i

PROJECT REPORT ON

MODIFICATION OF A FORMULA STUDENT CAR GEARBOX

Submitted in Partial Fulfilment of the requirements

for the degree of

BACHELOR OF ENGINEERING

BY

Inder Singh Sehra

&

Munjal Savla

UNDER THE GUIDANCE OF:

Prof. B.M. Pradhan

Department of Mechanical Engineering

K. J. SOMAIYA COLLEGE OF ENGINEERING, MUMBAI

UNIVERSITY OF MUMBAI

2009 - 2010

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Certificate

This is to certify that the project entitled

MODIFICATION OF A FORMULA STUDENT CAR GEARBOX

submitted by

Inder Singh Sehra

&

Munjal Savla

in Partial Fulfilment of the degree of B.E. in Mechanical Engineering is approved.

K. J. SOMAIYA COLLEGE OF ENGINEERING, MUMBAI

UNIVERSITY OF MUMBAI

2009 - 2010

Prof. B. M. Pradhan

Guide & Head of the Department

Department of Mechanical

Engineering

Internal examiner

Dr. (Mrs.) Medha Dixit

Principal

External examiner

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Abstract:

A number of racing series have seen the influx of motorcycle engines as basis power

plants which incorporate a performance-oriented sequential shift transmission. Most

of these engines have a 1-N-2-3-4-5-6 International standard shift pattern. This shift

pattern was introduced in bike due to safety concerns and gives a lot of problems

when the engine was used in a car.

This project focuses on changing the shift pattern of the gearbox to N-1-2-3. It also

explains the development of a lift-less low cost light weight electronic shift system

which uses a 24V DC motor to change the gears. This system reduces the shift time

from 1500ms (Manual Shift) to 500ms.

iv

Acknowledgements:

We are very grateful to Mr. Samir Somaiya for starting this project and for his

constant support all these years. We are also grateful to our team members both

juniors and seniors without whose support this project would not have been possible.

We are grateful to our project guide Prof. B.M.Pradhan for giving us an opportunity

to work under his guidance. We express our sincere thanks to Mr. Saiju (Proprietor)

of Trepko Micron Industries and Mr. Gurmeel Singh Washist (Proprietor) of

Ambota Steel Sales for advice, support and sponsoring the material and the

manufacturing process. We are also thankful to Protosys Technologies Pvt. Ltd. for

sponsoring the manufacturing of the prototype of the shifter drum.

Without their support this project would not have been possible, for which we

express our gratitude.

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Table of Contents

List of Figures ................................................................................................. 3

List of Tables .................................................................................................. 5

Nomenclature ................................................................................................. 6

Introduction… ................................................................................................. 8

Formula Student Competition ....................................................................... 8

Orion Racing India ........................................................................................ 9

Objectives of the Project ............................................................................. 10

Chapter 1: The Constant - Mesh Gearbox .............................................. 11

1.1 The Constant - Mesh Gearbox and its Working................................. 11

1.2 Constant-Mesh Sequential Gearbox ................................................. 12

1.3 Selector Mechanisms ........................................................................ 13

1.3.1 Sliding-Type Selector Mechanisms ............................................. 13

1.3.2 Ball-type selector mechanism ..................................................... 14

1.3.3 Drum Shifter Mechanism ............................................................ 16

1.3.4 Camplate and Shift Quadrant Shifter Mechanism ....................... 17

1.3.5 Ball-Lock Shifter Mechanism ...................................................... 18

Chapter 2: The Honda CBR600 F4i Gearbox.......................................... 20

2.1 Working and Construction ................................................................. 20

2.2 The Shift Pattern ............................................................................... 25

2.3 Problem with the current selector mechanism ................................... 26

2.4 Solution Adopted ............................................................................... 26

Chapter 3: Cam Design ............................................................................ 27

3.1 Design Methodology .......................................................................... 27

3.2 Design of the Shifter Cam ................................................................. 27

3.2.1 Selection of the S V A J Functions .............................................. 28

3.2.2 Calculations for Cam Path 1 ....................................................... 30

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3.2.3 Calculations for Path 2 ................................................................ 35

Chapter 4: CAD Modelling and Prototyping........................................... 41

4.1 Design of the Prototype ..................................................................... 41

4.2 Final Model of the Shifter Drum ......................................................... 42

Chapter 5: Material Selection and Manufacturing ................................. 45

5.1 Material Selection .............................................................................. 45

5.2 Manufacturing of the Shifter drum ..................................................... 46

Chapter 6: Electronic Gearshift Mechanism .......................................... 47

6.1 Requirements of the Shifter ............................................................... 47

6.2 Pneumatic v/s Electric shifting systems ............................................. 47

6.3 Previous year’s Electronic shifting mechanism.................................. 48

6.4 Initial Concept and Testing ................................................................ 48

6.5 The Final Product .............................................................................. 49

6.6 Testing .............................................................................................. 50

Chapter 7: Future Work and Conclusion................................................ 51

7.1 Future work ....................................................................................... 51

7.2 Conclusion......................................................................................... 51

References.… ............................................................................................... 52

Books .......................................................................................................... 52

Reports ....................................................................................................... 52

World Wide Web ......................................................................................... 53

Appendix A… ................................................................................................ 54

Production Drawings ................................................................................... 54

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List of Figures

Figure 0.1: CAD model of ORI2010 .................................................................. 9

Figure 1.1:Simple Constant Mesh Gearbox.................................................... 11

Figure 1.2: Sliding type Selector ..................................................................... 14

Figure 1.3: Ball Type Selector ........................................................................ 15

Figure 1.4: Shifter Fork for Ball type Selector ................................................. 16

Figure 1.5: Shifter Fork for Ball type Selector ................................................. 17

Figure 1.6: Camplate and Shift Quadrant Shifter Mechanism ........................ 18

Figure 1.7: Ball Lack Shifter Mechanism ........................................................ 19

