2005 SAE Supermileage® Design Competition - College of ...

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University of Massachusetts


Amherst


Department of Mechanical & Industrial Engineering

2005 SAE Supermileage® Design Competition

Design
Report






University of Massachusetts


Amherst

Department of Mechanical and Industrial Engineering


March 31, 2005




Design Team:


Engine/Powertrain Team


Keith Washburn


Team Leader

Dave Bielunis

Steve Bofill

Mike Hirsch

Ryan Lepard

Rob Mastria

Zack Shanley


Electronic Fuel Injection Team


Jordan Crafts

Mike Flanagan

Dorinda Garcia

Andy McCarron

Ji
m MacDonald









Body/Chassis Team


Sean Stewart


Team Leader

Alex Fok

Katie Hartig

Caroline Linehan

Chiu Yip




Project Advisor


Professor Robert W. Hyers






___________________________ _________

Prof. Robert W. Hyers Date

Universit
y of Massachusetts, Amherst

Dept. of Mechanical & Industrial Engineering



University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

2

Table of Contents




Mission Statement and Design Goals

................................
................................
.............................

3

In
troduction

................................
................................
................................
................................
.....

3

Basic Vehicle Configuration

................................
................................
................................
...........

4

Engine

................................
................................
................................
................................
.............

6

Overview

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................................
................................
................................
.....

6

Cylinder Head: Combustion Chamber

................................
................................
........................

7

Cylinder Head: Intake and Exhaust Valves

................................
................................
................

8

Cylinder Head: Valve Train

................................
................................
................................
........

9

Cylinder Head: Connecting Rod

................................
................................
................................
.

9

Electronic Fuel Injection

................................
................................
................................
...............

11

Design Process

................................
................................
................................
..........................

11

Parts Selection

................................
................................
................................
...........................

12

Engine Control Unit

................................
................................
................................
..................

13

Power Electronics

................................
................................
................................
.....................

14

Tuning Notes

................................
................................
................................
.............................

14

Experimental Projects

................................
................................
................................
...................

14

Flywheel

................................
................................
................................
................................
....

14

Heat Exchanger

................................
................................
................................
.........................

16

Powertrain Configuration
................................
................................
................................
..............

17

Brake S
ystem and Vehicle Performance

................................
................................
.......................

18

Running Gear

................................
................................
................................
................................

18

Body/Aerodynamic Shell

................................
................................
................................
..............

20

Aerodynamic Considerations

................................
................................
................................
....

20

Safety Considerations

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................................
................................
...............

22

Cost Estimate and Manufacturing Methods

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................................
..................

24

Body Team

................................
................................
................................
................................

24

Engine Team

................................
................................
................................
.............................

24

Fuel Injection Team

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................................
................................
..................

25

Total Donations

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................................
................................
.........................

25


Appendix A

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................................
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................................
...

26

Appendix B

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................................
...

27











University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

3

Mission Statement and Design G
oals


The overall goal of this project is to design and fabricate a single
-
person, fuel efficient vehicle

powered by a

modified, iso
-
octane burning

Briggs and Stratton four
-
cycle 3.5 hp motor.
The
objective

of the U
niversity of Massachusetts

Supermileage
Vehicle (
UMass
SMV) Team
is
to
apply our engin
eering knowledge and experience

to design and build a vehicle that will achieve
in excess of 1000 miles per gallon while s
atisfying all of the SAE Supermileage® Competition
rules and requirements. This
target

will be met by designing and
constructing

a lightweight,
aerodynamic shell that will house a custom

built,

fuel
-
injected motor designed around the
existing
Briggs and Stratton crankcase.



Introduction


The
UMass SMV

design team
split themselves into

three

individual
groups

who each
concentrate
d

on optimizing a particular aspect of the vehicle. The Body and Chassis team was
comprised of five
senior
mechanical engineering
students

who
took

on the task of designing
and
building
the vehicle chassis,
body
shel
l, steering
assembly
,
and
braking system as well as
choosing driver positioning

and fabricating other important vehicle components to ensure drive
r

safety.
The main goal for this group was to construct an aerodynamic shell around a rigid,
lightweight chas
sis that would sufficiently support the driver, engine, and all required vehicle
components.


The Engine and Powertrain team consisted of seven

senior

mechanical engineering
students
dedicated to maximizing the efficiency of the motor by designing and fabr
icating custom engine
components with the goal of reducing the motor’s fuel consumption while still providing
adequate power to propel the vehicle.

These students have extensive knowledge in the areas of
machining and engine tuning which allowed them to e
xperiment with some new and
innovative

id
eas for this year’s competition such as a heat exchanger and a flywheel.


T
he third group was comprised of four
senior
electrical engineering students and one

senior

mechanical engineering student who
a
re
committed
to developing an Electronic Fuel Injection
(EFI) system to deliver fuel to the engine. EFI provides the ability to adjust air
-
to
-
fuel ratios
automati
cally which c
ould

dramatically

reduce
fuel consumption when tuned properly. This is
the first year the UM
ass team has attempted to incorporate EFI into their Supermileage® vehicle
.
Incorporating EFI will bring UMass to the next level to compete

with other more successful
teams who have already taken advantage of using
fuel injection in past Supermileage®
com
petitions
.


In order to achieve these goals, t
he students made use of the resources provided to them
by

the
University of Massachusetts including, but not limited to, the Department of Mechanical and
Industrial Engineering’s stude
nt machine shop as well as

the U
nive
rsity
-
owned wind tunnel.
Additional resources were provided through generou
s donations by vehicle sponsors interested in
supporting such a

unique project.



University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

4

Basic Vehicle Configuration


The overall vehicle configuration illustrates the efforts pu
t into making the car as light
weight

and
compact

as possible while minimizing the aerodynamic drag. The vehicle uses three wheels; one
rear wheel which
is
powered by

the motor and two
front
wheels which steer the vehicle. This
orientation
provides
a bala
nced

foundation which satisfies the
stability
requirements while
allowing for maximum
driver
visibility through the
unobstructed front windshield. This t
hree
-
wheeled configuration
was determined to be the

best design

for

an aerodyn
amically efficient
body
shell.


