Design and Construction of a Fuel Efficient Drive Train System of a

Design and Construction of a Fuel Efficient
Drive Train System of a Prototype Vehicle
for the Shell Ecomarathon 2011
A Thesis presented to the
School of Mechanical Engineering
Mapua Institute of Technology
In Partial Fulfillment
of the Requirements for the Degree of
Bachelor of Science in Mechanical Engineering
By:
Dujunco, Karlo Antonio D.
Mendoza, Johndelon P.
Radovan, Cyril R.
October, 2011
i
APPROVAL SHEET
This thesis,
Design and Construction of a Fuel Efficient Drive Train System
of a Prototype Vehicle for the Shell Eco-marathon
by
Dujunco, Karlo Antonio D., Mendoza, Johndelon P.
& Radovan, Cyril R.
has been approved for oral defense
Engr. Hans Bosshard
Thesis Coordinator
Sherwin S. Magon
Thesis Adviser
Thesis Evaluation Committee
Engr. Igmedio F. Isla Jr.
Panel Member
Engr. Jaime P. Honra
Panel Member
Engr. Godofredo C. Salazar
Panel Member
Accepted and approved in partial fulfillment of the requirements for the Degree of
Bachelor of Science in Mechanical Engineering.
Dr. Manuel Belino
Dean
School of Mechanical Engineering
ii
ABSTRACT
The Shell Eco-Marathon is a competition wherein teams from different universities
design and fabricate vehicles with the goal of going the furthest with the least amount of
fuel consumed. Proper design and fabrication methods must be employed in order to
increase the efficiency of the vehicle and use less energy. In this study the group designed
and fabricated a prototype vehicle that uses a gasoline internal combustion engine.
The group employed the use of a four stroke Honda GX35 grass cutter engine. The
engine is tested using a hydraulic engine test bed and a dynamometer in order to gather
information regarding the engine’s performance, i.e. its torque, RPM, and power. A chain
drive system is used to increase the gear ratio, resulting in a final ratio of 13.5.
Under the given factors and conditions the team worked with, the vehicle produced,
named Amihan, was able to achieve a fuel consumption versus mileage of 186 km/L at
its best run.
iii
ACKNOWLEDGMENT
This thesis would not be possible without the help of several individuals who shared their
expertise and knowledge.
The authors want to express their gratitude to their adviser, Professor Sherwin Magon for
the help, support, and guidance throughout the project and for the technical insights
provided by Engr. John Judilla
The authors want to express their deepest gratitude to Dr. Manuel C. Belino and Engr.
Igmedio Isla, Jr. for their unrelenting support and boost of morale for the team; and to
MIT’s President, Dr. Reynaldo B. Vea, for believing in the team.
The authors want to thank Jaylord Jauod and Eliseo Capili for helping in the construction
of the vehicle. The help of the ME laboratory assistants during the course of the thesis is
also very much appreciated.
The group would also want to give their gratitude to Pilipinas Shell Petroleum
Corporation for giving the opportunity to be part of the competition that molded their
character in every aspect of life.
Most importantly, none of this is possible without God. To God be the GLORY.
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Table of Contents
Title Page
Approval Sheet
Abstract
Acknowledgement
Table of Contents
i
ii
iii
iv
v
Chapter
1
Introduction
1.1 Overview
1.2 Statement of the Problem
1.3 Objectives of the Study
1.4 Significance of the Study
1.5 Scope and Limitations
2
Review of Related Literature
2.1 Review of Related Studies
2.1.1 Modeling of a Vehicle with Continuously
Variable Transmission
2.1.2 Team Atalanta Technical Innovations
2.1.3 Rose-Hulman Institue of Technology
SEM
2.1.4 Mixture Distribution in a Multi-Valve
Twin-Spark Ignition Engine
2.1.5 Dalhousie University Supermileage
Team Final Report
2.2 Review of Related Literature
2.2.1 Internal Combustion Engine
2.2.1.1 Four Stroke Engine Cycle
2.2.2 Transmission System
2.2.2.1 Roller-Chain Drive System
2.2.2.1.1 Chains
2.2.2.1.2 Chain Selection
2.2.2.1.3 Chain Lubrication
2.2.3 Clutches
2.2.3.1 Centrifugal Clutch
3
Theoretical Considerations
3.1 Shell Eco-marathon Rules
v
1
2
2
3
4
6
7
8
9
10
11
12
12
13
14
15
15
16
16
18
3.2 Vehicle Dynamics
3.3 Torque and Power
3.4 Roller Chain Drive Design
3.4.1 Drive Ratio
3.4.2 Drive Arrangements
3.4.3 Shafts Center Distance
4
5
6
Methodology
4.1 Development Process
4.2 Engine Selection and Initial Testing
4.2.1 Engine Selection
4.2.2 Initial Testing
4.3 Initial Design and Computation
4.3.1 Design of Drive Train System
4.3.2 Computation of Speed Reduction Ratio
4.4 Construction and Installation
4.4.1 Chains and Sprockets
4.4.2 Clutch System
4.5 Reliability and Efficiency Testing of the System
4.6 Dynamometer Testing
4.7 Additional Tests
Discussion of Results
5.1 Test Runs
5.1.1 Mapua Test Runs
5.1.2 Malayan Test Runs
5.1.3 Batangas Racing Circuit Test Runs
5.2 Sepang International Circuit Run
5.3 Dynamometer Test
5.4 Additional Tests and Findings
5.4.1 Clutch Temperature Gradient Test
5.4.2 Clutch Power Loss
Conclusion and Recommendations
6.1 Conclusion
6.2 Recommendations
vi
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19
21
21
22
23
25
25
26
27
28
31
34
36
37
38
41
41
42
42
43
44
46
48
49
References
51
Appendices
A
B
C
D
E
F
G
Proposed Timetable (Gantt Chart)
Chain Selection Graph
Sprocket Selection Table
Computations
Clutch Slip: Temperature Gradient
Clutch Slip: Power Loss
List of Figures, Equations and Tables
vii
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54
55
56
59
60
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CHAPTER 1
INTRIDUCTION
1.1.
Overview
The Shell Eco-marathon began in 1939 as a friendly wager between the scientists at
Shell research laboratory in the United States to see who could get the most mileage
per gallon of fuel from their vehicle.
In its current form, Shell Eco-marathon is an event that challenges students to design,
build and drive the most energy efficient vehicle. It is not a competition of how fast
the vehicle goes; but rather how far the vehicle would go using the least amount of
fuel. (Shell Oil Company 2010)
For the second Shell Eco-Marathon Asia, Mapúa participated again in this prestigious
competition which was held in Sepang Racing Circuit in Kuala Lumpur, Malaysia on
July 2011.
The group decided to enter the competition under the category of designing and
fabricating a prototype vehicle where the primary consideration is focused on
reducing frictional constraints and maximizing performance efficiency. Specifically,
this study focused on the engine, drive train, instrumentation, and innovations to
improve the fuel efficiency of the vehicle.
1
1.2.
Statement of the Problem
This study aims to answer the problem on how to improve fuel efficiency by using
different innovations are based on research and testing. The study tackled the
challenge of developing designs based on concepts of the internal combustion engine,
machine elements, mechanics of materials, and many other disciplines and came up
with a prototype vehicle that is fuel efficient and lightweight.
Also, along with a fuel efficient vehicle, the study may find a solution on decreasing
pollution regarding the carbon emissions as the byproduct of combustion. Such
outcomes may bring about environment friendliness, health consciousness, and life
preservation.
The team’s major challenge was to build a prototype vehicle along a strict compliance
with the Shell Eco-Marathon Asia 2011 Rules and Regulation. Design and innovation
is a vital introduction in the construction of an effective and efficient Power Train
System.
1.3.
Objectives of the Study
The study aimed to satisfy the following general and specific objectives as presented
subsequently:
2
General Objectives:
The study aimed to design and construct a power train system for the prototype
vehicle to maximize engine performance and improve fuel efficiency.
Specific Objectives:

To construct a prototype vehicle driven by a four-stroke single cylinder
gasoline engine.

