4*60)} x 0.68
= {( 23 in * RPM * π)/ (total ratio*10.4*60)} x 0.68
With these equations above we made the table below.

33
Engine rpm Torque
output(lb-ft)
CVT ratio Total ratio Torque on
wheel(lb-ft)
Speed(mph)
1800 13.20 2.107 22.251 293.719 5.52
2000 13.70 1.929 20.366 279.010 6.70
2200 14.10 1.750 18.480 260.568 8.12
2400 14.30 1.571 16.594 237.298 9.87
2600 14.45 1.393 14.709 212.539 12.06
2800 14.52 1.214 12.823 186.188 14.90
3000 14.50 1.036 10.937 158.589 18.72
3200 14.40 0.857 9.051 130.341 24.13
3400 14.20 0.679 7.166 101.753 32.38
3600 13.80 0.500 5.280 72.864 46.53
Table 3.1 this table displays our numerical data as it relates to our assumptions
and the equations.
The max torque is 293.7 lb-ft and the max speed is 46.53 mph which satisfy
the team’s intended goals. Thus our assumptions for the cvt ratio are realistic and
obtainable. Based on the 0.5 high ratio and the 3 low ratio, the team choose the CVT:
PULLEY SERIES 0600 AND DRIVEN PULLEY SERIES 5600 from CVTech- AAB
Inc. This CVT provides a range of 0.45 high and 3.1 low ratio that will be compatible
with the design. However, it changes the equation for CVT ratio slightly, thus:
CVT ratio = 3.1 – {(2.65*(rpm-800)}/2800 for 800<rpm<3600
With the new ratio we can adjust our data with the same calculations:
Engine rpm Torque
output(lb-ft)
CVT ratio Total ratio Torque on
wheel(lb-ft)
Speed(mph)
1800 13.20 2.154 22.742 300.191 5.40
2000 13.70 1.964 20.743 284.177 6.58
2200 14.10 1.775 18.744 264.290 8.01
2400 14.30 1.586 16.745 2339.456 9.78
2600 14.45 1.396 14.746 213.084 12.03

34
2800 14.52 1.207 12.747 185.093 14.99
3000 14.50 1.018 10.749 155.854 19.05
3200 14.40 0.829 8.750 125.996 24.96
3400 14.20 0.639 6.751 95.862 34.37
3600 13.80 0.450 4.752 65.578 51.70
Table 3.2 This table displays our numerical data as it relates to our assumptions
and the equations with a slight change in ratio.
The maximum torque applied on the sprockets are followed by the equations below
where: (T is the torque output from engine,T1 is the torque applied on the first
sprocket)[8]
T1= T * Rcvt * Ncvt = 13.20 lb-ft*2.154*0.88 = 25.02 lb-ft.
T2 = T1 * (N2/N1) = 25.02 * 4 = 100.08 lb-ft
T3 = T2 = 100.08 lb-ft
T4 = T3 * (N3/N2) = 100.08 *3 = 200.19 lb-ft.
3.3 Modelling of gears
For modelling of gears, we have used SOLIDWORKS software and to check
the meshing of gears. The gear box is of two stage single reduction type as seen in the
Figure 3.4. It is depicted in the Figure 3.8 and Figure 3.9 shown below:

35
Figure 3.8 First stage reduction
Figure 3.9 Second stage reduction

36
3.4 Modal Analysis:
Definition: We use Modal Analysis to determine the vibration characteristics (Natural
frequencies and mode shapes) of a structure of a machine component while it is being
designed. It also can be a starting point for another, more detailed, Dynamic Analysis,
such as a transient dynamic, a harmonic response analysis, or a spectrum analysis.
Uses for Modal Analysis:
The Natural frequencies and mode shapes are important parameters in the
design of a structure for Dynamic loading conditions. They are also required if you
want to do a spectrum analysis or a mode superposition harmonic or transient
analysis.
We can do modal analysis on a pre stressed structure, such as a spinning
turbine blade. Another useful feature is modal cyclic symmetry, which allows you to
review the mode shapes of a cyclically symmetry structure by modelling just a sector
of it.
Modal Analysis in the ANSYS family of products is a linear analysis. Any
nonlinearity, such as plasticity and contact (gap) elements, are ignored even if they
are defined. You can choose from several mode extraction methods: subspace, Block
Lanczos, Power Dynamics, reduced, unsymmetrical, and damped. The damped
method allows you to include damping in the structure. Details about mode extraction
methods are covered later in this section.[14]
3.4.1 Structural Static Analysis:
A static analysis calculates the effects of static loading conditions on a
structure, while ignoring inertia and damping effects, such as those caused by time-
varying loads. A static analysis can, however, include Transient inertia loads (such as
gravity and rotational velocity), and time-varying loads that can be approximated as
static equivalent loads (such as the static equivalent wind and seismic loads
commonly defined in many building codes).
3.4.2 Loads in a Static Analysis:
Static analysis is used to determine the displacements, stresses, strains and
forces in structures or components caused by loads that do not induce significant
inertia and damping effects. Transient loading and response conditions are assumed;