Figure 2.1: Components of the F4i Gearbox .................................................. 20

Figure 2.2: Drawing of the Assembled Gearbox ............................................. 21

Figure 2.3: Shifter Fork ................................................................................... 22

Figure 2.4: Assembled Shifter forks on shifter shaft in the Engine ................. 22

Figure 2.5: Shifter Drum ................................................................................. 23

Figure 2.6: Shifter Drum Assembled in the Engine ......................................... 23

Figure 2.7: Shifter Drum Assembled in the Engine ......................................... 24

Figure 2.8: Shift Pawl ..................................................................................... 24

Figure 2.9: The Shift Pawl Assembled in the Engine ...................................... 25

Figure 3.1: SVAJ graphs for Harmonic Motion ............................................... 28

Figure 3.2: SVAJ graphs for Cycloidal motion ................................................ 29

Figure 3.3: Displacement Graph for Path One (Harmonic) ............................. 30

Figure 3.4: Velocity Graph for Path one (Harmonic) ....................................... 31

Figure 3.5: Acceleration Graph for path one (Harmonic) ................................ 32

Figure 3.6: Pressure Angle Graph for path one (Harmonic) ........................... 33

Figure 3.7: Displacement Graph for Path 2 (Harmonic).................................. 35

Figure 3.8: Velocity Graph for Path 2 (Harmonic) ........................................... 36

Figure 3.9: Acceleration Graph for Path 2 (Harmonic) ................................... 37

Figure 3.10: Pressure Angle for Path 2 (Harmonic) ........................................ 38

Figure 4.1: CAD Model of the Prototype ......................................................... 41

Figure 4.2: Rapid Prototype of the Shifter Drum ............................................. 42

Figure 4.3: Final Model of Main Cylinder ........................................................ 43

Figure 4.4: Final Model of Main Cylinder ........................................................ 43

Figure 4.5: Final Model of the Cap ................................................................. 44

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Figure 4.6: Final Assembly ............................................................................. 44

Figure 6.1: ORI2009’s Electronic shift mechanism ......................................... 48

Figure 6.2: The Motor used as Actuator ......................................................... 49

Figure 6.3: Initial Testing ................................................................................ 49

Figure 6.4: Motor with the Motor Lever ........................................................... 50

5

List of Tables

Table 3.1: Tabulated Results of Calculations for Path 1 ................................. 35

Table 3.2: Tabulated Results of Calculations for Path 2 ................................. 40

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Nomenclature:

Symbol Description

s Displacement of follower (mm)

h Maximum Displacement of follower(mm)

θ Angle Traversed by the cam (o)

β Total angle traversed by the cam(o)

v Velocity of follower(m/s)

a Acceleration of follower(m/s2)

φ Pressure Angle(o)

ω Angular Velocity of Cam (rad/s)

Rp Pitch Circle Radius of Cam (mm)

j Jerk of follower(m/s3)

F Inertial forces on Cam(N)

W Weight of accelerated elements (Kg)

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g Acceleration due to gravity (9.81 m/s2)

P Max. Parallel force on Cam (N)

S Spring Force on Follower (N)

L External Force on Follower(N)

µ Coefficient of friction

Pn Max. Normal Force (N)

Sc Contact Stress (N/mm2)

Rc Radius of curvature of Cam(mm)

rf Radius of follower (mm)

L’ Width of contact surface between Cam and Follower (mm)

Ec & Ef Elastic modulus of the material for Cam and Follower Materials

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Introduction

Formula Student Competition:

Formula Student is an international competition held world-wide by the Society

of Automotive Engineers (SAE).

Students build a single seat formula race car with which they can compete

against teams from all over the world. The competition is not won solely by the

team with the fastest car, but rather by the team with the best overall package

of construction, performance, and financial and sales planning.

Formula Student challenges the team members to go the extra step in their

education by incorporating into it intensive experience in building and

manufacturing as well as considering the economic aspects of the automotive

industry. Teams take on the assumption that they are a manufacturer

developing a prototype to be evaluated for production. The target audience is

the non-professional Weekend-Racer, for which the race car must show very

good driving characteristics such as acceleration, braking and handling. It

should be offered at a very reasonable cost and be reliable and dependable.

Additionally, the car's market value increases through other factors such as

aesthetics, comfort and the use of readily available, standard purchase

components.

The challenge the teams face is to compose a complete package consisting of

a well constructed race car and a sales plan that best matches these given

criteria. The decision is made by a jury of experts from the motorsport,

automotive and supplier industries. The jury will judge every team’s car and

sales plan based on construction, cost planning and sales presentation. The

rest of the judging will be done out on the track, where the students

demonstrate in a number of performance tests how well their self-built race

cars fare in their true environment.

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Orion Racing India

Orion Racing India is a team of engineering students from K.J. Somaiya

College of Engineering, Mumbai and participates in Formula SAE. The

competition is held in various countries around the globe however, Orion

Racing India participates in the event held at the Hockenheim Ring, Germany.

Formula SAE provides us a fabulous learning opportunity and gives us

tremendous "hands-on" experience, not to mention benchmark ourselves

against the best teams through this exposure to international competition.

The team has been successfully participating at Formula Student Germany

annually since 2007. The ORI2010 will be our fourth car.

Figure 0.1: CAD model of ORI2010

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Objectives of the Project

At the Formula Student Competition, the ranking of the team heavily depends

on the performance of the car at the dynamic events. These events are:

Acceleration

Skid – Pad

Autocross

Endurance

The objective of a race car is to be the fastest around a circuit. To achieve

this, the speed should be as high as possible at all parts of the circuit and time

wasted on non-performance activities should be minimum.

The time taken for gear shifts is wasted time, because the car is not

accelerating but being carried forward by its own momentum. Thus, the car

can be faster around a circuit by reducing the time taken for a gear shift.