The driver position was
then
determined
,

which facilitated
the design of the frame
. In order to
most efficiently use the space as well as ensuring
the rules were adhered to
, the driver position
was chosen to be a reclined position with the knees b
ent over the front axle to reduce overall
vehicle length
. This orientation made

the most of the space that
will
be covered by the

body
shell.
T
he driver

s head and shoulders
are
angled slightly upward so that the driver
can look

straight between his or h
er knees while remaining
reclined
.
Figure 1

shows
the chassis and the
driver orientation.

This
body position

allow
s

the car to be at a minimal height off of the ground
where

the driver’s head would be the highest point anywhere along the car.

T
he engine

was
placed

directly behind the driver, separated by a firewall for the
protection of the driver
.


Figure 1



Chassis and driver positioning


The frame
design was

based upon the driving position and the prima
ry driver’s
physical

dimensions. A simple “lad
der” style frame
composed

of rectangular aluminum tubing
was
chosen based on the ability to make
the chassis

lightweight while still providing plenty of
strength to support the load of the driver and all of the vehicle components.

This is shown in
Figure
2
.

Aluminum
6061
-
T651
was chosen
for the frame material
because of

its low cost, high
strength
-
to
-
weight ratio, and the eas
e
of welding
. Other materials such as carbon fiber could
have provided better frame design alternatives, but in the end, mate
rial c
ost was the driving
force.


University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

5



Figure
2



Top view of the Chassis


T
he
vehicle
frame was
design
ed

so that the body shell would have three inches of ground
clearance

which was deemed
sufficient

to clear any major bumps and cracks at the Eaton
race
track
.
T
h
is

low center of gravity
made

the vehicle more stable and
improved vehicle handling
.

As show
n in Figure 2
,
the front wheels are located at the widest part of the frame.


From there,

the frame goes straight back where the driver and engine are located and

then
gets narrower as it
tapers back to the rear wheel.

This allow
s

for the body to have
a
shape whic
h starts out wide and
gets narrower, similar to an airfoil. This
type of
shape
is

optimal for provi
ding the least amount
of drag.

Figure 3

shows the pla
ster mold used

for wind tunnel testing and

also
to create the
fiberglass s
hell
.


Overall, the vehicle will measure 115 inches long and 33 inches in width at its
widest point. The vehicle will be 22 inches tall at its maximum height.



Figure 3
-

Plaster
Mold of Body




University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

6

Engine

Overview

A stock

Briggs and Stratton 3.5 h
p

motor

produces too much power for implementation into a
minimum speed car. So, the first major engine team goal was to reduce this power output to a
point where it would still be able to ac
celerate the car while burning less fuel than the stock
motor. The stock motor was measured to have approximately 1.2 brake specific fuel
consumption (BSFC) which represents the amount of fuel flow in pounds per hour divided by the
horsepower. Therefore,

a reduction in both BSFC and horsepower would be necessary
for the
goal of

minimal fuel consumption and optimal vehicle performance. Other major goals were to
reduce the power output by reducing the overall cubic inch displacement (CID) of the motor and
to increase efficiency by increasing the compression ratio, airflow, and redesigning the cam
profiles. These goals were achieved by completely designing and manufacturing a custom
overhead cam cylinder head for use on the
stock Briggs and Stratton crank
ca
se. The
compression ratio was increas
ed by using a domed piston and

minimizing the volume of the
combustion chamber. The cams were designed with minimal overlap
,

maximum lift and dwell
rates. Figure
4

shows the partial assembly of the race engine with t
he custom cylinder head.


Figure 4

-

Engine and Engine Head
-

Partial Assembly


Additionally, two experimental projects were developed this year to further increas
e the
efficiency of the motor by employing other energy
-
saving techniques. A heat exchanger was
designed t
o

increase

thermodynamic efficiency by heating the intake air before it enters the
engine. Also, a hybrid powertrain was investigated utilizing a me
chanical energy storage
flywheel for the purpose of providing energy to the vehicle when the engine is not in use.



University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

7

Cylinder Head
:
Combustion Chamber

The combustion chamber of the cylinder was designed by first analyzing various chamber shapes
such as he
mispherical, quench, and wedge. Due to the advantages of

quench
and hemi

designs
providing the most efficient combustion, they were selected as the base for which the design was
constructed. This decision immediately resulted in a h
igher initial compress
ion ratio
without
having to dome the piston and
provided

better air
-
fuel mixing properties. Since one of the
design goals
for

the cylinder head was to maximize airflow into and out of the engine, the intake
and exhaust runners were designed to be as strai
ght as possible which resulted in the valves
having to be angled. The intake runner was designed a
round using an 18mm (0.71 in
) carburetor
and
the
exhaust
runn
er was designed with a 0.52
-
in

diameter in order to provide an unrestricted
exit of the exhaust
gases through the cylinder head
.

The intake runner is cylindrical in shape
with rough walls and the exhaust runner is a “D” shape. The “D” shape promotes flow around
the valve seat and with a proper exhaust pipe setup, prevents exhaust gas from re
-
enteri
ng the
combustion chamber.


The engine bor
e was selected to be 1.70 in

due to the availability
of pistons and seals. The
piston modification, shown
below
in Figure 5,

assisted in reducing the overall engine
displacement. In order
to place the valves in t
he 1.70
-
in

combustion chamber, the valves were
angled 15 degrees from the horizontal. By placing

the valves opposite

each other, a closed
chamber was created, similar to the hemi head. This allow
ed

the greatest amount of energy upon
gasoline explosion to

be deflected directly on top of the piston.


Protruding from the valve seats are conic shapes which enhance the hemi effect

and eliminate
sharp surfaces
which conduct more heat and

can lead to melting. The point where the quench
pads and the combustion c
hamber meet is rounded to decrease th
e

threat of melting. However,
the rounding is not enough to negate the turbulent effect of the quench pads.


Using Pro/Engineer, the volume of the combustion chamber

shown in Figure 5

was found to be 9

cc.

When combine
d with a 1.70
-
inch bore, a 1.74
-
inch stroke, and a 5.64 cc domed piston the
compression ratio becomes 15.4
:1.



Figure 5

-

Combustion chamber with dua
l

q
uench pads and conic side walls;

Race Piston



University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

8

Cylinder Head: Intake and Exhaust Valves

B
oth the intake and exhaust valves were designed to have the maximum amount of airflow while
still retaining high strength properties. Due to the high stress conditio
ns found in valves along
with the exposure to high temperatures, material selection was very important. Nitronic 50
Stainless Steel was selected
because it is fully austenitic; therefore its strength is not affected by
heating cycles. Nitronic 50 is a mo
di
fied 300 series stainless steel

which has relativity high
strength
,

even at elevated temperatures.