37
that is, the loads and the structure’s response are assumed to vary slowly with respect
to time. The kinds of loading that can be applied in a static analysis include:
• Externally applied forces and pressures
• Steady- inertial forces (such as gravity or rational velocity)
• Imposed (non-zero) displacements
• Temperatures (for thermal strain)
• Fluencies (for nuclear swelling)
A static analysis calculates the effects of static loading conditions on a
structure, while ignoring inertia and damping effects, such as those caused by time-
varying loads. A static analysis can, however, include static inertia loads (such as
gravity and rotational velocity), and time varying loads that can be approximated as
static equivalent loads (such as static equivalent wind and seismic loads commonly
defined in many building codes.
3.4.3 Linear vs. Non linear Static Analysis:
A static analysis can be either linear or non linear. All types of nonlinearities
are allowed-large deformations, plasticity, creep, stress stiffening, contact (gap)
elements, hyper elastic elements etc.
3.5 OVERVIEW OF STEPS IN A STATIC ANALYSIS:
The process for static analysis consists of three main steps: -
1. Build the Model: - To build the model, specify the job name, analysis title and
then define the element types, element real constants, material properties, and
the model geometry. The structural elements can be linear or nonlinear.
Material properties can be linear or nonlinear, isotropic or orthographic, and
constant or temperature-dependent. The Young’s modules should also be
defined.
2. Apply the loads obtain the solution:- In this step, define the analysis type and
options, apply loads, specify load step options, and begin the finite element
solution. The loads that can be applied are :
1. Displacements- degree of freedom, constraints usually specified at
modal boundaries to define rigid support points.
2. Forces- concentrated loads usually specified on the model exterior,
moments.

38
3. Pressures- surface loads usually applied on the model exterior,
temperatures.
4. Fluencies- applied to stuffy the effects of swelling or creep.
5. Review the results: - Results from a static analysis include the nodal
displacements, nodal and element stresses, nodal and element strains,
element forces, nodal reaction forces etc.
3.6 STATIC ANALYSIS
Static analysis is concerned with determination of response of a gear to steady
loads whose response remains unchanged with time. The response of the gear is
expressed in terms of stress, strain, displacement. The tool used in the static analysis
is Static structural. The finite element analysis procedure of the spur gear was given
below:
• A three-dimensional model of the spur gear was created using the SOLIDWORKS
software.
• The material properties were defined for gears.
• The model was meshed using finite element software.
• Boundary conditions for ANSYS Workbench as mentioned below:
• Fixed displacement constraint was applied on gear
• Moment was applied on gear
• In order to arrest the displacement on x, y, z directions and rotations
on x, y directions remote displacement constraint is applied on pinion
surface.
Analysis Procedure:
1. After designing the basic model of gears considering all conditions,
save an IGES in a separate folder location.
2. Open Ansys design modeler and import this IGES file.

39
3. In import settings, urn off the surface and volume body import options.
Make sure that line body settings are only on.
4. Now click generate button
5. Mesh the models as given below :
1. Pinions with size 2 mm
2. Gears with size 4 mm.
The meshed model looks like this.
Figure 3.10 showing the model the pinion gear.
Figure 3.11 Meshing of pinion gear.

40
Figure 3.12 Loading scenarios on pinion.
Equivalent stress:
Figure 3.13 Equivalent stress on pinion gear 1

41
Total deformation:
Figure 3.14 Total deformation
Figure 3.15 Intermediate gear

42
Figure 3.16 Meshing of intermediate and output gears.
Figure 3.17 Loading scenarios

43
Equivalent stress:
Figure 3.18 Output gear
TOTAL DEFORMATION:
Figure 3.19 Intermediate gear

44
Figure 3.20 Output gear
Figure 3.21 Input shaft and gear

45
Figure 3.22 Analysis on intermediate gears
Figure 3.23 Analysis on output gear
3.6.1 FEA Results of Gears:
Displacement results of FEA of gears are shown above; forces are calculated
according to the engine torque transmitted to gearbox, due to proper material selection

46
all gears are safe with FOS value of more than 2 and maximum displacement is less
than 0.15 mm at maximum load.
3.7 Technical data:
The overall gear features:
Gear No. of Tooth Pitch Dia. (mm) Module
Input 18 36 2
Intermediate 1 48 96 2
Intermediate 2 18 36 2
Output 70 140 2
Table 3.3 Overall gear features
3.7.1 Input shaft dimensions
Figure 3.24 Input shaft

47
3.7.2 Output spline data
Figure 3.25 Output shaft (Rzeppa Joint)
S.No Particulars Features Size
1 Gear box output spline 18 teeth 20 mm ID
2 Rzeppa Joint 18 teeth 20 mm OD
Table 3.4 Output shaft data
3.8 Gearbox casing
According to the design of gears and shaft the outline is made accordingly. CAD
drafts are shown below:
The gear housing’s machining was completed on a CNC machine, using a number of
tools, jigs and fixtures. The tolerances are critical for the bearings to be press fitted
and to ensure that the shafts had the correct spacing. Therefore, these surfaces are
bored. Ball bearings are held in place on the shafts by grooves and incorporated snap
rings. Similarly, the gears are held in place laterally on the shafts by grooves and
circlips, but rely on a standard keyway to prevent rotation. The casing of the gearbox
is machined out of 6061-T6 Aluminium and utilized mounts on the base to secure the
gearbox to the sub frame assembly.