This project concentrates on achieving this through two means:

Changing the shifting pattern of the sequential gearbox

Implementing an electronic shifting system

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Chapter 1: The Constant - Mesh Gearbox

In order to better understand what function a selector mechanism performs in

a gearbox it was necessary to know how a Constant-Mesh Gearbox works.

1.1 The Constant - Mesh Gearbox and its Working:

A constant mesh gearbox has various gears of which some can slide axially

on the shaft and some that have no axial freedom. The most common form of

constant-mesh gearbox is shown in the figure below. (The working of the

constant mesh gearbox is explained after the figure)

Figure 1.1: Simple Constant Mesh Gearbox

Here the engine shaft A is integral with the pinion B which meshes with the

wheel C on the layshaft .The latter is, therefore, driven by the engine shaft.

Wheels E ,F and G are fixed to the layshaft just as in a sliding mesh gearbox,

and the main shaft D is also Similarly Arranged .The gears E,F and G( the

latter through a reverse idler ) are , however free to turn on the main shaft

,bronze bushes , or ball or roller bearings, being provided between them and

the shaft. The gears H, J and I are therefore constantly driven by the engine

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shaft, but at different speeds, since the wheels E, F and G are of different

sizes.

If any one of the gears H, J or I is coupled up to the mainshaft then there will

be driving connection between that shaft and the engine shaft. The coupling is

done by means of dog clutch members L and M, which are carried on splined

portions of the mainshaft. They are free to slide on those splined portions, but

have to revolve with the shaft. If the member M is slid to the left it will couple

the wheel I to the mainshaft giving the first gear. The drive is then through

wheels B, C, F and I and the dog clutch M. The outer dog clutch is meanwhile

in its neutral position, the member L is slid to the right, it will couple the wheel

H to the mainshaft and give second gear, the drive being through the wheels

B,C,E and H and the dog clutch L .If the member L is slid to the left it will

couple the mainshaft directly to the pinion B and gibe direct drive, as in a

Sliding-mesh gearbox.

This type of gearbox has several advantages over the ordinary form of sliding-

mesh box. It facilitates the use of helical or double helical gear teeth which are

quiter than straight teeth ;it lends itself to the incorporation of synchronising

devices more readily than the sliding-mesh box ; the dog clutch teeth can be

made so that they are easier to engage than teeth of gear wheels ,and any

damage that results from faulty manipulation occurs to the dog clutch teeth

and not to the teeth of the gear wheels .Now, when once the dog clutches are

engaged there is no motion between their teeth ,whereas when gear teeth are

engaged the power is transmitted through the sliding action of the teeth of one

wheel on those of the other. The teeth have to be suitably shaped to be able to

transmit the motion properly, and if they are damaged the motion will be

imperfect and noise will result .Damage is, however, less likely to occur to the

teeth of the dog clutches, since all engage at once, whereas in sliding a pair of

gears into mesh the engagement is between two or three teeth.

1.2 Constant-Mesh Sequential Gearbox

A Constant-Mesh Sequential Gearbox is exactly the same as the Simple

Constant-Mesh Gearbox in which gears can only be engaged in a specific

pattern in an ascending or descending sequence i.e. no gear can be selected

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randomly (A simple constant-mesh gearbox can be converted into a

Sequential one by changing its selector mechanism). This type of a gearbox is

mostly used in motorcycles with an international shift pattern of 1-N-2-3-4-5-6.

1.3 Selector Mechanisms

The engagement and disengagement of gears or the sliding of the gears in a

gearbox is controlled by a selector mechanism. The sliding movement of the

gear is controlled by selector forks. The fork fits into a groove formed in the

boss of the gear to be moved, so that although the gear is left free to revolve,

it must partake of any sideways movement that is given to the fork .There will

be a selector fork for each sliding member in the gearbox. The selector forks

either slide on rods fixed in the gearbox casing or are fixed to rods which can

slide in that casing, the rods being parallel to the shafts upon which the gears

slide. The Necessary sliding motions are given to the selector forks by the

motion of a gear change lever actuated by the driver.

1.3.1 Sliding-Type Selector Mechanisms

The Sliding type selector mechanism is as shown in the figures below. Figure

1.2(1) shows the sectional elevation, the plane of the section being indicated

by the line SS in the end view Figure 1.2(2). The latter is a section on AB. The

third view is a part plan there are three moving members in the gearbox into

which this particular mechanism is fitted so that there are three selector forks,

C, D and E. The forks C and D slide on rods F and G fixed in the casing, while

E is carried by a pivoted lever Q which is actuated by a member that slides on

the third rod H. The forks are moved by a fore and aft rocking motion of the

gear lever J which is carried by a shaft L pivoted in the casing and to the inner

end of which is secured the striking lever K. The particular fork that is to be

moved is selected by a sideways sliding motion of the member JLK. To hold

the forks in their various positions spring plungers, one of which is seen in

1.2(1), and which spring into grooves cut in the rods FGH, are fitted.

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Figure 1.2: Sliding type Selector

To prevent two forks from being moved at once a locking piece M is provided.

This slides on a cross rod N fixed in the casing, and is provided with horns O

and P which project into the slots in the sliding members. Between the horns

O and P is situated the end of the striking lever K so that the sideways

movement of the latter causes the member M to slide on its rod. The gap

between the horns O and P is only slightly wider than each of the sliding

members, so that the latter can be moved only one at a time.

1.3.2 Ball-type selector mechanism

In this form of selector mechanism, shown in Figure 1.3, the control lever is

mounted on the transmission casing. The selector forks A and B slide on rods

fixed in the gearbox lid, which in this design carries the whole of the selector

mechanism. The shape of the forks is shown by the perspective sketch Figure

10; they are provided with slots C to receive the end D of the striking arm. The

latter is the lower end of the gear lever E which is ball jointed in the casing at

F. By rocking the lever sideways its end D may be brought into engagement

15

with either of the selector forks, when a fore and aft rocking motion will slide

that fork along its rod.