To design and construct a drive train system: Roller Chain Driven System.

To determine the chain and sprocket dimensions through power and load
calculations and other parameters based on the official rules of the said event.

To determine fuel consumption of the designed power train system in
kilometer per liter.
1.4.
Significance of the Study
According to Takahashi (1998), the concern of the environment has recently led to
regulations on fuel consumptions and exhaust emissions. Fuel consumption is
particularly important as it represents the technology of automobile industry.
Different innovations were presented by many automotive manufacturers that
promise high engine efficiency ratings and significant fuel economy. This study can
give many opportunities including development of new designs and innovations for
urban road use. Smite is a prototype of an ultra-fuel efficient car for urban road use
3
presented at the Shell Eco-marathon 2010. Sohrab Kazamahvazi, from Vehiconomics
Company quotes that: “Many cars on today’s city roads are totally unsuitable to the
urban environment. On the other hand some of the Prototype cars that the Shell Ecomarathon has showcased would be great for urban use, but of course they are not
roadworthy. With the SMITE we are filling that gap and reshaping urban mobility.”
Concepts from this study can help promote new ideas that could end up on the roads.
This study may also serve as a future reference for individuals are be interested in
joining competitions or anyone who would be engaged in the same study. And lastly,
this may also be an innovation for designing efficient and effective transmission
system.
1.5.
Scope and Limitations
The study covered the design and fabrication of the drive train system for the
prototype vehicle to maximize fuel efficiency. The vehicle was driven by a fourstroke, single cylinder gasoline engine with a chain driven transmission system.
Adapting the modification of the Team Atalanta (Mapua Institute of Technology,
2010) and the Supermileage Teams (Dalhousie University, 2009), a modified velocity
stack was also integrated in the air intake. A velocity stack with a curved inlet was
used to allow a smooth flow of air in the intake side of the engine.
4
The drive train was designed and constructed according to the results that was
obtained after the power and load computations. Basically, the drive train was chain
driven and was able to vary in torque output.
For the vehicle design, dimensions, and safety appendages, the study was under the
strict compliance of the Shell Ecomarathon Asia 2011 official rules.
5
CHAPTER 2
REVIEW OF RELATED LITERATURE
2.1.
Review of Related Studies
2.1.1. Modeling of a vehicle with Continuously Variable Transmission
The team of Dragos, S. Preitl did a study on modeling of a vehicle with
continuously variable transmission wherein it was concerned about the modeling
and driving scenarios of a class of vehicular power train systems. The modeling
included an analysis of the system components which included the spark-ignition
internal combustion engine, the torque converter, the metal v-belt type
continuously variable transmission and the wheels. Moreover, the model
produced was subjected to four simulation tests: acceleration, overtaking
maneuver, ramp encounter, and the brake engine. With the results and illustrations
obtained by the team, they concluded that the model produced a good
performance when the vehicle was subjected to the four driving scenarios. Thus,
the study provided a new mathematical model for a vehicle power train system
with spark ignition internal combustion engine and a continuously variable
transmission. (Variable Transmission IEEE Xplore Digital Library)
The study provided some concept on how shifting speed and torque variation can
affect the performance of the engine. The study gave no direct bearing on the fuel
efficiency of the vehicle but stated the engines performance through mathematical
modeling.
6
2.1.2. Team Atalanta Technical Innovations
Based on the study made by the Shell Eco marathon team, Atalanta, the following
technical innovations for the prototype vehicle were conducted:
The rollers and guides were made of HDPE (engineering plastic) and was
machined function as a roller for the chains to provide low friction and rigidity for
the moving chains also. All chained components were fitted with roller chain
guides to maintain constant tension and keep the chains on track.
Figure 1: Agimat Transmission System
The transmission system used was single speed and had two stages of speed
reduction. First, speed reduction was from the engine to the clutch and the second
was from the clutch to the rear wheel. The overall speed reduction was 14. The
rear wheel uses a Shimano bicycle hub that also free wheels when there is no
torque applied which was used to make use of the car’s momentum when the
driver releases the gas lever.
7
The Team Atalanta fabricated a velocity stack to increase the power of the engine
for a specific range of engine rpm. The velocity stack was fitted in the engine’s air
intake and allows smooth entry of air to the intake tract. The velocity stack also
increased the volume of air intake.
The study of Team Atalanta on their prototype vehicle was very useful for the
group. Some concept of their study were integrated in the group’s study and tested
further.
2.1.3. Rose- Hulman Institute of Technology SEM
The team of Rose-Hulman Efficient Vehicle from Rose-Hulman Institue of
Technology in the United States of America provided a study of their technical
innovations used for their Shell Eco marathon car. The team used a TMAP
(temperature-manifold air pressure) sensor to measure how much the air enters
the intake. After knowing the time to keep the fuel injector open, the appropriate
spark and fuel was set to advance based on the engine rpm, which was read by
two sensors located on the cam and crank. These signals output the controller and
activated the injector and spark coil as instructed. After a few cycles, the
controller looked at the exhaust oxygen sensor which served as the correction in
the control loop.
Integration of electronic sensors in the engine was being considered; however,
budget and time constrictions were also taken in account. The team’s principle
8
was to keep things simple but effective; adding other variables in the process may
cause delay.
2.1.4. Mixture Distribution in a Multi-Valve Twin-Spark Ignition Engine
According to the study made by Mitroglou, et al., the injector timing was found to
control spray impingement on the piston and cylinder wall, thus contributing to
quick and efficient fuel evaporation. It was confirmed that in-cylinder charge
motion plays a major role in engine’s stable operation by assisting in the
transportation of the air-fuel mixture towards the ignition locations (i.e. spark
plugs) in the way of uniformly distributed charge or by preserving stratification of
the charge depending on operation mode of the engine. (Spark Ignition Engine
IEEE Xplore Digital Library)
The sequential ignition system is a promising concept since it may provide more