48
Figure 3.26 Side view of the gear box
Figure 3.27 Rear view of the gear box

49
3.9 Rendered Images
Figure 3.28 Rendered image - I
Figure 3.29 Rendered image - II

50
CHAPTER 4
DESIGN CONSIDERATIONS FOR BRAKING SYSTEM
4.1 BRAKE TORQUE
Brake torque in in-lbs. (for each wheel) is the effective rotor radius in inches
times clamping force times the coefficient of friction of the pad against the rotor.
Brake torque is the force that actually decelerates the wheel and tire. There are two
components how hard the pads clamp the rotor (clamping force) and how far that
clamping takes place from the centre of the wheel hub. The larger the effective rotor
radius, the further the clamping takes place from the wheel centre and the torque
generated by this longer “lever effect”. This is very similar to the manner in which a
longer handle on a ratchet generates more torque than a short handle ( for the same
input). To increase brake torque it is necessary to increase the hydraulic pressure, the
caliper piston area, the coefficient of friction between pad and rotor, or the effective
rotor diameter.
4.2 CLAMPING FORCE
The clamping force that a caliper exerts, measured in pounds, is the hydraulic
pressure(in psi) multiplied by the total piston area of the caliper (in a fixed caliper) or
two times the total piston area ( in a floating caliper), in square inches. To increase the
clamping force it is necessary to either increase the hydraulic pressure or the caliper
piston area. Increases the coefficient of friction will not increase clamping force.
4.3 COEFFICIENT OF FRICTION
The coefficient of friction between pad and rotor is an indication of the
amount of friction between the two surfaces. The higher the coefficient, the greater
the friction. Typical passenger car pad coefficients are in the neighbourhood of 0.3 to
0.4. Racing pads are in the 0.5 to 0.6 range. “Hard” pads have a lower coefficient but
wear less;”soft” pads have a higher coefficient but can wear quickly. With most pads,
the coefficient is temperature sensitive-which is why sometimes racers need to
“Warm “P” the breaks before they work well and also why most brakes will “fade”

51
when they overheat- the coefficient of friction is reduced as the temperature rises. For
more info in coefficient of friction, see section on pads.
4.4 THERMAL CAPACITY
The brake rotors must be capable of absorbing the heat generated by the
brakes as they convert the moving car’s kinetic energy into heat. The amount of
kinetic energy a car has (and, therefore, the amount of heat the rotors must be able to
absorb) depends on the weight of the car and the square of the speed of the car. The
rotor’s ability to absorb this heat depends on its mass(weight), and on how well it
cools. Exposed as they are to cooling airflow, this is one area where discs are superior
to drums.
The following equations and examples will help to clarify the concepts:
1) Brake torque
2) Brake torque required is calculated as:
TBr = Friction force on tire x Rolling radius of tire
Where TBr = Brake torque required (in. lbs.)
Friction force on tire = Vertical force on tire x grip
Grip = coefficient of friction between the tire and road
The tyre’s grip is difficult to measure, and can vary from 0.1 on wet ice to
about 1.4 for a racing slick on a hot, dry track. If you don’t have a value for your
tyres, use 1.0 as an average value.[16]
The calculation of friction force on tyre is different for front and rear tires,
takes into account weight transfer, and requires the calculation of vertical force on
both tyres first.
Front: FF = µ Ff/2
Where: FF= Friction force on front tyre
µ= grip (use 1.0)
Ff = vertical force on the both front tyres
And Ff = We [1-(Xcg/1) + (uYcg/1)

52
Ff= vertical force on the both front tyres
We = weight of car (in lbs.)
Xcg = distance from front axle to car’s centre of gravity (in)
I = wheelbase (in)
P = grip (1186 1.0)
Ycg = height above ground of car’s centre of gravity
Rear: FR = µ Ff/2
Where FR = friction force on rear tyre
µ = grip (use 1.0)
Fr = vertical force on the both rear tyres
And Fr = We – Ff
Where Fr = vertical force on the both rear tyres
We = weight of car (lbs.)
Ff = vertical force on the both front tyres
By examining the above equations carefully, we can learn some valuable
things. For example, note that the equation for brake torque required:
• Doesn’t involve vehicle speed in any way
• Will vary for different tyres & roads( or trail) conditions
• Does involve tyre/wheel radius
Clamping force is calculated as:
CF = PM * AT
Where
CF = clamping force (lbs.)
PM = maximum hydraulic pressure (psi)

53
AT = total effective area of caliper piston (sq. in.) ----
for fixed caliper this is the actual
Area of the piston, for floating caliper this is equal to 2 x the actual area of the
piston
Brake torque developed is calculated as;
TBd = CF (µL) Re
Where
TBd = Brake torque developed (in-lbs.)
CF = clamping force (lbs.)
µL = coefficient of friction between brake pads and
rotors (use 0.3, manufacturer’s specs, or estimate
derived from the pad’s DOT edge code (see section on
pads))
Re = effective rotor radius (in.) measured from the
centre of the rotor to the canter of the brake pad.
The maximum hydraulic pressure developed in your braking system can either
be measured.
With an inline pressure gauge, or can be calculated as:
PM = Fp / Ap
Where
PM = maximum pressure (psi)
Fp = force on master cylinder piston (lbs.)
= pedal effort x pedal ratio (i.e. how hard the driver
pushes the pedal multiplied by the pedal ratio.)
AP = area of the MC piston (sq. in.)