Figure 1.3: Ball Type Selector

No gate is provided, but small plungers G and H prevent both forks from being

moved at once. When both the forks are in the neutral position and the slots C

are opposite each other, the plungers are forced by small springs into holes in

the forks, and before either fork can be moved the plunger that locks it must

be pressed back into the casing. This is done by the sideways motion of the

gear lever, obviously when one plunger is pressed in to release one of the

forks the other plunger is out, and is locking the other fork. These plungers

also serve to lock the forks in position when the gears are properly engaged,

being arranged to spring into shallow recesses NN in the forks.

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Figure 1.4: Shifter Fork for Ball type Selector

The above mentioned selector mechanisms were mostly used on cars and are

now obsolete .Now newer and simpler sequential selector mechanism are

used, some of these are explained hereafter

1.3.3 Drum Shifter Mechanism

This mechanism is the most widely used mechanism in constant-mesh

gearboxes; the Honda CBR600 F4i also uses a similar kind of a mechanism. It

is simple, easy to operate and reliable and hence is a favourite among bike

manufacturers. It comes in two flavours, one type carries the shifter forks on

the drum (shown in Figure 1.5) and with the other type, the forks are carried

on their own shafts (shown in Figure 2.6). Both types use grooves cut into the

shifter drum, to move the forks back and forth. When the drum turns, the forks

move back and forth, moving the gears in and out of engagement. These

drums are turned by a mechanism known as the shift pawl.

To shift gears, the drum should only turn by a small amount and then stop.

The shift pawl does that, it only turns the drum a set amount, each time the

pawl is moved. Each time the gear shift lever is pressed down the pawl turns

the drum the same amount. If the shift lever is pulled up, the pawl reverses

and moves the drum in the opposite direction, the same amount. The shift

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lever moves the pawl and the pawl moves the drum. The drum, in turn, moves

the shift forks and the shift forks move the gears in and out of mesh. The pawl

presses against pins in the shift drum to rotate it, then a spring pulls the pawl

back to a middle position and the drum rotates once and stops there. To keep

the drum from turning, a wheel called as a Shifter Detent, Shifter Drum

stopper, or Shifter Cam Stopper, moves into grooves on the shifter drum,

locking it into position. This wheel is spring loaded shown in Figure 2.7. This

enables the wheel to move in and out of position, as the drum is turned by the

pawl. Sometimes, instead of a wheel, a spring loaded plunger is used. This

plunger has a rounded head that fits into Shifter Cam or holes in the shifter

drum. This cams, or holes, have bevelled sides allowing the plunger to move

smoothly in and out of the hole as the drum is turned. This type of shifter

mechanism is mostly used by Japanese manufacturers.

Figure 1.5: Shifter Fork for Ball type Selector

1.3.4 Camplate and Shift Quadrant Shifter Mechanism

In Camplate and Shift Quadrant shifter mechanisms the shifter forks are

carried on a plate, Grooves in this plate allow the forks to move the gears back

and forth. There are two styles of this type of shifter. One has the grooves for

the shifter forks cut right through the Shift Quadrant itself. The other has the

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grooves machined into a Cam Plate that the Shift Quadrant moves. This

mechanism is used mostly on British and European Bikes.

Figure 1.6: Camplate and Shift Quadrant Shifter Mechanism

1.3.5 Ball-Lock Shifter Mechanism

In ball lock systems, one gear shaft has all the gears machined on the shaft.

The other shaft is hollow and has four holes in each gear position. There is a

ball bearing in each of these holes. A gear rides around each set of four holes;

the gears have four indentions cut on the inside of the gear. A Shifter

Head moves back and forth, inside the shaft. It pushes the balls out and into

the indentions in each gear. This changes the gear. When the next gear is

chosen, the balls fall back into the shaft releasing the gear. Because all the

gears are always engaged with each other, one does not have to use the

clutch to change gears; the clutch is only used in first gear to get started, but

after that it is not used. This mechanism was developed by Hodaka Motorcycle

Company and is not used anymore. This is because of the high horsepower of

today’s motorcycle engines.

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Figure 1.7: Ball Lack Shifter Mechanism

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Chapter 2: The Honda CBR600 F4i Gearbox

This chapter explains with pictures the working and construction of the

gearbox whose selector mechanism is to be modified.

2.1 Working and Construction

The FSAE Competition permits use of engines whose displacement is no

more than 610cc. Hence, the Honda CBR600 F4i Engine with its simple

design and lightweight was the obvious choice for the competition. It has a

Constant-Mesh Sequential Gearbox, with a 1-N-2-3-4-5-6 shift pattern, which

is an integral part of the engine.

The construction and working details of the gearbox are explained as follows.

Figure 2.1: Components of the F4i Gearbox

As seen in the picture A through D are the bearings on which the main and the

countershafts rotate.

Gears E through I are mounted on the Mainshaft which gets its drive from the

crankshaft through the clutch , the gears E,I are fixed with the mainshaft and

gears F and H are free to rotate on the shaft and the gear G is free to slide on

the mainshaft.

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Similarly gears E’ through I’ are mounted on the countershaft .Gears E’, G’,

G”, I’ are free to rotate on the countershaft and gears F’ and H’ are free to

slide.

Figure 2.2: Drawing of the Assembled Gearbox

The gears are engaged and disengaged with the help of shifter forks and a

shifter drum as explained previously in the Section 1.1.

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Figure 2.3: Shifter Fork

Figure 2.4: Assembled Shifter forks on shifter shaft in the Engine

23

Figure 2.5: Shifter Drum

Figure 2.6: Shifter Drum Assembled in the Engine

24

Figure 2.7: Shifter Drum Assembled in the Engine

The Shifter Drum is rotated (to Shift a gear) with the help of a Shift Pawl

mechanism. This shift pawl is such a mechanism that at one time only one

shift can take place. The shift pawl is shown below.

Figure 2.8: Shift Pawl

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Figure 2.9: The Shift Pawl Assembled in the Engine

2.2 The Shift Pattern

The gearbox has an international standard 1-N-2-3-4-5-6 gearshift pattern.