combustion efficiency depending on the setting and timing of the system. The
concept was still being discussed by the group and its use depended on the
resources that are available or will be available.
9
2.1.5. Dalhousie University Supermileage Team Final Report
Figure 2: Dalhousie Supermileage Team Drive train system
Based on the study made by the Supermileage team from Dalhousie University,
the following specifications were observed:
For the intake and exhaust, pulse waves travel down the length of the manifold
until they reach the end. At the end of the manifold, the waves create a reflected
pressure wave back into the manifold. In the case of an intake manifold, one looks
for the waves to reinforce at the intake valve and create a higher-pressure area to
force more air into the combustion chamber. This gives the engine a greater
volumetric efficiency.
The team decided that a disk clutch would be more efficient than a centrifugal
clutch because of the chosen driving strategy which involves a “burn and coast”
method where the driver periodically starts the engine, and accelerates it to an
optimum top-end speed, shuts the engine off and coasts it until the bottom-end
speed is reached. Once the speed is reached, the driver will start the engine and
10
repeat the process. With this driving strategy, the driver can keep the disk clutch
engaged at all times because of the free wheel at the rear. On the other hand, in a
centrifugal clutch, the clutch will only engage when the engine shaft is rotating
above a specific RPM which would lead to inefficiencies during engine start-up
and when the driver lets off the throttle during cornering.
This study gives information on how the prototype was fabricated their prototype
and gives further concept and strategy of how to run the vehicle to achieve better
performance. Their modification may be used; however, further testing and
alteration should also be done.
2.2.
Review of Related Literature
2.2.1. Internal Combustion Engine
The purpose of an internal combustion engine is to convert chemical energy
contained in the fuel into useful mechanical energy. In internal combustion
engines, as distinct from external combustion engines, energy is released by
compressing and then burning or oxidizing the fuel inside the cylinder of the
engine. The work produced after combustion transfers and provides the desired
power output directly between the mechanical components and working fluids of
the engine. The air-fuel mixture before combustion and burned products after
combustion are considered as the actual working fluids. The internal combustion
engine which is the subject of this study is a spark-ignition engine (sometimes
called Otto engines, or gasoline or petrol engines). (Heywood 1988)
11
2.2.1.1.
Four-stroke Engine Cycle
In the four stroke cycle engine, the five parts of combustion occurs every
four strokes (two upstrokes and two downstrokes) of the piston. Two
revolutions of the crankshaft are required to accomplish the five parts of
combustion so that every other time the piston approaches the top of the
upstroke, combustion and power occur. The four strokes are: Power stroke,
Exhaust stroke, Intake stroke, compression stroke. (Schuster 1999)
2.2.2. Transmission System
Mechanical power transmission system, also known as drive train, is a device that
is composed of linkages, shafts, bearings, gears, pulleys, or other components that
transmits and controls the force and motion from one device to another. The
device it receives power from is called the driver, or prime mover, and the device
it transmits power to is called the driven device. (Amatrol 2000)
The device for transmitting power from the engine to the drive shaft is the
transmission system. This power is eventually transmitted from the shaft to the
wheels. The device must convert the power from the relatively high angular
velocity and low torque to lower velocity and higher torque needed at the wheels
(Lexicon Universal Encyclopedia 1993).
12
2.2.2.1.
Roller Chain Drive System
A chain drive consists of an endless chain of links meshing with the driving
and driven sprocket. This type of drive gives a positive speed ratio between
driving and driven sprockets, so that tension on the slack side is
unnecessary, because the links of the chain engage with the sprocket teeth
and drive them, chain drives can operate with the small arcs of contact and
short center distances.
Power chain must articulate over the sprockets and this movement results in
wear in the chain. As a result, the chain elongates with use, chain elongation
can be continued until the chain pitch no longer is compatible with the
sprocket pitch. The sprockets are designed to accept 3 to 6 percent chain
elongation. This elongation will be influenced by chain tension and by
sprocket size. Larger sprockets reduce wear because the chain flexes
through a smaller angle on a large sprocket.
Basic features of chain drive include a constant ratio, since no slippage or
creep is involved. Long life and the ability to drive a number of shafts from
a single source of power is also a feature of the system. (Patton 1980)
13
2.2.2.1.1. Chains
A chain is a reliable machine component, which transmits power by
means of tensile forces, and is used primarily for power transmission
and conveyance systems. The function and use of chains are similar
to a belt. There are many kinds of chain. It is convenient to sort types
of chain by either material of composition or method of construction.
A standard roller chain consists of alternating roller links and pin
links. The roller link consists of two rollers with two bushings, the
bushings being press fitted into the link plates (side members). The
bushings cannot turn in the link plates as the chain articulates. A
standard roller chain has three principal dimensions: pitch, chain
width and roller diameter. Chain width is the minimum distance
between the link plates, and the pitch is the distance between the
centers of adjacent bushings (Patton 1990)
Roller chains have been standardized as to sizes by ANSI. The pitch
is the linear distance between the centers of the rollers. The width is
the space between the inner link plates. These chains are
manufactured in single, double, triple and quadruple strands.
(Shigley, Mischke 2001)
14
2.2.2.1.2. Chain Selection
The allowable value of the force observed in the chain is a function of
many variables, including its type and size of chain, the type of driven
load, type of power source, type and degree of lubrication present.
The four commonly recognized modes of failure of roller chain are: 1.
Wear, 2.Fatigue failure of the link plates, 3.Fatigue failure of roller or
bushings, and 4.Failure in pins and bushings. (Phelan 1975)
2.2.2.1.3. Chain Lubrication
Roller chain consists of series of connecting travelling metallic
bearings, which must be properly lubricated to obtain the maximum
service life of the chain. Although many slow speed drives operate
successfully with little or no lubrication beyond the initial factory
lubrication, proper lubrication will greatly extend the useful life of
every chain drive. The chain drive requires lubrication for six
purposes:

To resist wear of the pin-bushing joint.

To cushion the impact loads

To dissipate any heat generated

To flush away foreign materials

To lubricate chain-sprocket contact surfaces
15

To retard rust or corrosion
There are two basic types of lubrication for roller chain drives. Close
adherence to the recommended type of lubrication is essential to
obtaining maximum service life of a chain drive.
1. Manual or Drip Lubrication
Oil should be applied periodically between the chain link plate
edges with a brush, spout can, or drip lubrication.
2. Oil Bath or Oil Slinger
With bath lubrication, the lower strand of chain runs through a
sump of oil in the drive housing. The oil level should reach
the pitch line of the chain at its lowest point while operating.
Only a short length of chain should run through the oil.
(maintenanceresources.com 2007)
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2.2.3. Clutches
All standard or manual transmissions have a clutch to engage or disengage the
transmission. There are times when the driving wheels should not rotate while the
engine is running. The clutch is used to engage or disengage the transmission and
driving wheels (Schwaller 1999).
In a car, you need a clutch because the engine spins all the time, but the car’s
wheels do not. In order for a car to stop without killing the engine, the car’s
wheels need to be disconnected from the engine somehow. The clutch allows us
to smoothly engage a spinning engine to a non-spinning transmission by
controlling the slippage between them (Bryant, Nice 2011)
2.2.3.1.
Centrifugal Clutch
A typical centrifugal clutch has a set of shoes that are forced out against the
drum by centrifugal force. The shoes may be loosely held within the drum,
but in more refined designs the shoes are connected to the input member by
means of a floating link. Other advantages: these clutches are good starters
for high-inertia loads. They tolerate a considerable amount of manufacturing
variations and are well suited to drives that undergo vibrations and heavy
shock loads. (Parmley 2000)
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CHAPTER 3
THEORETICAL CONSIDERATIONS
3.1.
Shell Eco-marathon Rules
According to the Shell Eco-marathon Rules 2011, the vehicle should follow these
specifications:
Article 26: Dimensions
1. The maximum height must be less than 100cm.
2. The maximum height measured at the top of the driver’s compartment must be
less than 1.25 times the maximum track width between the two outermost
wheels.
3. The track width must be at least 50 cm., measured between the midpoints
where the tires touch the ground.
4. The wheelbase must be at least 100cm.
5. The maximum total vehicle width must not exceed 130cm.
6. The maximum total length must not exceed 350cm.
7. The maximum vehicle weight, without the driver, is 140kg.
Article 36: Clutch and Transmission
1. Vehicles with internal combustion engines must be equipped with a clutch
system, so that during inspection and fuelling operations the vehicle remains
stationary with the engine running.
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a. All clutch systems used prior to 2011 must be reviewed with respect to
whether the clutch system remains not engaged at engine start up.
2. The starter motor speed must always be below the engaged speed of the
clutch.
3. The installation of an effective transmission chain or belt guard(s) is
mandatory.
3.2.
Vehicle Dynamics
The drive ratio or the speed ratio of the vehicle will be obtained through the analysis
of the vehicle dynamics. Parameters such as aerodynamic design, forces acting on the
tires, and weight distribution are factors in determining the drive ratio of the vehicle.
(Genta, 1997)
3.3.
Torque and Power
Mathematically, the torque on a particle (which has the position r in some reference
frame) can be defined as the cross product:
𝑇𝑇 = 𝐹𝐹 × 𝑟𝑟
eq. 1
19
Where:
𝑟𝑟 = The particle’s position vector
𝐹𝐹 = The force acting on the particle
𝑇𝑇 = The torque
Torque has dimensions of force multiplied by the distance and the SI units of torque
are stated in “Newton meters”. (Serway et al. 2008)
Power is defined as the rate or speed at which work is done. Engine power is the
power available from the crankshaft to do work (Crouse et al. 2004, 180).
Power can be calculated using the formula:
𝑃𝑃 = 2𝜋𝜋𝜋𝜋𝜋𝜋
Where:
eq. 2
𝑃𝑃 = Power
𝑁𝑁 = Revolutions per minute
𝑇𝑇 = Torque
The size of chain for the application was made by referring to selection charts relating
shaft speed and design power. The design power, Pd, is determined from the indicated
power and application factors such as service factors, sprocket size factor, and
temperature factor.
20
𝑃𝑃𝑑𝑑 = 𝑃𝑃𝑓𝑓1 𝑓𝑓2 𝑓𝑓3
Where:
eq. 3
𝑓𝑓1 = Service factor
𝑓𝑓2 = Sprocket size factor
𝑓𝑓3 = Temperature factor
𝑃𝑃
3.4.
= Indicated power
Roller Chain Drive System
Roller chain is designed for transmitting high torque loads, and provides the ideal
system to connect shafts. Selection and application is done by following normal
engineering practices.
3.4.1.
Drive Ratio
Roller chains operate at high efficiency on drive ratios of 3:1 up to 5:1. Higher
ratios are not recommended for chain drives with two sprockets. However, twostage drive system can be applied if higher reduction ratio is required.
The gear ratio of a gear train is the ratio of the angular velocity of the input gear
to the angular velocity of the output gear, also known as the speed ratio of the
21
gear train. The gear ratio can be computed using the number of teeth of the
various gears that engage to form the gear train. (Pennock, Shigley 2003)
Where:
𝜔𝜔1 × 𝑁𝑁1 = 𝜔𝜔2 × 𝑁𝑁2
eq. 4
𝜔𝜔 is the angular velocity
𝑁𝑁 is the number of teeth
3.4.2.
Drive Arrangements
It is preferred to use Roller Chain on drives with horizontal shafting, although
vertical shaft drives can be accommodated. Shaft centers may be displaced
horizontal at an incline, or vertical, with each arrangement having its own specific
requirement. Horizontally displaced shafts, and drives with centers inclination up
to 60°, are the best and most common arrangements. On inclined drives the driver
can be either above or below the driven sprocket, but it is preferable to have the
driving strand of the chain uppermost.
22
Figure 3: Example of a Drive Arrangement
3.4.3.
Shaft Center Distances
For optimum chain life shaft centres within the range 30 to 50 times chain pitch
should be used. The center distance for chain drives may of course be relatively
short, but a minimum of wrap of 120° is desirable; this condition is inevitably met
when mw < 3. An average good center distance would be D2 + D1/2, where D2 is
the pitch diameter of the larger sprocket, D1 of the smaller. The approximate
length of the chain is:
𝐿𝐿 = 2𝑐𝑐 +
𝑁𝑁1 + 𝑁𝑁2 (𝑁𝑁1 + 𝑁𝑁2 )2
+
2
40𝐶𝐶
eq. 5
Where C is in pitches (may contain a fraction). The length should be an even
number of pitches to avoid an offset link. There is the usual matter of adjusting
the chain length, center distance, and sprocket sizes so that everything fits.
23
Contrary to belting practice, the slack side of the chain is preferred on the bottom
of horizontal drives, especially for long centers, because, if the slack strand is on
top, the strands may touch after the chain has lengthened in service. Lengthening
occurs because of wear in the joints. (Faires 1969)
24
CHAPTER 4
METHODOLOGY
4.1.
Development Process
The following steps were followed in this study :
1. Engine selection and initial testing
2. Computation and initial design of the system
3. Construction and Installation of the designed system
4. Testing of the designed system
5. Overall testing and determination of fuel consumption
4.2.
Engine Selection and Initial Testing
4.2.1. Engine Selection
The engine used for the system was selected from a number of small engines
available in the market. Any engine can be used; however, since efficiency is the
focus, it was decided to use the Honda GX35. This engine was also used in
several studies, including Team Atalanta of Mapua Institute of Technology, which
in turn gave promising results. The engine is compact, light weight, and has
enough torque to drive over 100 kg of load.
25
4.2.2. Initial Engine Test
An engine analyzer was used to determine basic parameters of the engine such as;
fuel consumption, torque developed, and power output at any given speed. The
data gathered in this test was compared to the manufacturer’s specification.
Further information about how the engine behaves or perform will be a factor on
designing the system.
7000
1.4
6000
1.2
5000
1
4000
0.8
Speed
3000
0.6
Torque
2000
0.4
1000
0.2
0
0
0
Speed Trend
Torque Trend
50 100 150 200 250 300 350 400 450 500 550 600
Time, s
Figure 4: Initial Engine Test Results
The graph shows the torque developed and the speed of the engine at a given test
time. Results show that the GX35 has a maximum torque of 1.2 N-m, less than the
specification of the manufacturer. However, it was developed at a lower speed of
5000 rpm.
From the test conducted, the fuel consumption at this speed was 4.9 mL per
minute. Knowing this, prediction of how much fuel can be consumed is possible
but it is not recommended due to different factors still affecting the engines
performance.
26
4.3.
Initial Design and Computation