54
= 0.785 x Dp
2
(Dp= the MC piston diameter, in inches)
This is a very important equation. Note how the maximum pressure
determines the brake torque, and the maximum pressure is the force applied
DIVIDED by the area of the MC piston. This means, all other things being equal,
the bigger the MC piston, the LESS pressure developed, and the LESS brake
torque generated. The trade- off is, the smaller the MC, the less fluid is displaced
per inch of travel, and therefore the greater the pedal travel require more on this
later.[16]
Thermal capacity:
As previously discussed, the brake rotors must be capable of absorbing the
heat generated by the brakes as they convert the moving car’s kinetic energy into
heat.
The formula for the kinetic energy (K) in the moving car is;
K = (W*S2
)/29.9
Where
K = kinetic energy (ft.lbs.)
W = weight of car (lbs)
S = speed of car (mph)
(Note: the root equation for kinetic energy is actually K = 1/2mv2
, the above
version has conversion factor included so we can use weight instead of mass and
so the result is given in force units instead of energy units)
Nothing really surprising here we know from instinct and experience that
how much brake you need depends on how heavy the car is and how fast it’s
going. Note though, that speed is squared in the equation – meaning that as the
speed increases, the kinetic energy developed goes up by the square of the speed
increase for example, if speed doubles, kinetic energy increases by a factor of 4. If
speed triples kinetic energy goes up by a factor of (32
) of 9!! This is a bitch for

55
race –car drivers but not so much trail rigs. However if you’re beginning to “race”
your rock crawling buggy be aware that your brake required are going to increase
exponentially.
OK, so, moving rig has kinetic energy, kinetic energy must be turned into heat,
rotors must absorb said heat.
Well, the equations that follow are used to calculate the temperature increase
in the rotor for a given kinetic energy. Remembering that kinetic energy depends
on weight and speed they also explain in incontrovertible terms exactly why
“pinion brakes and why parolee’s rotors keep trying to melt remember- to add
insult to injury “pinion brakes” are often used on Rockwell – axled rigs- which are
big and heavy, and weight is a multiplying factor in the equation for kinetic
energy.
[As an aside the previous discussion on brakes torque also explains why
some feel that pinion brakes work well “ as slow speed”. It’s because the rotor is
placed before the axle.
Differential, meaning the pinion brake’s brake torque is calculated as above
& then multiplied by a factor equal to the axel ratio. This means, even with small
caliper and rotors, they can develop tremendous brake torque. But remember what
we said about the requirements of a braking system. It must also have sufficient
thermal capacity – and they simply do not. In fact, they’re dangerously inadequate
in this regard!]
The formula for temperature rise is:
TR = K=/77.8*Wb
Where
TR = temperature rise (0
F)
Kc = kinetic energy change (from start of braking to end
of braking) (Ft. lbs.)
Wb = weight of all rotors (lbs.)

56
And Kc = kinetic energy change (ft. lbs.)
=Kb- Ka
Where
Kb = kinetic energy before stop
Ka = kinetic energy after stop
Let’s do an example, imagine I’m stopping my 5000lb buggy from 40 mph to
a dead stop. First, let’s calculated the change in kinetic energy for a 5000 lbs.
buggy form 40 mph to 0mph – this remains the same regardless of the brake
system.
Kb = We*S2
/29.9
=5000*(40)2
/29.9
=267559ft-lbs.
Ka = 0
Kc = Kb – Ka
=267 599 – 0
=267 599
Now, with % ton truck disc brakes at each wheel, each rotor weighs above 21
Lbs. for a total rotor weight of 84 lbs,
In this configuration, the temp rise of the rotor will be:
TR = Kc/77.8*WB
= 267 559/ (77.8*84)
= 267 559/ 6535
= 410
F
If it were a hot day, after the stop my rotors could be 14101;

57
Now, imagine I have pinion brakes with 2 small rotors, weight, gay 12 lbs each
Now,
TR = Kc /77.8*Wb
=267 559/(77.8*24)
= 267 559 / 1867
= 1430
F
On the same hot day, the rotors are now at 2420
F.
Make a few more stops, with insufficient time for the rotors to cool fully between
them, and
It’s easy to see how you can seriously overheat the small pinion brake rotors.