The reason this pattern is used in most of the bikes made nowadays is

SAFETY. This can be explained using the following example , Suppose that

you have a bike which has a N-1-2-3-4-5 pattern and you are cruising at a

good speed and you see a red light and have to stop .As you slow down you

downshift to neutral at the right gear speeds , now say that the light changes

to green and you are still carrying considerable speed and you shift from

neutral to first , Since you are carrying considerable speed , your rear wheels

will lock due to engine braking phenomenon , this can result in an accident

and serious injury . Now consider the same scenario with a 1-N-2-3-4-5-6

pattern. Here you shift into first instead of neutral with the clutch disengaged to

stop , now when you see the green light you upshift to neutral or second gear

and hence you experience no or less engine braking . Hence, this pattern is

safer than the previous pattern. The gear shift pattern 1-N-2-3-4-5-6 is only

advantageous in a bike and causes problems when the same engine was

used on a car.

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2.3 Problem with the current selector mechanism

Drivers complained about the car accidentally shifting into neutral instead of

2nd when shifting from 1st to 2nd gear, this is somewhat tolerable in the

endurance event but can increase acceleration times in the acceleration event,

which is not acceptable. Hence a solution to this problem had to be found out

in order to improve the performance of the car.

2.4 Solution Adopted

Teams from other universities who faced the same problem removed the first

gear and welded of the grooves for the first gear on the shifter drum. The first

gear was ground off and 2nd gear was now used as the first gear. To

compensate for the loss in acceleration, the final drive ratio was increased.

This is the easiest, cheapest and most widely used solution; however when

the final drive ratio was increased so did the size of the sprockets, this in turn

increases the weight of the car and also causes packaging issues .Hence it

was decided to design and manufacture a new shifter drum with N-1-2-3 shift

pattern, Since in the competition the highest gear that the car goes into is 3, it

was decided to eliminate the remaining gears and design a shifter for only 3

gears.

27

Chapter 3: Cam Design

This chapter explains the design of the cam grooves on the shifter drum and

the calculations and approximations involved in it.

3.1 Design Methodology

Anything that goes onto the car has to be put through a rigorous design

process to ensure that it is the best part for the desired function. To check the

reliability of the designed component a lot of testing has to be done. The part

designed should be light in weight, should be reliable, easy and cheap to

manufacture. Keeping in mind all of these factors one should come up with a

balanced design. A component that is too bulky, difficult to manufacture will

not find its way onto the car.

3.2 Design of the Shifter Cam

The shifter cam is basically a critical extreme position cam CEP i.e. only the

start and the end positions of the follower matter and the path that was taken

by the follower to travel from start to the end does not matter. Since the type of

path of the original cam is not known, we had to select our own path from the

various types e.g. Cycloidal, Harmonic, Double harmonic etc. Thus, a

comparison was made between all the available paths.

The fundamental law of CAM design is

“The cam-follower function must be continuous through the first and second

derivatives of second derivatives of displacement i.e. Velocity and acceleration

across the entire interval”

Corollary:

The jerk function must be finite across the entire interval.

In the competition the highest gear that the car goes to is 3, hence the path for

the centre fork is eliminated and only the paths at the ends are used. Various

types of motions were analysed and it was decided to make two cams, one

with harmonic paths and the second with cycloidal paths. The following part

explains the reason these paths were selected.

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3.2.1 Selection of the S V A J Functions

The application has the following requirements

Path 1:

Displacement: 5.3 mm

Single Dwell Cam (Rise-Fall-Dwell)

Path 2:

Displacement: 5.3mm (to either side)

Single Dwell Cam (Dwell-Rise-Fall)

The following graphs show the comparison between Harmonic, Cycloidal

motions

Figure 3.1: SVAJ graphs for Harmonic Motion

From the above graph it is seen that the simple harmonic function has an

infinite jerk function at the ends. Thus it violates the fundamental law for Cam

design.

29

Figure 3.2: SVAJ graphs for Cycloidal motion

Since, acceleration function for cycloidal motion is continuous at the end, it

does not have infinite jerk at the endpoints. Thus it does not violate the

fundamental law of cam design.

The shifter cam is not a continuous motion cam; it has intermittent motion i.e.

after 60o of rotation the cam stops for a few seconds until the next gear

change occurs.

Thus, it can be considered as a very low speed intermittent motion cam and

hence even if a function violates the fundamental law of cam design it can be

used in the design. The advantage of the harmonic function over cycloidal

function is that it has a comparatively low maximum pressure angle value due

to a smoother curve than the cycloidal function which results in lower side

thrust on the follower, due to a lower side thrust on the follower a less amount

of torque is required to rotate the shifter cam. The only problem with the

harmonic function is the infinite jerk at the ends of the motion. Since the

behaviour of the cam and follower when they experience a high value of jerk is

unknown, it was decided to go ahead with the harmonic function and test the

shifter cam in actual operation to test if it operated correctly. The cycloidal

shifter cam would also be manufactured later and the best of the shifter cams

would be implemented on the car. The calculations for the harmonic function

are shown below

30

3.2.2 Calculations for Cam Path 1:

Displacement calculations at critical points:

Where, β outside sine is the time taken by the cam to turn through angle βo

S30=5.3(1-cos(π*30/60)) = 2.533mm

S60 = 5.3mm , S90 =2.75mm , S120=S180=0mm

Figure 3.3: Displacement Graph for Path One (Harmonic)

3.2.2.2 Velocity Calculations at Peak points:

It was found out that the shifter cam takes about 60 to 100ms to rotate through

an angle of 60o.

V+max = π*5.3*0.5/0.1*sin(π*30/60) = 0.083m/s.

V-max = -0.083m/s.