In order to increase the airflow, the stem diameter was decreased. The upper valve stem was
selected to be
0.25
-
in diameter

simply
because bronze valve
guides were available

for this size
.
If the
0.25
-
in
diameter had been carried throughout the valve it would have encompassed 12.
4
%
of the intake port and 23.1% of the exhaust port. These numbers are much too high for optimal
airflow.


T
o compute the red
uction

needed

in
the
valve stem diameter
,

three methods were analyzed;
fatigue life, energy, and dynamic impact.

Prior to performing

any
calculations,
the
Comp Cams
GP960
-
kit

valve spring

was compression tested using a Universal Instron #4401
. The force

versus displacement

curve

was graphed
for the spring
and the integral was taken from 0.2 to 0.6
deflections to compute the total energy.


The
0.120
-
in
valve diamete
r was selected because it performed well under the parameters set
forth during testing
.
The valve diameter is optimal

for fatigue life;
it can

with
s
tand 240.3 lb of
repeated force

and has a total deflection of about
0
.0005

in
.


To verify calculations, Pro/Mechanica w
as used and a force of 240.3 lb was placed on the valve
head. T
he valve was
constrained at the larger diameter of the stem to simulate the valve in an
engine. The FEA
, shown

below

in Figure 6
,

confirm
ed

a maximum stress in the stem of 21 k
si,
h
owever a slightly higher stress of 23 k
si was found in the stem near the valve head. T
his radius
is very large when compared to the valve stem
. A possible explanation for this is that the

large

mesh
used to solve for the stress
es

attributed to
a

possible error.



Figure 6
-

FEA result for 0.120
-
in valve stem diameter showing
a stress of 2
1 ksi

in
the
valve stem.


D
esigning conservatively, the 0.120
-
in diameter valve stem

takes

into account the worst
case

University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

9

scenarios found in an engine. Since

the valve
s

w
ere

designed to meet these parameters
,

it is
more tha
n adequate for normal service.


Cy
linder Head: Valve Train

The valve train was designed to eliminate a high p
ercentage of frictional losses while providing
durabil
ity

and reliability.

A

du
a
l overhead cam was utilized to eliminate the need for rocker
arms and push rods. To virtuall
y eliminate the tangential force on the solid lifters, a roller lifter
was used. This allows for greater valve opening and closing rates,
which further contributes

to
the volumetric efficiency.


In order to lower the overall height of the cylinder head, t
he minimum lifter bearing length was
used.
T
o prevent rocking of the lifter
, the length was calculated to be

0.75 in
. At

this length,

the
opposed ends of the bearing e
xperience a force of 136 lb. This is achieved by neglecting
the
effect of the roller l
ifter.


Figure
7

shows the valve train assembly in the cylinder head.



Figure 7

-

Cylinder head valve train assembly


Cylinder Head: Connecting Rod

The original

Briggs and Stratton connecting rod
supplied with the

3.5
hp

was designed for basic
use in a
n

environment

where efficiency is not the major goal
.


The driving force to redesign the
connecting rod was based on a need for a longer profile to account for a lifted head design.

The
design parameters
were

to
create

a longer connecting

rod, but also to m
inimize the weight and
maximize
its

strength. All of the concepts were focuse
d on the redesign of the rod in
between
the two bearing surfaces. The factory connectin
g rod is shown below in Figure 8
.




University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

10


Figure 8

-

Factory Connecting rod


Four basic

concepts were considered for the shape of the cross section for the redesigned
connecting rod: a circular, rectangular, I
-
Beam, and H
-
Beam profile.

These are shown in Figure
9
.

The I
-
beam connecting
rod is the most popular in industry and can be found i
n almost every
automotive engine manufactured today. This profile is relatively cheap and very reliable. H
-
beam profiles are becoming more popular and a
re said to perform well in high
-
power, low
-
speed

applications where com
pressive loads are very large.


Figure
9

-

Various Design Considerations for Connecting Rod Profile (circular,

rectangular, I
-
Beam, and
H
-
Beam; respectively)


The first concept analyzed was the cylindrical beam cross section, which changed the profile to a
pure rod. This allow
ed

for a

greater clearance of the cylinder walls during piston motion by
aligning the highest point of the rod with the curved surface of the cylinder wall. The design is
simple and does not involve a complicated FEA analysis

and the

profile would have also been
easy to cast for final production. The on
ly

area of concern when

using the cylindrical beam

rod
is experiencing bending from the motion of the piston and crankshaft. The cylinder profile
provides the same resistance to bending in all directions, where it
is not needed.



University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

11

The next
idea

utilizes a rectangle profile

having

a wider profile
which faces

the highest
bending
stress
es. This compensates for the high

bending stresses in the rotating direction and smaller
bending forces in the perpendicular direction
. With proper choice of material properties, the
physical dimensions could be reduced resulting in a reduction in overall weight and
increasing
rod
-
to
-
wall clearance.

An alteration of this concept (I beam), which incorporates shallow fillets
to reduce we
ight, although not optimized
,

is also shown

in Figure 9
.


The H
-
beam design allows for a lightweight hollowed profile while providing maximized
strength against bending in the direction of motion. The higher resistance to bending is
maintained by ribs in
the material extending outward in the plane of motion. Pocketing and
fillets remove excess material reducing the mass. The removal of material however will not
reduce the overall profile, which increases the risk of collision between the rod and cylinder

wall. The H
-
beam concept is widely used in many high performance race companies to
maximize the strength
-
to
-
weight ratio.


The design team was

able to narrow down
their choice to the
I
-
Beam and H
-
Beam
profiles
by
eliminating the circular and rectangular

designs early in our research.
The 3.5 h
p Briggs engine
does not produce any appreciable compressive loads to show any difference in performance
between the I
-
beam and H
-
beam profiles. The I
-
beam was chosen as the final design concept
because it allows

f
or the best strength
-
to
-
weight ratio with a factor of safety of four. It also
allows for a total weight savings of 48% compared to a rectangular profile.
A
s a result, it was

manufactured

and installed on the
race
engine.

The race engine con
necting rod

is shown in
Figure 10
.