4.3.1. Design of the Drive Train System
The drive train system was designed based on the availability of materials and
components. The group initially planned a direct transmission system to lessen
frictional losses. However, sprockets with teeth number greater than 60 have
diameters same as the wheel.
Studying the designs of different teams, led us to a conclusion that the most
practical solution is to add a secondary system. The design would then be a twostage speed reduction system as shown below.
Figure 5: Initial Design of Drive Train System
This system lessens the length of the chain needed to connect the engine to wheel.
Decreasing the center to center distance of each sprocket will also lessen slacking
of chain. This system also decreases the probability for the chain to be derailed.
27
4.3.2. Computation of Drive Ratio
The required speed reduction was computed based on the needed torque for the
vehicle. Parameters such as vehicle weight, and road inclination were accounted
for the computation of the torque required for the vehicle to move. The Sepang
Circuit was known to have a maximum road inclination of 3 degrees. Since all
vehicles require more torque during uphill condition, the group decided to
compute the required torque within this condition and thus the needed speed
reduction.
Furthermore, the weight of the vehicle and the load it will carry were also taken
into account. The free body diagram (figure 7) was used to calculate the torque
required. The weight of the vehicle, W, was located at the center of gravity of the
vehicle. The CG of the vehicle was kept near as possible to the ground. This
would increase vehicle stability especially during cornering.
Figure 6: Free Body Diagram of the Vehicle
28
The weight of the vehicle is distributed among the wheels of the car, Wr and Wf.
The torque required for the vehicle was solved based on the tractive force that the
wheel generates to the ground. The weight acting the rear wheel Wr can be solved
by summing up moments at the front wheel.
𝐿𝐿𝑊𝑊𝑟𝑟 = ℎ𝑊𝑊𝑊𝑊𝑊𝑊𝑊𝑊𝑊𝑊 + 𝑙𝑙𝑙𝑙𝑙𝑙𝑙𝑙𝑙𝑙𝑙𝑙
eq. 6
The target weight of the vehicle was along 30 to 50 kg based on initial simulation
of chassis and weight of vehicle components. Additional weight of the driver of
50 to 60 kg was also recognized in the computation. At this stage, the actual
center of gravity of the vehicle was not determined since other components and
the vehicle body (vehicle shell) was not available just yet. So the group computed
the force as the center of gravity evenly distributed along the rear and front axles.
The height of the CG from the ground was placed on the hip level of the driver in
driving position, taking account the location of the CG of the human body.
𝐹𝐹𝑡𝑡 = 𝐹𝐹𝑓𝑓 + 𝑊𝑊 sin 𝜃𝜃
eq. 7
The previous equation was the summation of forces acting on the wheel. Ft
presents the tractive force of the wheel needed to move the vehicle on an inclined
plane and Ff presents the rolling friction acting on the wheel.
The coefficient of rolling friction was obtained based on standard production line
of bicycles tires. Coefficient of rolling friction ranges from 0.002 to 0.005 on very
smooth surfaces for bicycles tires. However, the manufacturer claimed that these
values tend to increase 100% or higher on road surfaces. On the other hand,
29
production car tires on road surfaces ranges from 0.015 to 0.02. The group then
decided to use a coefficient of rolling friction of 0.02 for car tires on road surfaces
since conditions are similar for the bicycle tires.
𝐹𝐹𝑡𝑡 = 𝐶𝐶𝑟𝑟𝑟𝑟 𝑊𝑊𝑟𝑟 + 𝑊𝑊 sin 𝜃𝜃
eq. 8
Crf presents the coefficient of rolling friction and the normal force was taken as
the weight acting on the rear wheel of the vehicle. With this the tractive force of
the wheel can now be determined.
𝑇𝑇𝑤𝑤ℎ𝑒𝑒𝑒𝑒𝑒𝑒 = 𝐹𝐹𝑡𝑡 𝑟𝑟
eq. 9
The torque is now then computed multiplying the tractive force to the wheel
radius. In determining the torque on the wheel, the speed reduction from engine to
wheel is now possible to determine.
𝑇𝑇𝑤𝑤ℎ𝑒𝑒𝑒𝑒𝑒𝑒 × 𝑁𝑁𝑤𝑤ℎ𝑒𝑒𝑒𝑒𝑒𝑒 = 𝑇𝑇𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒 × 𝑁𝑁𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒
eq. 10
The torque of the engine was obtained as the best efficiency point; 1.2 N-m at
5500 rpm. It was assumed that the engine will run at 5500 rpm to obtain the 1.2N-m torque to keep the engine running efficiently. The speed reduction ratio can
now be computed as follows.
𝑆𝑆𝑆𝑆𝑆𝑆𝑆𝑆𝑆𝑆 𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟 𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟 =
𝑁𝑁𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒
𝑁𝑁𝑤𝑤ℎ𝑒𝑒𝑒𝑒𝑒𝑒
eq. 11
The computed reduction ratio at this stage was 13.0 for 100 kg weight of the
vehicle with the driver. Aerodynamic parameters such as drag force and lift may
also be included in the computation. However, this was not included due to lack
of data at this stage.
30
4.4.
Construction and Installation
The materials in construction of the system should be, as much as possible, available
in the local market. The materials should also be easy to procure and should require
minimal adjustments or fabrication for the designed system. Quality materials were
recommended for best reliability and safety. Supports and mountings were designed
for the system. The design must securely hold the system components and assure
safety for the persons around it.
4.4.1. Chain and Sprocket Selection
The size of chain for the application was made by referring to selection charts
relating shaft speed and design power. The design power, Pd, is determined from
the indicated power and application factors such as service factors, sprocket size
factor, and temperature factor.
𝑃𝑃𝑑𝑑 = 𝑃𝑃𝑓𝑓1 𝑓𝑓2 𝑓𝑓3
Where:
𝑓𝑓1
= Service factor
𝑓𝑓3
= Temperature factor
𝑓𝑓2
=
𝑃𝑃
= Indicated power
eq. 12
Sprocket size factor
31
The service factor was determined from details of the driver and driven equipment
by selection from the table below.
Table 1: Service Factor Selection
Characteristics of
Driven Machine
Characteristics of Driver
Smooth Running
Slight Shock
Moderate Shock
Smooth Running
1.0
1.1
1.3
Moderate Shock
1.4
1.5
1.7
Heavy shock
1.8
1.9
2.1
The sprocket size factor was determined through the number of teeth of the driver
and referring to the graph below. The temperature factor was neglected since the
system does not reach high temperature elevation.
32
Figure 7: Sprocket Size Factor Selection Graph
The engine (driver) and the wheel (driven) were categorized to have a moderate
shock and a heavy shock, respectively. With the engine having 1.0 kW of power
output, the design power was determined to be 5.88 kW.
The chain was then selected from the chain drive selection power rating graph for
American Standard Chains. From the graph, ANSI 35 is sufficient to transmit
5.88 kW of power from the engine to the wheel
Sprockets were selected based on the needed speed reduction ratio. It was
recommended that ratio on a drive system should not exceed 5:1; however, it is
required for a ratio of 13:1. The group then decided to distribute the ratio from
5.5:1 and 2.4:1. The new total reduction ratio is 13.2:1, which is acceptable for
the application of the system.
33
Figure 8: Drive Train Layout
4.4.2. Clutch System
The group decided upon using a centrifugal clutch due to its compactness and
simplicity in design. It was also chosen because of its availability in the market
and low cost. Because of its simple design, installation was also easy to carry out.
There were two designs that were constructed for the clutch system; 1) the use of
the default centrifugal clutch that came with the engine, and 2) the use of
motorcycle centrifugal clutch. Both systems were tested for reliability.
Using the default centrifugal clutch, it required less fabrication of supports and
shafting. It was observed that other teams have done this method and proved that
it was possible to drive the vehicle using it. However as endurance tests were
conducted, it was observed that the vehicle started to show problems regarding on
accelerating the vehicle from a stationary position.
34
As the system was inspected, dark pigmentation was found indicating that
extensive heat is experienced through friction. From this, the group decided to
change the clutch to a more durable motorcycle clutch.
The clutch used was from Yamaha MIO scooters which is readily available as
purchase for spare. Using the clutch would require to fabricate an adaptor to
securely hold it. The clutch was studied and the group determined how it is
installed into a conventional scooter. A shafting was then designed that would
connect the clutch to the fly wheel of the engine.
Figure 9: Clutch System Layout and Dimension
The clutch (green component) and the flange are bolted directly to the fly wheel
of the engine. The orientation of how the system was assembled is shown below.
A bearing (red component) would be attached to the clutch bell (blue component)
to keep it free-wheeling. A sprocket would then be attached directly to the clutch
bell that would serve as the driver for the drive train system. Another sprocket
was also fitted through the shaft that is connected to the starter motor.
35
Figure 10: Clutch System
The second design was tested for a number of endurance run and gave promising
results;