31

Figure 3.4: Velocity Graph for Path one (Harmonic)

3.2.2.3 Acceleration Calculations at Peak Points:

A-max = π2 *0.5*5.3/0.12cos (π*60/60) = - 2.615 m/s2

A+max = 2.615 m/s2

32

Figure 3.5: Acceleration Graph for path one (Harmonic)

3.2.2.4 Jerk Calculations:

Jerk reaches infinity at the end points (Due to this reason it cannot be plotted).

The peak value of jerk is calculated as follows

Jmax = - π3*5.3*0.5/0.13 sin (π*30/60) = - 82.16 m/s3

3.2.2.5 Pressure Angle Calculations:

Where, V (θ) = Follower Velocity.

ω = Angular velocity of the Cam

Rp = Pitch circle Radius

Φ 30 = tan-1 (0.083/ (10.47*0.0021)) = 22o

Φ 60 = 0, Φ 90 = -22, Φ 120 = Φ 180 =0.

33

Figure 3.6: Pressure Angle Graph for path one (Harmonic)

3.2.2.6 Calculation of inertial forces (F):

F= W*a/g

Where, W = Weight of the follower system = 200 gms.

Forces at the peak acceleration = F = 0.2*10*2.6/9.81= 0.53 N.

3.2.2.7 Calculation of Max Parallel Force (P):

This force acts in a direction parallel to the Cam Axis

Here, L = External force is assumed to be 0

S = spring force = 0

µ = 0 (as the Shifter is very well lubricated)

Hence,

P = +-(W/g) a + W

Thus, Pmax = 0.53 + 0.2 = 0.73 N

Pmin = -0.53 +0.2 = -0.33 N

34

3.2.2.8 Calculation of Max Normal Force (Pn):

This force acts in a direction normal to the cam curved surface

P n = P/tan φ

P n max = 0.73 / cos 0 = 0.73N

3.2.2.9 Calculation of Max Contact Stress Sc:

Contact stress was approximated based on the hertz contact stress theory as

follows.

Calculation of the contact stress is a tedious task, hence it is better to

approximately calculate it.

Assuming that the contact occurs between a cylinder (the follower) and the flat

surface of the cam, then Rc = ∞, and assuming that Ec = Ef = 206 Gpa =

21x104 N/mm2. We get

Max Sc = √ ( (0.35*0.73/4.05)/(8*2/21*104)) = 28.7753N/mm 2 = 4333.72 PSI

Including a Factor of Safety to account for the approximation

f = 5

Max Sc =21666.5 PSI or 146 N/mm2

The contact stresses should not exceed 1/3rd the ultimate compressive

strength of the material. The value of Ultimate compressive strength of steels

is about 800 to 1500 N/mm2. Hence, the design is safe under contact stress.

35

3.2.2.10 Tabulated Results of Calculations for Path 1:

Angle s

(mm)

v

(m/s)

a

(m/s

2)

j

(m/s3) φ

Wa/g

(N)

P

(N)

Pn

(N)

Max Sc

(N/mm2)

0 0 0 2.6 ∞ 0 0.53 0.73 0.73

145.8

30 2.533 0.083 0 -81.8 22 0 0.2 0.215

60 5.3 0 -2.6 ∞ 0 -0.53 -0.33 -0.66

90 2.75 -0.083 0 81.8 -22 0 0.2 0.215

120 0 0 2.6 ∞ 0 0.53 0.73 0.73

180 0 0 0 0 0 0 0.2 0

Table 3.1: Tabulated Results of Calculations for Path 1

3.2.3 Calculations for Path 2:

3.2.3.1 Displacement calculations at critical points:

Where, β outside sine is the time taken by the cam to turn through angle βo

S30=0(1-cos(π*30/60)) = 0mm

S60 = 0mm , S90 =-2.641mm , S120=-5.27mm S180=5.84mm

Figure 3.7: Displacement Graph for Path 2 (Harmonic)

36

3.2.3.2 Velocity Calculations at Peak points:

It was found out that the shifter cam takes about 60 to 100ms to rotate through

an angle of 60o.

V+max = π*-5.3*0.5/0.1*sin (π*30/60) = -0.083m/s.

V-max = 0.166m/s.

Figure 3.8: Velocity Graph for Path 2 (Harmonic)

3.2.3.3 Acceleration Calculations at Peak Points:

A-max = π2 *0.5*5.3*2/0.12cos (π*60/60) = - 5.234 m/s2

A+max = π2 *0.5*5.3*2/0.12cos (π*0/60) = 5.234 m/s2

37

Figure 3.9: Acceleration Graph for Path 2 (Harmonic)

3.2.3.4 Jerk Calculations:

Jerk reaches infinity at the end points (Due to this reason it cannot be plotted).

The peak value of jerk is calculated as follows

Jmax = - π3*5.3*0.5*2/0.13 sin (π*30/60) = -164.333 m/s3

3.2.3.5 Pressure Angle Calculations:

Where, V (θ) = Follower Velocity.

ω = Angular velocity of the Cam

Rp = Pitch circle Radius

Φ 30 = tan-1 (0/(10.47*0.0021)) = 0o

Φ 60 = 0, Φ 90 = -22, Φ 120 = 0 ,Φ 150 =42 , Φ 180 = 0

38

Figure 3.10: Pressure Angle for Path 2 (Harmonic)

3.2.3.6 Calculation of inertial forces (F):

F= W*a/g

Where, W = Weight of the follower system = 200 gms.

Forces at the peak acceleration = F = 0.2*10*5.234/9.81= 1.061N.

3.2.3.7 Calculation of Max Parallel Force (P):

This force acts in a direction parallel to the Cam Axis

Here, L = External force is assumed to be 0

S = spring force = 0

µ = 0 (as the Shifter is very well lubricated)

Hence,

P = +-(W/g)a + W

Thus, Pmax = 1.06 + 0.2 = 1.26 N

39

Pmin = -1.06 + 0.2 = -0.86 N

3.2.3.8 Calculation of Max Normal Force (Pn):

This force acts in a direction normal to the cam curved surface

P n = P/tan φ

P n max = 1.261 / cos 0 = 1.261N

3.2.3.9 Calculation of Max Contact Stress Sc:

Contact stress was approximated based on the hertz contact stress theory as

follows.