Figure
10

-

Connecting Rod


Electronic Fuel Injection

Design Proces
s

Most commercially available fuel injector kits are designed for
four cylinders or more and
incorporate injectors with much larger flow rates than what is necessary or

even practical for the
Superm
ileage
®

Vehicle. Following a thorough examination of implementations
,

a combination
feed
-
forward/feed
-
back control scheme u
tilizing various sensors on the engine

was used
.
Please

University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

12

see

the block diagram in Appendix

A

for a complete overview of the system architecture.


On the intake manifold inlet, there is a temperature sensor which delivers the temperature of the
air outside
the engine. Behind the throttle valve, there is a manifold
-
absolute pressure sensor.
The pressure and temperature delivered by these

sensors allow one

to obtain a general picture of
the volume of atmospheric gases within the cylinder. Using some basic c
hemistry and the ideal
gas law, the amount of fuel necessary for stoichiometric combustion can be obtained. These two
sensors along with the utilization of the ideal gas law constitute the framework of the feed
-
forward portion of the control scheme.


The
fuel is metered by the electronic fuel injector mounted in the intake manifold. The timing of
the pulse is determined by information taken from the cam angular position sensor. Once the
feed
-
forward control has made its injector fuel volume “guess”, the
combustion process occurs.
The primary indicator of the accuracy of that guess is the exhaust gases which contain either
excess oxygen or fuel. The system utilizes a wide
-
band oxygen sensor to determine the content
of either of these elements and the nex
t “guess” is tuned accordingly.


Lastly, to aid in the system implementation and debugging process, a data acquisition system
was designed. The system is designed to give eight analog input channels and to sample at
various frequencies depending on the
frequency of the signal to be recorded. The sampling rate
can be configured anywhere from 30 Hz up to 6 kHz and implements storage for approximately
2 million samples.


Parts Selection

The most important part selection was the injector. To select the p
art, the
amount of fuel needed
for stoichiometric combustion at 1 atm within the cylinder and the amount of time available t
o
meter that fuel (at 3500 RPM) was calculated to
about 6.27 g/sec.
A

Synerject model used on
the Aprilia SR
-
50 scooter

was found t
o meet the specifications
. The flow rate specified by
Synerject for that injector is 6.7 g/sec at a fuel pressure of 43 psi, the closest to the ideal
.

From
the selection of this part followed the selection of the fuel pump, pressure regulator, and
sensor
s.
The intake manifold

for this project was custom machined and designed
to incor
porate the
chosen fuel injector,
required sensors

and throttle valve
. It is built around the Aprilia SR
-
50
injector with ports for the manifold pressure sen
sor and appropria
te fuel lines. While designing,
goals included aiming towards the fuel injector, positioning close to intake valve, mounting to
the head and contained in a small package. The injector itself had to sit in an open bath of
pressurized fuel. From this spec
ification, a pump, fuel lines and a pressure regulator were
selected in order to safely provide the injector with the required fuel pressure.



The primary sensors used in this system are the temperature sensor, manifold absolute pressure
sensor, cam angul
ar position sensor, and the exhaust gas oxygen sensor.

In accordance with the
manufacturer’s specifications, the oxygen sensor was mounted on the exhaust pipe 18 in. from
th
e exhaust valve. A Hall Effect

sensor was selected to monitor camshaft position.

This sensor
came as t
w
o parts: a magnetic ring and sensor. A camshaft extender was machined in order to
provide a place for the magnetic ring. A custom bracket was made to place the sensor close

University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

13

enough to the ring to sense the rings polarity, thus dete
rmining the position of the camshaft and
consequently when the intake/exhaust valves would be open.

The temperature sensor is
implemented with a simple tempe
rature controller transistor
but will likely be upgraded to a
much more durable sensor to accommod
ate the use of a heat exchanger on the finished engine.
T
he intake gases will reach temperatures of about 200 degrees F. The manifold absolute
pressure sensor was selected with durability and reliability in mind and is also equipped to
handle the high te
mperatures
.


The cam angular position sensor is implemented by a Hall Effect
sensor with a linear field response and a toroid magnet. The use of the toroid allows the sensor
to give a unique DC value for 180 degrees of the cam’s rotation and the slope of
the output sine
wave is used to determine which 180 degrees the cam is rotating through. The oxygen sensor
used in the system is a wide
-
band sensor which gives a roughl
y linear response for v
oltage v
s.
air

to
-
fuel r
atio.


Engine Control Unit

The Engine
Control Unit
(ECU)
is implemented via a Microchip dsPIC 16
-
bit digital signal
controller

which
uses a combination of peak detection and lookup tables to determine the timing
and pulse width for the injector. In terms of timing, the ECU looks

for the lower

peak of the
Hall
Effect

sensor output, one of the unique points on its angular position response. From there, it
looks for the threshold value at which to start the injection which is specified by the programmer.
The threshold is typically about halfway

up the positive slope of the sine wave. The magnet used
for the cam angular position sensor is positioned such that the intake valve opens in this range.
The approximate linearity of the sensor output at this point provides a good basis by which to
vary

the timing of the fuel injection. The injection timing is determined by the current RPM of
the engine which is obtained by a timer triggered by the detection of the previously stated lower
peak.


In terms of pulse width, the system takes samples from t
he temperature and pressure sensors (at a
frequency of about 6 kHz). The system looks for the lowest pressure value observed over the
last cycle which usually occurs when the intake valve is about to close. Following injection of
the fuel, this value is
reset and the low value previously recorded is used to determine the pulse
width for the next injection. The lowest pressure value over the last cycle is used because the
pressure is not consistent between cycles. There is sufficient air leakage around th
e throttle
valve such that the pressure is equalized before the next injection takes place and a sensed value
of 1 atm prior to the next injection provides little information to the computer. Note that there is
a lag of one cam revolution between the pres
sure sampling and its use to calculate pulse width.
The frequency response of the temperature sensor is sufficiently low in the frequency band to
allow an instantaneous sample. However, the implementation of the heat exchanger may require
more detailed a
nalysis of the sensor output. The actual pulse width of the system is determined
by a look
-
up table value corresponding to a particular range of temperature and pressure. A
successive subtraction algorithm determines which range the sensor data lies in a
nd the look
-
up
table is traversed appropriately.