58
CHAPTER 5
FABRICATION OF GEAR BOX
Spur gears are the most easily visualized common gears that transmit motion
between two parallel shafts. Because of their shape, they are classified as a type of
cylindrical gears. Since the tooth surfaces of the gears are parallel to the axes of the
mounted shafts, there is no thrust force generated in the axial direction. Also, because
of the ease of production, these gears can be made to a high degree of precision. On
the other hand, spur gears have a disadvantage in that they easily make noise.
Generally speaking, when two spur gears are in mesh, the gear with more teeth is
called the “gear” and the one with the smaller number of teeth is called the “pinion”.
The unit to indicate the sizes of spur gears is commonly stated, as specified by
ISO, to be “module”. In recent years, it is usual to set the pressure angle to 20
degrees. In commercial machinery, it is most common to use a portion of an involute
curve as the tooth profile.
Even though not limited to spur gears, profile shifted gears are used when it is
necessary to adjust the centre distance slightly or to strengthen the gear teeth. They
are produced by adjusting the distance between the gear cutting tool called the
hobbing tool and the gear in the production stage. When the shift is positive, the
bending strength of the gear increases, while a negative shift slightly reduces the
centre distance. The backlash is the play between the teeth when two gears are
meshed and is needed for the smooth rotation of gears. When the backlash is too
large, it leads to increased vibration and noise while the backlash that is too small
leads to tooth failure due to the lack of lubrication.
5.1 CLASSIFICATION OF MANUFACTURING PROCESSES OF
GEARS:
1. Milling process
• Disc type cutter
• End mill cutter

59
2. Gear planning process
• The Sunderland process
• The Maag process
3. Gear shapers
• Rack – type cutter generating process
• Pinion type cutter generating process
4. Gear hobbing
• Axial hobbing
• Radial hobbing
• Tangential hobbing
5. Bevel gear generating
• Straight Bevel – gear generator
• Spiral bevel –gear Generator
5.2 METHODS OF FORMING GEARS
5.2.1 Roll forming
In roll forming, the gears blank is mounted on a shaft & is pressed against
hardened steel of rolling dies. The rolls are fed inward gradually during several
revolutions which produce the gear teeth. The forming rolls are very accurately made
& roll formed gear teeth usually home both by not and cold. In not roll forming, the
not rolled gear is usually cold –rolled which compiles the gear with a smooth mirror
finish. In cold roll forming, higher pressures are needed as compared to not rolling
many of the gears produced by this process need no further finishing. It becomes
stronger against tension & fatigue. Spur & helical gears are made by this process.
5.2.2 Stamping
Large quantities of gears are made by the method known as stamping
‘blanking’ or ‘fine blanking’. The gears are made in a punch press from sheet up to
12.7mm think such gears find application in: toys, clocks 4 timers, watches, water &
Electric maters & some business Equipment. After stamping, the gears are shaved;
they give best finish & accuracy. The materials which can be stamped are: low,

60
medium & high carbon steels stainless steel. This method is suitable for large volume
production.
5.2.3 Powder metallurgy
High quality gears can be made by powder metallurgy method. The metal
powder is pressed in dies to convert into tooth shape, after which the product is
sintered. After sintering, the gear may be coined to in crease density & surface finish.
This method is usually used for small gears. Gears made by powder metallurgy method
find application in toys, instruments, small motor drivers etc.
5.2.4 Extrusion
Small sized gear can also be made by extrusion process. There is saving in
material & machining time. This method can produce any shape of tooth & is suitable
for high volume production gears produced by extrusion find application in watches,
clocks, type writers etc.
5.3 GEAR GENERATING PROCESS
5.3.1 Gear Hobbing
Hobbing is the process of generating gear teeth by means of a rotating cutter
called a hob. It is a continues indexing process in which both the cutting tool & work
piece rotate in a constant relationship while the hob is being fed into work. For in route
gears, the hob has essentially straight sides at a given pressure angle. The hob and the
gear blank are connected by means of proper change gears. The ratio of hob & blank
speed is such that during one revolution of the hob, the blank turns through as many
teeth. The teeth of hob cut into the work piece in Successive order & each in a slightly
different position. Each hob tooth cuts its own profile depending on the shape of cutter,
but the accumulation on the shape of cutter, but the accumulation of these straight cuts
produces a curved form of the gear teeth, thus the name generating process. One
rotation of the work completes the cutting up to certain Depth.
5.3.2 TYPES OF HOBBING
5.3.2.1 Arial hobbing
This type of feeding method is mainly used for cutting spur or helical gears. In
this type, firstly the gear blank is brought towards the hob to get the desired tooth

61
depth. The table side is them clamped after that, the hob moves along the face of the
blank to complete the job. Axial hobbing which is used to cut spur & helical gears can
be obtained by ‘climb noting’ or ‘conventional hobbing.
5.3.2.2 Radial hobbing
This method of hobbing is mainly used for cutting worm wheels. In this
method the hob & gear blank are set with their ones normal to Each other. The gear
blank continues to rotate at a set speed about its vertical axes and the rotating hob is
given a feed in a radial direction. As soon as the required depth of tooth is cut, feed
motion is stopped.
5.3.2.3 Tangential hobbing
This is another common method used for cutting worm wheel. In this method,
the worm wheel blank is rotated in a vertical plane about a horizontal axes. The hob is
also held its axis or the blank. Before starting the cut, the hob is set at full depth of die
tooth and then it is rotated. The rotating hob is then fed forward axially. The front
portion of the hob is tapered up to a certain length & gives the fed in tangential to the
blank face & hence the name ‘Tangential feeding’.
5.4 GEAR SHAPING PROCESS
In gear shapers, the cutters reciprocate rapidly. The teeth are cut by the
reciprocating motion of the cutter. The cutter can either be ‘rack – type cutter’ or a
rotary pinion type cutter’.
5.4.1 Rack – type cutter generating process
The rack cutter generating process is also called gear shaping process. In this
method, the generating cutter has the form of a basic rack for a gear to be generated
The cutting action is similar to a shaping machine. The cutter reciprocates rapidly &
removes metal only during the cutting stroke. The blank is rotated slowly but
uniformly about its axis and between each cutting stroke of the cutter, the cutter
advances along its length at a speed Equal to the rolling speed of the matching pitch
lines. When the cutter & the blank have rolled a distance Equal to one pitch of the
blank, the motion of the blank is arrested, the cutter is with drawn from the blank to
give relief to the cutting Edges & the cutter is returned to its starting position. The