The system withstands the heat generated due to friction much better.

No sign of excessive wear was found after every run.

The vehicle accelerates much smoother.

Tugging was eliminated during acceleration.
With these results the system was kept and tested for further improvements.
4.5.
Reliability and Efficiency Tests
After construction, the system was tested for reliability and performance. Several test
runs were conducted to determine the fuel consumption of the system and how will it
perform. Test runs were conducted in the open grounds of Mapua Institute of
36
Technology, in the oval track of Malayan Colleges Laguna, and in Batangas Racing
Circuit.
Tests in Mapua Institute of Technology
focused on the maneuverability and
reliability of the vehicle. Vibrations, noise, and other faults are noted down and
adjustments are made to resolve the problems. The system was made sure not to
break down during these runs. However, fuel consumption was still recorded.
In Malayan Colleges Laguna, fuel consumption of the system was recorded. The
vehicle was tested on two settings: 1) Maximum engine throttle (constant speed), 2)
“Burn and Coast” method. Fuel consumption is then computed based on the distance
traveled over the amount of fuel consumed. Similar test was also conducted on the
BRC race track. Track terrain and how will it affect the vehicle’s performance was
noted.
4.6.
Dynamometer Test
The system was tested in Autoplus Sportzentrium. The test was conducted after the
race in Sepang, Malaysia, due to time constraints. The dynamometer test was used to
measure the wheel horsepower and torque of the vehicle. Further tests were also
conducted to support the results from the dynamometer test. The tests would aim to
account what lead the power loss of the system. (See results on Figure 13, page 43)
37
4.7.
Additional Test
After the dynamometer test, results were studied and further tests were conducted.
Some hypotheses were formulated as the cause of the power loss. The following are;
1) roller chain friction from roller guides and chain slack could account for the power
loss in the system, 2) slip occurs in the centrifugal clutch at a given speed that
accounts for the power loss.
The first hypothesis would require realignment of chains and sprockets that would
assure the required tension for the chains and proper center distances of the sprockets.
Removal of the roller guides or other substitutes may be installed depending on the
configuration of the chain drives.
Figure 11: Roller Guide Design
38
However, the system would undergo another dynamometer test to see the result of the
adjustments made. Unfortunately, another dynamometer test would be inappropriate
due to time constraints.
The second hypothesis would require the group to measure the speed difference of the
engine and the clutch bell. The engine speed could be measured through a
tachometer; however, there is no available equipment to measure the speed of the
clutch bell during runs. Because of this, the group will not be able to show the slip in
a quantitative manner.
The group decided to present the occurrence of slip in the centrifugal clutch through
temperature increase during operation. Temperature increase is present when slip
occurs due to abrasion between the clutch bell and the clutch shoe lining. At the
same time, when slip is not present, the clutch hub and the clutch bell can be assumed
to be in synchronous motion thus eliminating heat due to abrasion of the two surfaces.
The test was conducted in two set up;
1. Obtaining temperature increase at a given speed with no load.
2. Obtaining temperature increase at a given speed with load.
The first set up was set as base line of the test. The vehicle was suspended for the rear
wheel to rotate freely, thus eliminating resistance from the ground. Initial temperature
of the clutch was measured at engine idle speed before increasing engine speed. The
engine was operated at a given engine speed for ten minutes and temperature of the
clutch bell was taken after the set time. The clutch bell was then allowed to cool
39
down to the initial temperature. The procedure from the first set up was then repeated
for the second set up; however, the vehicle will be driven on ground for 10 minutes.
From the gathered data, the heat produced in the clutch was then computed using the
heat energy equation.
𝑄𝑄 = 𝑚𝑚𝐶𝐶𝑝𝑝 ∆𝑡𝑡
eq. 13
Where 𝑄𝑄 is the heat absorbed by the clutch bell, 𝑚𝑚 as the mass of the clutch bell, and
𝐶𝐶𝑝𝑝 as the specific heat of the material. The clutch bell weighs about 45.35 grams and
was said to be made of high-carbon steel material. The specific heat of carbon steel
was found out to be 0.49 kJ/kg-K; from these values, the energy absorbed by the
clutch bell was obtained.
Heat transferred is dependent on the rate of abrasion between the surface of clutch
bell and the clutch shoe lining. Testing the system with similar duration of time, the
rate of abrasion would then be dependent on the rate of revolution of the clutch hub,
in which, directly coupled to the engine. Thus, the heat transfer rate is directly
proportional to the rate of revolution of the engine. With this, a direct relationship
between the engine speed and increase in temperature of the clutch bell is established.
40
CHAPTER 5
DISCUSSION OF RESULTS
5.1.
Test Run Results
5.4.1. MIT Test Runs
The first test run done in the Mapua Institute of Technology quadrangle was
under conditions of constant low speed along an oval route. The succeeding runs
were under wet grounds and followed a fixed route across obstacles moving at
varying speeds.
Table 2: MIT Test Runs and Results
Driver
Fuel Consumption (Km/L)
Classification
Joshua Coronel
53
Endurance testing
Karlo Dujunco
60
Endurance testing
Karlo Dujunco
70
Endurance testing
Alex Luminarias
70
Endurance testing
Alex Luminarias
70
Endurance testing
5.4.2. Malayan Colleges Test Runs
The first runs of both drivers were under conditions of maximum speed to
complete 20 laps around the oval track and field course in Malayan Colleges
Laguna. The second run included engine shut-off and coasting and observing
varying speeds depending on the driver.
41
Table 3: Malayan Colleges Test Runs and Results
Driver
Fuel Consumption (Km/L)
Classification
Joshua Coronel
80
Endurance testing
Joshua Coronel
100
Burn and Coast
Karlo Dujunco
92
Endurance testing
Karlo Dujunco
116
Burn and Coast
5.4.3. Batangas Racing Circuit Test run
The test run held in Batangas Racing Circuit observed conditions of many slopes
in the terrain at different elevations. The speed of the car was maintained at its
maximum observed speed.
Table 4: Batangas Racing Circuit Test Run and Result
Driver
Fuel Consumption (Km/L)
Classification
Karlo Dujunco
99
Endurance testing
5.2.
Sepang International Circuit Run