The feedback control of the system is implemented via a linear difference equation taking into
account a number of previous oxygen sensor samples. The output of this difference equation is

University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

14

either added o
r subtracted from the pulse width retrieved by the look
-
up table in order to tune the
next pulse width “guess”.


Power Electronics

Two DC
-
DC converters supply the power needed for the various components of the system.
Those components
,

which cannot be po
wered directly from the battery
,

utilize either a 5V power
supply or a 15V supply. More specifically, the 5V supply is used to power the computer and
sensors while the 15V supply is used for the fuel pump. Power supplies are designed around the
given 12V

1.4A
-
hour battery.


Tuning Notes

The initial look
-
up table was based on a series of assumptions that have seen been altered or
refuted. First, the assumption was made that the temperature at the intake valve at the time of
injection would also be or at l
east linearly related to the pressure inside the cylinder. This is not
true as the pressure gets continually lower from the throttle valve to the bottom of the cylinder on
the intake stroke. Also, the initial table failed to account for the increased tem
perature within the
cylinder which cuts down on the amount of available oxygen for combustion. Lastly, the table
did not take into account the exhaust gas left in the cylinder during the intake stroke which also
cuts down on the available oxygen.


With
these factors in mind, the engine ran extremely rich during our first tests. However, these
factors have now been compensated for. The compression ratio provided information on the
level of exhaust gases left in the cylinder during the intake cycle and t
he table was otherwise
altered through testing and tuning.



Experimental Projects

Flywheel

In order to maximize the thermodynamic efficiency of the engine, it should operate at full
throttle. In past competitions, this was achieved by accelerating the c
ar at full throttle to 27 mph,
a speed well above the required average minimum speed of 15 mph. The engine was then turned
off and the car slowed down to 7 mph, a speed lower than the 15 mph minimum. At this point,
the engine was restarted and run at ful
l throttle back up to above the minimum average speed and
the cycle was repeated. This technique, although necessary to maximize thermodynamic
efficiency, increases two main sources of energy loss: drag and engine starting. Since drag
increases proportio
nally to the square of velocity, exceeding the minimum speed significantly
increases drag, therefore severely increasing the amount of kinetic energy being lost. To
minimize this additional drag loss, the car could be accelerated to a speed that is not mu
ch over
the minimum, and then started again after less coasting in order to maintain the minimum
average. The problem with doing this is that the frequency of engine starting would be
increased, thus increasing energy losses associated with startup. Also,

the amount of time that the
engine is running per acceleration period was already very short, and it would be difficult for the

University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

15

driver to stay in control without error with even shorter acceleration periods.


For this year’s vehicle, the introduction of
a mechanical energy based hybrid drive train was
considered. This will help alleviate the problems associated with the burn
-
coast strategy. With a
flywheel coupled directly to the rear wheel of the car via a fixed ratio gearbox, the engine will
need to ru
n for a longer amount of time to increase the vehicle’s speed, since a large portion of
the energy will go to increasing the angular velocity of the flywheel. Also, because of this large
amount of momentum added by the flywheel, it will
take
a longer time

for the vehicle to lose
speed while coasting. This will allow the engine to run at full throttle for a longer time without
the increase in velocity that drastically increases drag. In addition, it will allow the vehicle to
coast for a longer amount of t
ime with a smaller reduction in vehicle speed, resulting in a
decrease in starting frequency which will decrease the losses associated with startup.
Implementing

the flywheel will also help to reduce the drag losses associated with the hill on the
course.

Without the flywheel, going down a hill adds extra speed that translates into higher drag
losses. With the flywheel, the energy gain associated with the decrease in elevation will be put
into rotating the flywheel along with the vehicle speed increase,
resulting in a lower speed
increase and therefore lower drag losses.


Figure
10



Flywheel Cross
-
section Profile


In the design of the flywheel, the most important goals were to maximize the energy storage to
weight ratio and minimize the amount of tota
l energy lost from the system without exceeding the
stress limits of the design. The amount of energy stored in the flywheel is proportional to the
radius to the 4
th

power and the weight is proportional to the square of the radius. Since this is
true, it

is advantageous to keep the radius as large as possible but within the limits of the design
of the body of the vehicle. The profile was then optimized to keep as much of the mass toward
the outer rim as possible in order to increase the energy to weight
ratio, since a mass close to the
axis of rotation stores less energy than a mass on the outer rim due to its lower velocity. This
does increase the stresses in the flywheel, but it was found through experimentation that the
energy losses associated with a
erodynamic drag and bearing friction on the flywheel cause the
most efficient operation speed. This speed is slow enough so that the stresses are small
which

allow
s

for

the use of

a thin inner cross section. The inner hub was left in order to add stabili
ty

University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

16

for when the system is subjected to gyroscopic effects resulting from cornering, and in the event
of failure, an aluminum burst shield was fabricated for protection. Unfortunately, due to the
extra kinetic energy in the system, braking distance would i
ncrease substantially. To remedy this
potential safety hazard, a clutch was connected to the brake line so that the flywheel
would

be
disconnected from the rear wheel in the event that braking is necessary. If a magnetic material
was used for the flywhee
l, a magnetic field could be created due to the magnetically sensitive
electronic equipment contained in the fuel injection system. Therefore, Monel, a Nickel
-
Copper
alloy, was chosen for the flywheel material because of the lack of magnetic properties as

well as
its high strength and density.


By implementing this mechanical energy storage system, some of the energy from the engine
will be able to be stored during full throttle acceleration and released at a later time while the
engine is not in use. This

will increase overall system efficiency by reducing drag losses as a
result of allowing a smaller speed differential while obtaining maximum engine efficiency
through use of the burn
-
coast strategy.


Heat Exchanger

With gre
ater fuel savings in mind, the
t
eam sought out to design a heat exchanger for the engine.
In theory, a heat exchanger would heat the air
entering the engine
in order to lower the density of
the air and the oxygen content. This provides for a leaner mixture of fuel and air, and therefor
e
results in the vehicle using less fuel. There is a power loss due to the new, leaner mixture but not
enough of a loss to cause a significant change. Thus, reintroducing the energy from the exhaust
in the form of heat to the engine will maintain a power

level close to what it would be if the
intake air was at ambient temperature.


Three concepts were created and then analyzed for the

design of the heat exchanger. These are
shown in Figure
11
.
The main concept employs concentric intake and exhaust pipes
. The
second idea was to maintain the intake and exhaust as separate parallel pipes with a carbon fiber
connection to transfer heat. The final concept, nicknamed “the crab
,
” utilized a reservoir for the
intake and exhaust gases to allow heat transfer to
occur over a longer period of time.