62
blank is next indexed & the next cut is started following the same procedure.
5.4.2 Pinion type cutter generating process
The pinion cutter generating process is fundamentally the same as the rack
cutter generating process, and instead of using a rack cutter, it uses a pinion to generate
the tooth profile. The cutting cycle is commenced after the cutter is fed radically into
the gear blank Equal to the depth of tooth required. The cutter is then given
reciprocating cutting motion parallel to its axis similar to the rack cutter and the cutter
& the blank are made to rotate slowly about their axis at speeds which are Equal at the
matching pitch surfaces. This rolling movement blow the teeth on the blank are cut.
The pinion cutter in a gear shaping m/c may be reciprocated either in the vertical or in
the horizontal axis.
5.4.3 Advantages
➢ The gears produced by the method are of very high accuracy.
➢ Both internal & external gears can be cut by this process.
➢ Non – conventional types of gears can also be cut by this method.
5.4.4 Disadvantages
➢ The production rate with gear shaper is lower than Hobbing
➢ There is no cutting on the return stroke in a gear shaper
➢ Worm & worm wheels can’t be generated on a gear shaper.
5.5 GEAR CUTTING BY MILLING
5.5.1 Disc type cutter
For cutting a gear on a milling machine, the gear lank is mounted on am arbour
which is supported between a dead centre & a lieu centre in the in dering head. The
cutter is mounted on the arbour of the cutter must be aligned exactly vertically with the
centre line of the indexing head spindle. The table of machine is moved upward until
the cutter just touches the periphery of gear blank. The vertical feed dial is set to zero.
The table is then moved horizontally until the cutter clears the gear b lank. The table is
then moved upwards by an amount equal to the full depth of the gear tooth The vertical
movement may be less if the gear is to be cut in two or more passes After this, the

63
longitudinal feed of the table is engaged. The gear blank moves under the rotating
cutter & a tooth space is cut. After this, the movement of the table is reversed so that
the cutter again clears the gear blank. The gear blank is then indexed to the next
position for cutting the second tooth space. This procedure is repeated until all the
teeth have been milled. There is a flat circular disc type cutter and the plane of rotation
of the cutter is radial with respect to the blank.
5.5.2 End Milling cutter
In this method the cutter rotates about am axis which is set racially with respect
to the blank & at the same time the cutter is traversed parallel to the axes of the blank
The cutting edge tie on a surface of revolution, So that any axial cross- section of the
cutter corresponds to the shape required for the space between two adjacent teeth on
the finished wheel. The milling machine used in this method is vertical milling
machine The End mill cutter is mounted straight on the milling m/c spindle through a
chuck.
1) The disc type of cutter is used to cut big spur gear of cutter is Employed for the
manufacture of pinion of large pitch.
2) This method is very slow since only one tooth is cut at a time. To overcome these
drawbacks, “multiple tools shaping cutter head” is used to cut all the tooth spaces of
the gear at the same time.
5.5.3 Advantages
➢ Gear milling is a simple, Economical & flexible method of gear making.
➢ Spur, helical, bevel gears and racks can be produced by this method.
The major disadvantage of this method is that a separate cutter must be used not
only for every piton but for every no. of teeth.
5.6 Bevel Gear Generating
The teeth of bevel gears constantly change in form, from the large to the small Encl
There are to common types of bevel – gear generators, on cuts straight teeth & other
cuts spiral teeth.

64
5.6.1 Straight Bevel – gear generator
For generating straight – bevel gears, the rolling motions of two pitch cones are
employed motions of two pitch cones are employed instead of pitch cylinder. In this
method, two reciprocating tools which work on top & bottom sides of a tooth & are
carried on the machine cradle. The cradle & work roll up together with the gear blank
at the top of roll, when a tooth has been completely generated, the work is withdrawn
from the tool and the machine inclined, while the cradle is rolled down to the starting
position. The operating cycle is repeated automatically until all the teeth in the gear
have been cut.
The advantages of this process are that a previous roughening cut is not necessary,
thus saucing one handling of the blank, longer cutter life, improved quality of gear and
less set – up time
5.6.2 Spiral bevel –gear Generator
In this method, a rotating circular cutter generates spiral teeth that are curved &
oblique proper tooth profile shapes are obtained by relative motion in the machine
between work cutter. The machine has adjustment by which both spiral – bevel gears
& hypoid gears can be generated.
Spiral bevel gears have an advantage have on advantage over straight bevel
gear is that teeth are Engage with one another gradually by eliminating any noise &
shock in their operation.
5.6.3 Gleason Method
In this method, two disc milling cutters are employed, fig. The tools form the
blanks of a tooth simulating the basic crown wheel. Cutter teeth are inter – meshing
and the discs are inclined to each other at the pressure angle (usually 20*).
The following motions are involved while cutting a tooth:
1. The rotating cutters revolve about their axes to provide the cutting action
2. They travel in planes passing through the sides of the teeth on the imaginary
crown gear to shape the teeth along their teeth.
3. At the same time, they participate in the relative rolling motion between the
cutters and blank to obtain the required tooth profile.