Calculation of the contact stress is a tedious task, hence it is better to

approximately calculate it.

Assuming that the contact occurs between a cylinder (the follower) and the flat

surface of the cam, then Rc = ∞, and assuming that Ec = Ef = 206 Gpa =

21x104 N/mm2. We get

Max S2c = (0.35*1.261/4.05)/(8*2/21*104) = 37.88N/mm 2 = 5494.02 PSI

Including a Factor of Safety to account for the approximation

f = 5

Max Sc =27470 PSI or 190 N/mm2

The contact stresses should not exceed 1/3rd the ultimate compressive

strength of the material. The value of Ultimate compressive strength of steels

is about 800 to 1500 N/mm2. Hence, the design is safe under contact stress.

40

3.2.3.10 Tabulated Results of Calculations for Path 2:

Angle s

(mm)

v

(m/s)

a

(m/s2)

j

(m/s3) φ

Wa/g

(N)

P

(N)

Pn

(N)

Max Sc

(N/mm2)

0 0 0 0 0 0 0 0.2 0.2

190

30 0 0 0 0 0 0 0.2 0.2

60 0 0 -2.6 ∞ 0 -0.53 -0.33 -0.33

90 -2.5 -0.083 0 81.8 -22 0 0.2 0.21

120 -5.27 0 5.234 ∞ 0 1.061 1.261 1.26

1

150 0 0.175 0 -164 42 0 0 0

180 5.84 0 -5.23 ∞ 0 -1.601 -0.86 -1.60

Table 3.2: Tabulated Results of Calculations for Path 2

Similar Calculations were performed for the Cycloidal Profile which would also

be given for manufacturing.

41

Chapter 4: CAD Modelling and Prototyping

This chapter talks about the CAD Modelling Process of the Shifter Drum and

manufacturing of its prototype.

4.1 Design of the Prototype:

The prototype design was started by measuring the major dimensions of the

OEM shifter drum. Using these major dimensions and the design calculations

the CAD Model of the Shifter was constructed.

Figure 4.1: CAD Model of the Prototype

To see how the drum would fit in the engine an initial model of the shifter drum

was made out of Plaster of Paris by using Rapid Prototyping Machines. The

Drum was then tested for its dimensions and fit by actually putting it in the

engine .The drum was then rotated to check if the drum could actually move

the gears. These initial tests were a success and this enforced the team

members’ belief that this would actually work.

42

Figure 4.2: Rapid Prototype of the Shifter Drum

4.2 Final Model of the Shifter Drum:

One problem that came up initially during the design of the drum was how to

design a drum in such a way that it could be made hollow. Initially it was

decided that the drum would be cast , but then that idea was scraped due to

the complexity involved in the operation and cost . The cast drum would also

be heavier, so the decision of casting the drum was dropped. Machining was

the only option that seemed feasible, but in order to make the drum hollow it

would have to be made on two parts. Since machining was easy, cheap and

would result in a lighter drum, it was decided to design the drum in such a way

that it would be easy to machine.

The drum was supposed to be made in two parts one was the main cylinder

which had the cam paths and the second was the cap to be fitted over the

open end of the drum. The CAD Model of the Main Cylinder and the Cap are

shown below

43

Figure 4.3: Final Model of Main Cylinder

Figure 4.4: Final Model of Main Cylinder

44

Figure 4.5: Final Model of the Cap

Figure 4.6: Final Assembly

45

Chapter 5: Material Selection and Manufacturing

This chapter describes the procedure followed for material selection and steps

on manufacturing.

5.1 Material Selection:

The original shifter drum material details could not be obtained as it is a trade

secret, so material selection was one of the main problems faced in designing

the drum. The Drum actually sees almost no force during operation as seen in

the calculations; wear was the major criteria for which the drum material had to

be selected. The four main kinds of wear in cam-follower mechanisms are:

adhesive wear, abrasive wear, corrosive wear, and surface fatigue wear.

Adhesive and Abrasive wear resulting due to metal to metal contact, corrosive

wear resulting due to oxidation of the surface.

Hardness tests were conducted on the OEM Drum and its hardness was found

out to be about 46 Rc (Surface Hardness). After a lot of research the following

procedure for material selection was adopted.

The OEM Drum was first examined for the wear it had undergone. It was

observed that the drum had small grooves (smooth wear) suggesting that it

has undergone abrasive wear due to metal to metal contact ,it also had a dull

appearance which suggests that the material has undergone burnishing i.e.

non adhesive wear . It was concluded that the drum undergoes abrasive wear,

corrosive wear, cutting wear and low cycle fatigue.

After having determined what the drum undergoes, a steel with suitable

alloying elements was to be selected. To increase wear and abrasion

resistance, a steel with a high content of carbon and manganese was to be

selected. After a lot of consulting with the manufacturers and material experts,

Steel En41B was selected for manufacturing the drum. En 41B was selected

because it can be easily hardened and nitrided . Its composition is as shown

below

En41B is easily machinable and its availability is not a problem. The problem

of material selection was thus solved.

46

5.2 Manufacturing of the Shifter drum:

The raw material was to be machined into two parts and the following

procedure was to be adopted for manufacturing, the two parts i.e. the main

cylinder and cap were to be machined roughly in a 4- axis CNC. The parts

would then be hardened to 25 Rc hardness and in the last step the two parts

would be welded, then precision machined and then nitrided to a hardness of

about 48 Rc.

47

Chapter 6: Electronic Gearshift Mechanism

This Chapter presents the construction of the electronic gearshift mechanism

and the design of various mechanical components involved in it.

6.1 Requirements of the Shifter:

The Engine used i.e. The Honda CBR600 F4i with integrated transmission

uses a foot lever to shift gears, the foot Interface is unnecessary for the car.