Figure
11

-

Heat Exchanger Concepts


Concentric, Parallel, and Crab


After an analytical analysis and simulation of the three conceived concepts, it was decided that
the concentric cylinder design would be the b
est choice. Design of this concept involved using
FIDAP simulation as well as Engineering Equation Solver programming. The simulations were
conducted using equations for counter flow heat exchangers and the flow of air through a pipe.
Tests were run on l
ast year’s engine to acquire the temperature of the exhaust gas leaving the

University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

17

engine via thermistors.


T
he theory behind the heat exchanger
was then

tested.
A
n experiment was performed in which
the intake pipe was heated using a blowtorch and the temperatur
e of the intake air was measured
by a thermistor close to the intake port as well as the temperature of the exhaust gas. This
experiment was run on a dynamometer and the amount of fuel used and the power of the engine
were carefully measured. These values

were compared against the stock engine power and fuel
consumption with no intake heating.


To acquire the best heat transfer for the prototype, the flow of the air was kept turbulent. To do
this, the inner and outer surfaces of the exhaust pipe were made

extremely rough. From a
previous analysis, it was found that to increase the heat transfer; the pipes should be made as
small as possible. However, with smaller pipes there is a greater chance of choking the engine.
A simulation was run to calculate th
e head loss of the system to determine if the engine would
choke or if the vacuum of the engine
would be

great enough to pull the air in and push it out.
The result of this test showed the major area of choking was the point where the exhaust pipe
entered

the intake pipe. The exhaust pipe blocked the majority of the flow cross section. As a
result, an expansion chamber was instituted to relieve restriction in area. The appropriate size of
the expansion chamber wa
s found using FIDAP. Figure
12

shows the

completed prototype heat
exchanger.


Figure
12

-

Prototype Heat Exchanger


Powertrain Configuration


The drive train configuration significantly influences the fuel efficiency of the vehicle
. This
year’s SMV team sought to perform testing to compare a

single speed direct drive power train to
a continuously variable transmission (CVT) power train. In a direct drive system, the gear ratio
is fixed, so as the engine shaft speed increases, so does the output power and naturally, the fuel
consumption. The

CVT transmission is composed of two conical pulleys. As the engine speed
increases, one of the pulleys remains fixed while the other changes the size of its “cone” through
rotational forces and springs. The idea of the CVT is that unlike direct drive, a
utomatic, or
manual transmissions which have a set number of gear ratios, the CVT has infinitely adjustable
gear ratios within a given range but is constrained by the size of the pulleys being used. This
continuous change in ratio automatically optimizes
engine efficiency.


University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

18


The biggest concern with using the CVT was that a fair amount of belt slippage was occurring on
the pulleys using the recommended belt tension as well as a tighter tension. This led to the
testing of both CVT and direct drive transmiss
ions using a dynamometer to determine the most
efficient system. Unfortunately, due to various e
ngine and problems with the dynamometer
,
relevant data has yet to be obtained to accurately compare and support a decision for choosing
the most effective driv
e train. Although testing has not been completed at this time, the drive
trains have been initially designed such that when a decision on which system to use is made, it
can be implemented immediately onto the vehicle.


For a direct drive system, it was c
alculated that a gear reduction of 6:1 would be required to
obtain a maximum vehicle speed limit of 30mph. This system would be composed of a 2
-
inch
diameter pinion on the crank shaft and a 12
-
inch diameter gear at the rear wheel. The gears
would be conn
ected by a heavy duty chain such as those used on motorcycles. A friction clutch
system would be integrated at the rear wheel so that the engine could be disengaged when not
being fired, allowing the vehicle to coast.


The CVT system, which was used in l
ast year

s
vehicle
, would incorporate a belt that rotates
between two V
-
shaped pulleys. The pulley at the crank shaft is rigid while the pulley at the
output is spring loaded allowing it to move axially to vary the radius of the contact point between
the
two pulleys which would change the tangential velocity. The CVT would allow the engine to
operate at its optimum RPM regardless of the vehicle’s speed. A clutch is not needed because
when the engine is off the belt is not tensioned and will just slide ov
er the pulleys.




Brake System

and Vehicle Performance


The braking system on the vehicle consist
ed

of a single disc brake mounted to the rear wheel of
the car. The brake is an Avid mechanical disc brake
setup
commonly found on high
-
end
mountain bikes an
d it features a 6
-
inch rotor and a single piston caliper.

The caliper
is

cable
-
actuated by way of a brake lev
er attached to the handlebars.


The stopping distance of the vehicle was determined assuming friction coefficient values and
that braking force wa
s equal to traction force. Estimates were made concerning vehicle mass and
normal forces acting on the braking wheel. The maximum speed of the vehicle (assuming drag
limitation) was estimated to be approximately 60 mph
.

H
owever, the max speed in competiti
on
trim (assuming gear/engine speed limitation) was estimated to be 40 mph (17.9 m/s). The
stopping distance from
a speed of
40 mph was estimated to be 90.2 ft.


This distance was
determined based on the kinetic energy of the vehicle traveling at
that spee
d of 40 mph
.


Stopping distance was also estimated for the braking section of the maneuverability course.


From a speed of 11.2 mph
, more than two times the minimum speed required for the slalom
section, 6.9 feet will be required to stop the vehicle. This
distance is well within the maximum
stopping distance required.


Running Gear


University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

19


The vehicle’s running gear wa
s made primarily
of
bicycle components and custom
fabricated
items.

Due to the
el
ev
ated

cost
for high performance

vehicle wheels,
the wheels

were r
eused
from the previous years’ vehicle and adjusted to fit the current vehicle’s
configuration
.

These
wheels
were

only a year old and still appeared to perform as well as
they did
when new
.

To
prove that the wheels would still perform well for this year’
s vehicle, rolling resistance tests were
conducted
.
The tests

resulted in the wheels having a
sufficiently
low coefficient of rolling
resistance
which would

provid
e

optimal performance for the vehicle.


All three wheels were built using standard tires fit
ted to heavy
-
duty mountain bike hubs which
were mounted to standard 20” rims. The two front wheel hubs made use of a 20
-
mm through
-
axle design which allowed for the fabrication of custom precision aluminum steering knuckles
and axles as shown in Figure
13
. Figure
14

below illustrates how the wheel hub mounts to the
knuckle.