65
Indexing takes place after each tooth space has been completed and the machine is
fully automatic in its motions. When gear has been completed, the machine stops, the
cutters withdraws the work piece can be changed with little delay. This type of
machine is a high production rate machine and very useful for dealing with large
batches of identical gears.
5.7 Gear finishing process
The following processes are generally used for finishing of gears
5.7.1 Gear shaving
Gear sharing is the most common method for gear finishing. In this method, a
very hard gear is used to ramous fine chips from the gear – tooth profile. The sharing
cutter can be: Rotary type or Rack type in rotary shoring, the cutter & the gear are run
in mesh. As they rotate, the gear is traversed longitudinally across the shaving cutter or
vie versa. The rotary sharing cutter has a member of peripheral gashes or grooves to
from a series of cutting Edges. The cutter & Gear are set up in a gear shoring m/c with
crossed axes in the form of spiral gearing. The usual angles are 10* to 15*.
In rock sharing, the cutter is in the form of a rack. During the operation, the
gear is rolled in mesh with the cutter. The cutter is reciprocated & at the End of Each
stroke is fed into the year
5.7.2 Gear grindings
Grindings is the most accurate method of gear finishing. By grinding, teeth can
be finished either by generation or forming. In forming, the work is made to roll in
contact with a fiat faced rotating grinding wheel, corresponding to the face of the
imaginary rack meshing with the gear. One side of the tooth is ground at a time after
the grinding wheel is given the shape by space b/w two adjacent teeth. Both flanks are
finished together.
The second method tends to be rather quicker, but both give equally accurate
results and which of the methods is to be used depends upon the availability of the type
of grinding machine.

66
Disadvantages
➢ Considerable time is consumed in the process
➢ Low production capacity
➢ Grinding wheels are Expensive.
➢ Gear lopping
It is another extensively used process of gear finishing & it is accomplished by
having the gear in contact with one or more cast iron lap gear of true shape the work is
mounted b/w centre & is slowly driven by rear lap. It is in term driven the front lap &
at the same time both laps are rapidly reciprocated across the gear face. Each lap has
individual adjustment & pressure control. A fine abrasive is used with kerosene or
light oil to assist the cutting action. The largest time of gear lapping is about 15
minutes. Prolonged lapping damages the profile.[13]
5.7.3 Shot blasting
It provides a finishing process resembling that produced by lapping although it
has other functions, such as removing slight burrs, reducing stress concentration in
tooth fillets & sometimes providing slight tip & root relief to teeth
5.7.4 Phosphate coating
It is a chemical process which attacks the treated ferrous surface and leaves a
deposit on it about 0.01 mm. in thickness. It prevents from scuffing, particularly in
hypoid gears, by permitting the Engaging tooth Surface under the prevailing boundary
lubrication conditions.
5.8 Gear planning
This is one of the oldest methods of gear production but is still extensively used.
It employs rack type cutters for generation of spur & helical gears. Involutes rack has
straight Edges & sharp corners can be (Easily) manufactured easily & accurately There
are two types of gear planning machines, one based on ‘The Sunderland process & the
other on ‘The Maag process’ Both the methods are identical in principle but differ in
m/c configuration & detail.
5.8.1 The Sunderland process
In this method, the work (gear balance) is mounted with axis horizontal & the

67
cutter slide is carried on a saddle position that moves vertically downward as cutting
proceeds. For cutting super gears, the cutter reciprocates parallel to the work axis (but)
because it can be swivelled in the vertical plane to any desired angle. The m/c is also
used for cutting single helical gears. The cutter is gradually fed to the desired depth of
teeth after which the depth remains constant. Simultaneously the gear blank is rotating
& rack is traversed at a tangent, the motion of rack & blank being geared to act on their
respective pitch lines. This relative motion beings fresh part of the blank & rack into
contact & thus causes the teeth of the cutter to generate wheel teeth of the cutter to
generate wheel teeth. The indexing really consisting slopping the rotation of the blank
& causing the rack to moue. The process is repeated until the blank has completed one
revolution.
5.8.2 The Maag process
In this method, the work is mounted on the m/c table with its axis vertical. The
rack cutter is carried in a cutter head: that is made to moue in a vertical plane but the
actual direction of motion can be set at any desired angle.
5.8.3 Principle of gear planning
The cutter during its cutting stroke is in contact with several teeth at the same
time but with different part of each tooth, it planes comparatively a narrow strip on
each tooth at each stroke and a different part of each tooth is submitted to the action of
the cutter at the next stroke.
5.8.4 Photographs taken while fabrication
Figure 5.1 Raw material