This lever has about 120mm arm length and requires approximately 70

Newtons of force (depending on whether the engine is running or not), with

15mm of travel in either direction.

The system must be reliable and provide shifts for at least an entire endurance

race. In the endurance event the no. of shifts that a driver could perform is

about 880 shifts (considering endurance is 22 laps and there are 20 shifts per

lap and the factor of safety is 2).

The system should be simple, light weight, easy to use and should shift with

times less than 0.5 secs.

6.2 Pneumatic v/s Electric shifting systems:

Pneumatic shifters provide a great amount of force and this force remains

constant with the travel of the lever. Their response and time for actuation is

also less. The major disadvantages of these systems are that they require a

constant fluid source (CO2), i.e. they require a storage tank, and due to this

reason the weight of the system increases drastically. They are also a lot more

complex.

On the other hand electric systems are light weight and provide just enough

force to perform the shifts. The major problem with the electric system is that it

puts an additional load on the already overloaded electrical system.

After of a lot of thought process it was decided to go ahead with the electric

system and if a battery problem arises, an individual light weight battery would

be used to drive the shifter.

48

6.3 Previous year’s Electronic shifting mechanism:

The Electronic shifting mechanism used in the previous year’s car used a

solenoid as the actuator. A single solenoid is used for both pushing and pulling

the gear lever. The solenoid could not provide enough force to shift the gears

and hence would sometimes shift successfully but failed the rest of the times.

This system was also very heavy at 2.5kgs .Hence; it was decided to replace

the solenoid with a light weight and reliable actuator.

Figure 6.1: ORI2009’s Electronic shift mechanism

6.4 Initial Concept and Testing:

As a replacement for the solenoid, a no. of solenoids and motors were tried

and finally a compressor motor rated at 12V weighing 1.5kgs was selected.

The initial tests conducted on the previous year’s car with the motor yielded

good results and hence it was decided to go ahead with the initial design and

refine it.

49

Figure 6.2: The Motor used as Actuator

Figure 6.3: Initial Testing

6.5 The Final Product

It was decided that relays be used for driving the motor, which would be

triggered by buttons placed on the steering wheel. It was also decided to use

50

the Gear Change Ignition Cut Feature of the newly procured MoTeC M400

ECU which senses when a gear change occurs and shuts the engine for a few

milliseconds by cutting power to the coils, so that upshifts can be made liftless

(Since when the engine is shutdown the load on the driveline reduces).To

signal the ECU that a gear change has occurred a SPDT Switch was used. A

new motor lever was also designed as shown in the figure below.

Figure 6.4: Motor with the Motor Lever

6.6 Testing

As stated earlier that a part which has not undergone rigorous testing will not

be put on the car. The finished shifter was tested for in acceleration and gave

promising times. The car underwent a no. of testing sessions and luckily the

shifter did a commendable job in almost all of them. The one time that it failed

was due to low battery voltage, this problem will be further looked into and a

viable solution will be found out and applied. The shift times were also

approximately determined, which were about 500 to 600ms which are

acceptable.

51

Chapter 7: Future Work and Conclusion

7.1 Future work:

Future developments on the gearbox include changing of the current ratios to

optimal ratios so as to improve lap times and reduce the final drive ratio.

Further weight reduction of the gearbox could be carried out by changing the

dog gears with the lighter weight dog gears.

7.2 Conclusion:

The designed systems i.e. the Shifter Drum and the Electronic gearshift

system were tested successfully and will be implemented on the ORI 2010

Car. So far the systems have shown promising results. The Shifter Drum

(458gms) manufactured is lighter than the OEM drum (~600gms). The

Electronic shifter at 1700gms is lighter than previous year’s shifter at 2500gms

and shifts in about 500ms in upshifts with comparable time in downshifts.

This document presents the designs and the decisions that were made during

the development of the systems, so as to help future team members to

understand why certain decisions were made, so that they can find the areas

where improvement can be made. One Important lesson learnt during the

design and fabrication of the above systems was that even the most planned

out approaches can fail due to difficulty with the supplier , fabrication etc. . To

avoid such hiccups a very well planned, tried and tested approach should be

followed so that everything is completed on time.

52

References:

Books:

The Motor Vehicle – T.K.Garett

Automotive Transmissions: Fundamentals, Selection, Design and Application

by Gisbert Lechner and Harald Naunheimer

Cam Design by Clyde H. Moon

Cam Design Handbook by Harold A. Rothbart

Cam Design and Manufacturing Handbook by Robert L.Norton

Mechanical Engineer’s Data Handbook by James Carvill

Handbook of Materials Selection by Myer Kutz

Reports:

Methodical design of a selector for automotive semi-automated gearbox by

Francesco Cappellt, Luca Ciulla, Antonio Mancuso

Simplification of the Shift/Clutch Operations for Formula SAE Vehicles by

Hiroshi Enomoto , Hironari Morita, Yousuke Fukunaga and naoki UOTA

Estimation procedure of increasing the speed of gear shift in sports cars at the

design stage by Dudnikov, Andrey

Honda CBR600 F4i 2001 Service Manual

Small Engine Performce Limits – Turbocharging, Combustion or Design by

William Attard

Design of a Pneumatically Assisted Shifting System for Formula SAE Racing

applications by Andrew J. Kennett

53

World Wide Web:

Gearbox Basics -

http://en.wikipedia.org/wiki/Transmission_(mechanics)

Cam Mechanism Basics -

http://www.cs.cmu.edu/~rapidproto/mechanisms/chpt6.html

Motorcycle Gearshifting Patterns -

http://www.timberwoof.com/motorcycle/faq/riding.html

Motorcycle Gearboxes and Selector Mechanisms -

http://www.dansmc.com/

Material Data -

http://www.westyorkssteel.com/

Material Data -

www.matweb.com

54

Appendix A

Production Drawings

55

56