Figure
13

-

Aluminum Steering Knuckle


Fi
gure
14

-

Steering knuckle with wheel hub


(shown in

black).


The rear wheel uses the

same rim and tire as the front whee
ls

along with

a heavy
-
duty rear hub
,
as
opposed to a mountain bike hub.
The rear hub also accepts I
nternational
S
tandard

(IS) bicycle
disc brakes. As discussed
in the Brake System section
, a mechanically actuated bicycle disc
brake
was

fitted to the vehic
le to satisfy the braking requirements.

The hub
wa
s also a through
-
axle design

with a 12
-
mm axle. The design is shown below in Figure
15
.



University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

20


Figure
15

-

Model of rear wheel through
-
axle.

The final drive gear of the vehicle’s powertrain
was m
ounted to a

freewheel
which wa
s
attached
to the rear wheel hub. This freewheel allows the vehicle to coast without
running the engine.



The steering apparatus was chosen to be a rack and pinion design. Using the rack and pinion set
-
up minimize
d

backlash
in the as
sembly
as well as guarantee
d

a precise turning
feel for the driver
.

This setu
p is depicted below in Figure
16

and provides the vehicle with a turning radius well
under the required 50
-
ft turning radius for the competition.



Figure 16

-

Rack and Pinion S
teering


Body/Aerodynamic Shell

Aerodynamic Considerations

T
o minimize drag forces, a number of factors were kept in mind during the design process of the
shell. The general idea was to minimize the cross
-
sectional area, avoid sharp corners or negative
curvatures (concavity), and to keep the air flow laminar around the body for as long as possible.
Of course, the design was also constrained by the placements and sizes of the internal
components as well as the driver. Ultimately, wind tunnel testing ver
ified the theories behind the

University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

21

designs and allowed refinements of the shell when necessary.


This year
,

it was decided to have the shell enclose
all of
the wheels instead of leaving the two
front wheels exposed. The team used bicycles wheels from a road bi
ke because of
their

minimal
rolling resistance. However, like many typical bicycle wheels, the spokes on the wheels
contributed to a portion of the drag resistance since it behaves like a cylinder moving through
fluid. Therefore, each spoke created its o
wn wake and the two front wheels would have caused
the airflow to begin separating at the front end of the vehicle. Initially, covering the spokes on
each wheel was considered but the wheel itself would create wakes since its shape was still
cylindrical (
or a thin disk). Also, with the front wheels being outside of the shell, a majority of
the steering assembly would also be exposed and contribute to drag effects.


The placement of the wheels within the body provided benefits as well as drawbacks. For
in
stance, the cross
-
sectional area of the shell increased to account for the two front wheels,
thereby, increasing the drag force linearly. Still, it was predicted that the benefit from keeping
the airflow around the body laminar was more significant, espec
ially in the front end of the
vehicle.


The sides of the vehicle were kept relatively flat to minimize the overall cross
-
sectional area.
This design was especially important for the front end of the vehicle, where the front wheels are
placed. Since the i
nclusion of the front wheels made this portion of the body widest, it was
essential to keep from making the cross
-
sectional area any wider. This design w
ould

be
disadvantageous when the vehicle experiences crosswinds. The main objective was to design the

vehicle to travel forward and not sideways; it was not optimally designed to handle crosswinds.
This reasoning relies on the assumption that the relative wind speed in the longitudinal direction
is insignificant to the relative wind speed in the axial di
rection. Furthermore, if the sides were
protruded to create a rounded bulge, it could potentially cause lift or drag force in the
longitudinal direction. This is analogous to an airplane wing.


Boundary layer separation occurs when there are sharp corner
s. Therefore, a gradual reduction
in cross
-
sectional area for the rear half of the vehicle was required to make the boundary layer
stick to the shell. This prevents wakes from forming prematurely; thereby, reducing drag.
However, if the area reduction r
ate is too slow, then the shell would be impractically long. The
increase in surface area would also increase viscous drag. Based on previous experience and
knowledge, it was decided that a 7 degree taper in the back end of the vehicle is the ideal
compr
omise.


A port on the rear face of the vehicle was cut out for the exhaust pipe. By using this placement,
unnecessary protrusions sticking from the side of the shell were avoided. Otherwise, the
protrusion would have caused the flow of air to separate pr
ematurely from the body’s surface.


For safety reasons, kill switches were placed on either side of the vehicle’s outer surface. The
lever arms of the toggle switches were cut to the minimum length of 0.75 inches and were
sanded down to a smaller diameter
. Again, the intention was to prevent the external components
from tripping the boundary layer flow.



University of Massachusetts


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22

Safety Considerations

Above all else, safety is the first and last consideration of the body design. Specifically in the
case of an emergency, the dri
ver needs to be able to get out of the vehicle on his/her own ability,
the pit crew needs to be able to get the driver out if the driver w
as

unconscious, and the fuel
bottle needs to be accessed within their corresponding time limit. Another fundamental c
oncern
was that the vehicle needed to be securely fastened to the frame or other pieces of the shell and
that the body would be sufficiently stiff. The last thing the driver should be worried about is that
pieces of the shell m
ight

fall off. The shell co
nsist
s

of three separate pieces: tail section, bottom
section, and front section. On the edges of each piece, a flexible L
-
bracket was attached to create
a surface that was perpendicular to the tangent at a connecting edge
, shown in Figure
17
. This
desig
n allowed the adjustment of contact surface area. The amount of Velcro area used is
directly related to the amount of force required to separate the adjacent contacts. It was then
possible to define the amount of Velcro used to ensure that the driver sep
arated the adjacent
contacts with ease (in case of emergency) while preventing separation during normal ride
experience. Furthermore, the stiffness of the shell was increased by embedding several long
strips of sheet metal into the underside of the fiberg
lass.



Figure 17

-

Cross sectional view of how sections of the body are joined together


The tail section extended from the end of the vehicle to the firewall/roll
-
bar (length) and from the
top of the roll bar to the chassis (height). Since this section

was not required to be opened in a
short time, it was designed to hold in place more securely. An access port for refueling the fuel
bottle was designed for quick accessibility and minimum drag. Since the number of external
protrusions on the shell w
ere

to be minimized for aerodynamic efficiency, a handle
-
less design
was required. The solution was to partially cut a circle into the fiberglass shell, while meeting
the 300cm
2
minimum area, near the vicinity of the fuel bottle.