The first official run was invalid due to electrical problem that occured while on the
track. The second run was for driver familiarization of the track moving at different
speeds and minimum coasting. The third and fourth run was under maximum coasting
and different speeds depending on the terrain.
42
Table 5: Sepang International Circuit Runs and Results
Driver
Fuel Consumption
Classification
Result
Alex Luminarias
-
-
Invalid
Karlo Dujunco
174
Driver Familiarization
Passed
Alex Luminarias
178
Burn and Coast
Passed
Karlo Dujunco
186
Burn and Coast
Passed
5.3.
Dynamometer Test
Data was obtained by conducting a dynamometer test for the vehicle. The obtained
result was a performance curve of wheel horsepower and torque. The result obtained
is shown in the figure below.
Figure 12: Dynamometer Result
Power transmitted to the wheel was found out to be 0.76 hp and torque at 1.13 ft-lbs
as seen in the graph. The indicated power of the engine was 1.3 hp at 7000 rpm and
1.2 ft-lbs at 5500 rpm as stated in the engine specification. It was observed that
43
maximum power of 0.76 hp was reached at 3750 rpm; and beyond this point, the
power started to decrease drastically.
The dynamometer test did not show any similarities of how the engine should
perform. The engine did not even enter the speed range of the indicated performance
curve. However, the transmission system provided sufficient torque to drive the
vehicle despite the amount of power loss.
5.4.
Additional Tests and Findings
5.4.1. Clutch Temperature Gradient Test
The occurrence of slip was observed through the difference in rotational speed of
the engine and the clutch bell; however, there was no available equipment to
Temperature, C
measure the speed of the clutch bell while the vehicle was being driven.
43
41
39
37
35
33
31
29
27
25
With Load
No Load
2000
2500
3000
3500
4000
4500
Engine Speed, Rpm
Figure 13: Clutch Temperature Increase
44
The graph presents the temperature reached by the clutch bell at a given engine
speed. It was observed that when the system was operating without load, the
temperature remained almost at constant level of 30℃. The system readily
engaged at 2,000 rpm with no sign of slipping. However, when operated with
load, the vehicle remained stationary at 2,000 rpm. The temperature reading after
the test was 41.4℃. Slipping at this point was clearly evident.
Temperature level decreased significantly at engine speed 2,500 to 3,000 rpm. It
was observed that at 2,500 rpm the vehicle started to move slightly in a tugging
motion. At 3,000 rpm, temperature level reached to 31.2℃ which were almost the
same as the test without load. Furthermore, the vehicle started to run smoothly at
this speed. Higher engine speeds were able to drive the car smoothly; however,
temperature level started to increase gradually.
From the obtained results, it can be concluded that the centrifugal clutch is
creating friction before engagement at 2,000 rpm. At this point, frictional loss is
gained in the system, decreasing transmitted power. The clutch engages at 3,000
rpm minimizing slip and frictional loss; however, this does not suggest that the
system is running efficiently. It is said that the engine must run at 5,500 rpm to
generate its maximum torque and power.
The system tends to slip when engine speed reached 3,500 rpm. The increase in
heat transfer rate suggests that the rotation of the engine and the clutch bell is not
synchronous, in a sense that the engine is rotating faster than the clutch bell.
Running at this speed range would cause constant power loss and inefficient
45
power transmission. Furthermore, the friction at the range of 2,000 to 2,500 rpm
causes power loss as the driver constantly increase speed from zero during “burn
and coast” conditions.
5.4.2. Clutch Power Loss
The group studied the results from the initial engine test and the results from the
dynamometer test to determine the amount of power loss from the centrifugal
clutch.
From the initial engine test, torque was observed to increase gradually up to 5,000
rpm. However, from the engine performance curve, torque would decrease at
5,500 rpm onwards. Test runs indicate that the engine starts to transmit power
(drive the vehicle) at 3,000 rpm onwards.
From the dynamometer test, it was observed that torque output at 2,000 to 3,000
rpm is higher than the initial test. This is due to the additional force applied in the
flywheel of the dynamometer to surpass the inertial force required for it to start
rotating. Since power is transmitted at 3,000 rpm onwards, it can be said that
torque developed by the engine is not affect at 3,500 rpm onwards.
Engine can transmit power of 0.418 kW at 4,000 rpm. From the dynamometer
test, the engine starts to lose power at this speed. Temperature of the clutch bell at
this speed starts to increase as indicated from the temperature gradient test.
46
From 0.418 kW to 0.397 kW, power loss was 5.1% at 4,000 rpm; furthermore,
percent power loss increases as engine speed increases. Observed power loss was
13.7% and 20.9% at 4,500 rpm and 5,000 rpm, respectively. It was also observed
that the frictional force required was not sufficient to lock the clutch shoe and the
clutch bell, causing slip and inefficient power transmission.
47
CHAPTER 6
CONCLUSION AND RECOMMENDATIONS
6.1.
Conclusion
In designing a drive train system, the proper speed ratio is required to transmit
sufficient amount of torque and power from the engine to move the vehicle. Based
from track characteristics, vehicle dynamics, and design factors, the speed ratio was
determined to be 13:1 reduction for a 100 kg combined vehicle and load weight.
The selection of the drive train system dictates how the engine will perform. Higher
speed reduction ratio would increase torque applied on the wheel; however, energy
would be wasted if the applied torque is in excess of the required amount. On the
other hand, lower speed reduction ratio would transmit lower output of torque on the
wheel. The problem with this system is that if the torque being transmitted is enough
to move the vehicle. Neither high nor low reduction ratio is efficient since power
transmitted by the engine is both wasted on the said settings.
The dynamometer test shows the designed system was only able to deliver power of
0.76 hp and torque of 1.13 ft-lbf. This yields to only 58 percent of the maximum
power that the engine is able to generate. The main factor for the power loss is due to
friction and slip from the centrifugal clutch. Losses from the chain drive system can
also be accounted.
48
The vehicle was able to complete three runs in which the highest performance
obtained is 186km/L from the Shell Eco-Marathon Asia 2011 contest. The fuel
consumption is affected by factors such as driving behavior, aerodynamic parameters,
weight of the car, and the track layout. Although there was no direct correlation