68
Figure 5.2 CNC milling
Figure 5.3 Pinion
Figure 5.4 Gear box casing under machining

69
5.9 INSTALLATION TO THE VEHICLE
Figure 5.5 Gearbox installed to the vehicle

70
CHAPTER 6
RESULTS & DISCUSSION
Our team suggested three concepts of design, which are automatic, manual
and CVT transmission design. Top two choices were chosen to further analysis. After
that we analysed the overall system to find what desired torque and speed which
turned to be 290 lb-ft and 40 mph respectively. As for the CVT system, it has 300.19
lb-ft torque and 51.70 mph which exceeds our expectations and should meet our goals
considering friction force and power lost. Based on our results we decided to
complete the transmission system with the Continuous Variable Transmission.
S.No Specifications Value
1 Purpose SAE BAJA competition
2 Compatible with CVTech CVT
3 Differential NO
4 Gear reduction 10.4:1
5 Kerb Weight 6.0 Kgs
6 Length (in) 12.84
7 Height (in) 7.73
8 Width (in) 4.13
9 Lubrication system Splash
10 Oil capacity (ml) 1000
11 Mounting Flat base
12 Mounting points 4
13 Input – Output Centre distance 160 mm

71
Moreover, the axle braking is to slow down or stop the vehicle safely and effectively
by converting kinetic energy into heat. It is the one of the main safety of a vehicle.
The goal of the project is to develop a lightweight, compact gear reduction that will
increase the efficiency and durability of the vehicle. And also, in addition we have
provided hydraulic braking for the drive shaft (i.e. known as inboard braking) for
which the calliper is mounted to the gearbox itself.
Finally, a high level of manufacturability was incorporated to ensure feasibility for
mass-production.

72
CHAPTER 7
CONCLUSION
This being Team Wraith Racing, our objective was to design a customised
gearbox with inboard braking is typically designed as a reduction after the
Continuously Variable Transmission and build an All Terrain Vehicle that can
complete all competition events without any failure. All designs and calculations were
done to realise this aim.
Reliability and safety were considered paramount, keeping the nature of the
end-user in mind. They also need to demonstrate economic viability of the project by
doing cost analysis and marketing presentation. It is technically challenging and in the
meantime, involves many other aspects of a modern engineering enterprise, such as
people skills. This year long project was consummation and highlight of an
undergraduate education experience. We learned so much more than what we would
normally learn from a classroom setting.

73
REFERENCES
[1] Bajasaeindia.com(2018).[online] retrieved from: http://bajasaeindia.com
[2] Khurmi, R.S & Gupta, J.K – 2003. Theory of Machines. S.Chand & Co.
[3] Khurmi, R.S & Gupta, J.K – 1996. Text Book of machine design.
[4] PSG college of Design data book Kalakhathir.Coimbatore 2010.
[5] Shigley, J.E (1956). Machine Design. New York: McGraw-Hill.
[6] Swapnil R. Abhang, D.P Bhaskar, International journal of engineering trends and
technology (IJETT), “Design analysis of disc brake”, Volume 8 Number 4-Feb 2014.
[7] CVTech-AAB
Available:
http://www.numericquetechnologies.com/cvtech/CatalogueCVTech
AAB_US_%202013.pdf
[8] Seamless AMT offers efficient alternative to CVT
Available:
http://www.zzeroshift.com/pdf/Seamless%20AMT%20Offers%20Efficient%20Altern
ative%20To%20CVT.pdf
[9] Richard Budynas, and J Keith Nisbett. Mechanical Engineering Design. 9th
.1021.
New York: McGraw Hoill, 2011. Print.
[10] ‘Milliken, W. F., and Milliken, D. L.,’ “Race Car Vehicle Dynamics”, SAE Inc.
Milliken, 1995
[11] Gill, Simren., Hay,Philip., Mckenna, patrick., and Thibodeau, Philippe;”Final
Design” Dalhouse Formula SAE ” April 8, 2011.
[12] ‘George E. Totten, D. Scott MacKenzie’, “Handbook of Aluminum: Vol. 1:
Physical Metallurgy and Processes”, CRC Press, 2003.
[13] ‘Bandari’, “Design of Machine Elements”, McGraw Hill, 2017.
[14] ‘Suhaimi,Khalis bin’ ”Design And Fabrication Of An Upright With Brake
Caliper Mounting For Formula Versality Race Car” dissertation, april 2011.
[15] Thomas D. Gillespie; Fundamental of Vehicle Dynamics; ISBN: 978-1-56091-

74
199-9; February 1992
[16] ‘Carroll Smith,’ “Weight, Mass Load and Load Transfer” in Tune to Win, Aero
Publishers, INC, 329 West Aviation Road, Fallbrook, CA 92028
[17] Ramamurti V., Sukumar T., Mithun S., Prabhakar N. & Hudson P. V., Design
analysis of Hub, Rim and Drum in Brake Assembly, Mechanical Engineering
Research; Vol. 3, No. 1; 2013
[18]https://www.researchgate.net/publication/228595865_Gear_Noise_and_Vibration
-A_Literature_Survey
[19] shodhganga.inflibnet.ac.in/bitstream/10603/144009/10/10_chapter%202.pdf
[20] http://www.irdindia.in/journal_ijmer/pdf/vol3_iss4/7.pdf

75
APPENDIX
➢ This Appendix proved the visual aids for complete assembly of transmission
system of CVT .
i. Side view
ii. Rear view

76
iii. Top view
iv Isometric view

77
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