between the drive system efficiency and fuel efficiency, it can be concluded that the
design and variations of the drive system of the vehicle is still factor.
6.2.
Recommendations
For the improvement of the designed power train system or for students wanting to
conduct the same study, the group has the following recommendations:
1. Study the feasibility of applying toothed belt drive as transmission
system for the vehicle.
Toothed belts may be more advantageous because they weigh less than roller
chains. However, belt size and length should be considered in designing the
system. Designs may be restricted because belt sizes and lengths are available
only in standards.
2. Establish a comparison of performance between the centrifugal clutch
and friction plate clutch.
49
Study the difference in performance of a centrifugal clutch and a frictional
plate clutch. Knowing further the applications and behaviour of the two types
of centrifugal clutch, would help improve the design of the system.
3. Study the feasibility of applying a Continuous Variable Transmission
(CVT) system to an engine.
The CVT system would vary from different gear ratios that depend on the
amount of torque required to move the vehicle at a certain condition. A
system that can vary gear ratios could maximize torque output at any speed.
50
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Drake, George R. Small Gasoline Engines maintenance, troubleshooting, and repair,
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Ganesan V. Internal Combustion Engines, 1st ed. Tata McGraw-Hill Publishing Company
Limited, 1994. Pages 266-267
Ganesan V. Internal Combustion Engines, 1st ed. Tata McGraw-Hill Publishing Company
Limited, 1994. Pages 273-275
Heywood, John B. Internal Combustion Engine Fundamentals, International ed.
McGraw-Hill Book
Company,1988. Page 1.
IEEE Xplore Digital Library. http://ieeexplore.ieee.org/Xplore
Faires, Virgil M. Design of Machine Elements, 4th ed. The Macmillan Company, 1969.
Pages 464 – 469.
Introduction to Mechanical Drive Systems, 2nd ed. Amatrol Incorporated, 2000. Page 2.
Introduction to Mechanical Drive Systems, 2nd ed. Amatrol Incorporated, 2000.
Page 25-26
Kensey, William. How to Calculate the Pulley and Belt Ratio, 2010.
http://www.ehow.com/how_6196301_calculate-pulley-belt-ratio.html.
Kharagpur IIT, Module 13 Belt Drives, Version 2.
http://nptel.iitm.ac.in/courses/Webcourse-contents/IIT%20Kharagpur/
Machine%20design1/pdf/mod13les1.pdf. Pages 1 - 7.
Phelan, Richard M. Fundamentals of Machine Design, 3rd ed. McGraw-Hill Incorporated,
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Shigley, Joseph and Charles Mischke. Mechanical Engineering Design, 6th ed.
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Uicker, J.J. and et.al. Theory of Machines and Mechanisms,
Oxford University Press, 2003.
51
APPENDIX A: Proposed Time Table
52
APPENDIX A: Proposed Time Table (Continued)
53
APPENDIX B: Chain Selection Graph
54
APPENDIX C: Sprocket Selection Table
55
APPENDIX D: Computations
Speed Reduction Ratio Computation
𝐿𝐿𝑊𝑊𝑟𝑟 = ℎ𝑊𝑊𝑊𝑊𝑊𝑊𝑊𝑊𝑊𝑊 + 𝑙𝑙𝑙𝑙𝑙𝑙𝑙𝑙𝑙𝑙𝑙𝑙
Where:
𝐿𝐿 = 200 𝑐𝑐𝑐𝑐
𝑙𝑙 = 100 𝑐𝑐𝑐𝑐
ℎ = 38.1 𝑐𝑐𝑐𝑐
𝑊𝑊 = 𝑚𝑚𝑚𝑚; 𝑚𝑚 = 100 𝑘𝑘𝑘𝑘
𝜃𝜃 = 3°
56
APPENDIX D: Computations (Continued)
Solving for 𝑊𝑊𝑟𝑟 ;
𝑊𝑊𝑟𝑟 =
0.381𝑚𝑚[(100𝑘𝑘𝑘𝑘 )(9.81) sin(3°)] + 1𝑚𝑚[(100𝑘𝑘𝑘𝑘) �9.81 𝑚𝑚�𝑠𝑠 2 � cos(3°)]
2𝑚𝑚
𝑊𝑊𝑟𝑟 = 499.61 𝑁𝑁
Solving for the Tractive Force on the wheel;
𝐹𝐹𝑡𝑡 = 𝐹𝐹𝑓𝑓 + 𝑊𝑊 sin 𝜃𝜃
Where:
𝐹𝐹𝑓𝑓 = 𝐶𝐶𝑟𝑟𝑟𝑟 𝑊𝑊𝑟𝑟
Then;
𝐶𝐶𝑟𝑟𝑟𝑟 = 0.02
𝐹𝐹𝑡𝑡 = 𝐶𝐶𝑟𝑟𝑟𝑟 𝑊𝑊𝑟𝑟 + 𝑊𝑊 sin 𝜃𝜃
𝐹𝐹𝑡𝑡 = 0.02 [(499.61 𝑁𝑁)(cos(3°))] +(100𝑘𝑘𝑘𝑘) �9.81 𝑚𝑚�𝑠𝑠 2 � (sin(3°))
Determining the required torque;
𝐹𝐹𝑡𝑡 = 61.32 𝑁𝑁
𝑇𝑇𝑤𝑤ℎ𝑒𝑒𝑒𝑒𝑒𝑒 = 𝐹𝐹𝑡𝑡 𝑟𝑟
𝑇𝑇𝑤𝑤ℎ𝑒𝑒𝑒𝑒𝑒𝑒 = (61.32 𝑁𝑁)(0.254 𝑚𝑚)
Speed reduction ratio;
𝑇𝑇𝑤𝑤ℎ𝑒𝑒𝑒𝑒𝑒𝑒 = 15.58 𝑁𝑁 − 𝑚𝑚
𝑇𝑇𝑤𝑤ℎ𝑒𝑒𝑒𝑒𝑒𝑒 × 𝑁𝑁2𝑛𝑛𝑛𝑛 𝑔𝑔𝑔𝑔𝑔𝑔𝑔𝑔 = 𝑇𝑇𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒 × 𝑁𝑁𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒
Take engine torque and engine rpm at best efficiency point, at 1.2 N-m and 5500 rpm
Leading to a ratio of
𝑁𝑁2𝑛𝑛𝑛𝑛 𝑔𝑔𝑔𝑔𝑔𝑔𝑔𝑔 = 423.75 𝑟𝑟𝑟𝑟𝑟𝑟
𝑆𝑆𝑆𝑆𝑆𝑆𝑆𝑆𝑆𝑆 𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟 𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟𝑟 =
𝑁𝑁𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒𝑒
= 𝟏𝟏𝟏𝟏. 𝟎𝟎
𝑁𝑁2𝑛𝑛𝑛𝑛 𝑔𝑔𝑔𝑔𝑔𝑔𝑔𝑔
57
APPENDIX D: Computations (Continued)
Chain and Sprocket Selection
𝑃𝑃𝑑𝑑 = 𝑃𝑃𝑓𝑓1 𝑓𝑓2 𝑓𝑓3
Where:
𝑃𝑃 = 1.0 𝑘𝑘𝑘𝑘
𝑓𝑓1 = 2.1
𝑓𝑓2 = 2.8
Then;
𝑓𝑓3 = 1.0
𝑃𝑃𝑑𝑑 = (1.0𝑘𝑘𝑘𝑘 )(2.1)(2.8)(1.0)
𝑃𝑃𝑑𝑑 = 5.88 𝑘𝑘𝑘𝑘
Characteristics of
Driven Machine
Smooth Running
Moderate Shock
Heavy shock
Characteristics of Driver
Smooth Running
Slight Shock
Moderate Shock
1.0
1.1
1.3
1.4
1.5
1.7
1.8
1.9
2.1
58
APPENDIX E: Clutch Slip: Temperature Gradient
No Load Condition
Engine Speed, rpm
Temparature, C
Specific Heat,
kJ/kg-K
Heat, J
Heat Transfer Rate,
J/min
Initial
Final
dt
2000
27.8
30.3
2.5
0.49
55.56
5.56
2500
28.0
30.1
2.1
0.49
46.67
4.67
3000
28.3
30.5
2.2
0.49
48.90
4.89
3500
28.1
30.2
2.1
0.49
46.67
4.67
4000
28.4
30.2
1.8
0.49
40.01
4.00
4500
28.2
30.4
2.2
0.49
48.90
4.89
Specific Heat,
kJ/kg-K
Heat, J
Heat Transfer Rate,
J/min
With Load Condition
Engine Speed, rpm
Temparature, C
Initial
Final
dt
2000
28.1
41.4
13.3
0.49
295.60
29.56
2500
28.3
32.5
4.2
0.49
93.35
9.33
3000
28.2
30.8
2.6
0.49
57.79
5.78
3500
28.4
31.2
2.8
0.49
62.23
6.22
4000
28.0
33.3
5.3
0.49
117.80
11.78
4500
28.3
35.8
7.5
0.49
166.69
16.67
59
APPENDIX F: Clutch Slip: Power Loss
60
APPENDIX G: List of Figures, Equations and Tables
List of Figures
Label
Figure 1
Figure 2
Figure 3
Figure 4
Figure 5
Figure 6
Figure 7
Figure 8
Figure 9
Figure 10
Figure 11
Figure 12
Figure 13
Title / Caption
Agimat Transmission System
Dalhousie Supermileage Team Drive-train system
Example of a Drive Arrangement
Initial Engine Test Results.
Initial Design of Drive Train System
Vehicle Free Body Diagram
Sprocket Size Factor Selection Graph
Drive Train Layout
Clutch System Layout and Dimensions
Clutch System
Roller Guide
Dynamometer Result
Clutch Temperature and Engine Speed Relationship
Page
7
10
23
26
27
28
33
34
35
36
38
43
44
List of Equations
Label
Equation 1
Equation 2
Equation 3
Equation 4
Equation 5
Equation 6
Equation 7
Equation 8
Equation 9
Equation10
Equation 11
Equation 12
Equation 13
Title / Caption
Torque
Power
Design Power
Drive Ratio
Length of Chain
Summation of Moment
Tractive Force
Tractive Force and Frictional Force
Torque on wheel
Speed Ratio
Speed Reduction Ratio
Design Power
Heat Energy Equation
Page
19
20
21
22
23
29
29
30
30
30
30
31
40
61
APPENDIX G: List of Figures, Equations and Tables (Continued)
List of Tables
Label
Table 1
Table 2
Table 3
Table 4
Table 5
Title / Caption
Service Factor Selection Table
MIT Test Runs and Results
Malayan Colleges Test Runs and Results
Batangas Racing Circuit Test Run and Result
Sepang International Circuit Runs and Results
Page
32
41
59
42
43
62