As shown in
Figure 18
, the
vertical slit,
which

remain
ed

attached to the fiberglass shell, face
d

the
front of the vehicle. By doing so, when air moved from the front to the rear of the vehicle, the
“flap” (fuel port) of fiberglass had a tendency to remain flush to the rest of the b
ody. If the
orientation were flipped, the “flap” would potentially scoop air and cause itself to open, which

University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

23

significantly increases drag. As a secondary measure, the fuel port utilizes magnets to keep the
door closed during the race and quick access whe
n refueling. With both measures at hand, a
reduction in drag and compliance to the competition rules were achieved.

Next, the bottom
section of the shell consists of the entire undercarriage, beginning from the chassis, and the front
nose of the vehicle,

including the region over the front wheels.




Figure 18
-

Fuel Port Design


The undercarriage separates the driver from the pavement, therefore, acting as a safety barrier.
This section was securely fastened to the chassis by means of Velcro but was no
t intended to be
easily removable. By avoiding permanent fasteners, there is significantly more flexibility in
future repairs on the vehicle. The use of certain fasteners, such as rivets, may extend beyond the
outer surface of the shell. Again, this may

increase the drag force.


Finally, the front section of the shell consist
s

of the driver’s entrance hatch. Unlike other
sections of the shell, the hatch was fastened to the chassis and other sections with several
magnets. This design allow
s

the driver t
o open the hatch with little force and time. Similarly,
members of the team
a
re able to open the hatch from the outside effortlessly. A section was cut
out and replaced with a plastic window that gave the driver sufficient visibility.

The final body
des
ign is shown in Figure
19
.



University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

24


Figure
19



A Pro/Engineer model of the Body shell of the vehicle.



Cost Estimate and Manufacturing Methods

Body Team

The total cost accrued was $1541 which included materials to construct the frame, tools to shape
the body,

fiberglass for the shell, and other miscellaneous parts. These costs are shown in Table
1

in Appendix B. There were significant donations of $1085 from NE Foam & Sealed Air and

WR Sharples as shown in Table 1. These companies provided the foam and wate
r jet cutting
services, respectively, which were used to manufacture the foam mold used to produce the shell.


The fiberglass body was produced by creating a male mold from sections of foam specifically
cut to form the shape of the body by WR Sharples. Th
e assembled rough foam mold was then
shaped using plaster of paris and then sanded smooth to produce the final body shape. The
fiberglass shell was then constructed by layering multiple sections of fiberglass over the mold
and applying the epoxy and harde
ner. By adding 0.75
-
in rubber tubing to the inner face of the
fiberglass sections, the rigidity of the body was increased substantially.


The chassis was constructed using TIG
-
welded 6000 series Aluminum rectangular tubing and the
windshield was construct
ed using a 0.005
-
in thick piece of Lexan sheeting.


Engine Team

The engine team required the most amoun
t of funding for this project. The e
xpenses

for the team
are

shown in
Table
2

in
Appendix B and total $3489.


There w
as

$2320 in donations
to the
engi
ne team
which

included a
generous contribution from JE Pistons in the form of a
billet
cylinder

of
2618

aluminum
from
which the piston was machined.


The piston, connecting rod, and cams were designed using Pro
-
Engineer and Pro
-
Mechanica and
were speciall
y machined on a Bridgeport CNC mill. The cylinder sleeve was made from cast
iron pressed into an aluminum sleeve that was pressed into the
existing

Briggs
and Stratton

University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

25

crankcase. The engine block was heated
to expand the cylinder diameter

and the sleeve
was then
press
-
fitted into the block. This formed a tight seal as th
e block cooled and contracted. The
cylinder head was manually machined using a Bridgeport milling machine and the valves were
machined on a lathe.


Fuel Injection Team

The fuel injection

team purchased $661 in parts which included various sensors, circuitry, and
pumps

which can be seen in Table 3
. They
received $668 in part donations, as shown in Table
4
,

which included the fuel injector, printed circuit boards, sensors, and driver chips
. The team has
their own lab where computers, oscilloscopes, solder irons, and other various tools are available
to them.


Total Donations

O
verall, the entire project has cost $6007. Total donations are shown in Table 4 and stand at
$13,073. It was also

estimated the machine shop staff labor totaled approximately $10,000.
Outside of part and material donations, a substantial amount of cash has been donated to the
entire team from supporters such as TI, General Dynamics, SAE, and the UMass Commonwealth
C
ollege.
Miscellaneous expenses, shown in Table 5, result in an overall spending of $304 for
registration and sponsor acknowledgement and appreciation.
All expenses and donations are
stated to date and are expected to increase as the project wraps up towa
rds June.

























University of Massachusetts


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Department of Mechanical &

Industrial Engineering

26

Appendix A


INTAKE
Intake
Gas
Temp
Sensor
Manifold
Absolute
Pressure
Sensor
Fuel
Injector
EXHAUST
Wideband
Oxygen
Sensor
ENGINE
Cam
Angular
Position
Sensor
Combustion
Process
ENGINE CONTROLLER
Controller
Algorithm
DRIVER INTERFACE
DATA RECORDER I
/
O
Bus Arbiter
A
/
D
Conversion
Micro
Controller
PC Serial
Interface
PC
STORAGE
Memory
Controller
Memory
PC Serial
Interface
POWER
ELECTRONICS
Battery
12
V
Fuel
Pump
Sensor Bus
UART PC Serial
Connection
Memory Bus Control
Data
Atomized
Fuel
Exhaust
Gases
Injector
Pulse
5
V
Power
Supply
Fuel
5
V
12
V
For Simplicity Here
,
we
have noted the need for
5
and
12
Volt power nodes
.
A component with a
5
V or
12
V in the lower right
corner uses
5
or
12
volts
respectively
.
In cases
where we do not yet know
the power requirement
,
we
have simply used a P
.
5
V
5
V
5
V
5
V
P
P
12
V
12
V
5
V
Starter
12
V
Ignition
Control
Starter
/
Ignition
Control
12
V
System Block Diagram
:
SMV Electronic Fuel Injection
LEGEND
5
V
P
12
V
Control Signal
Information Signal
Sub
-
Component
Top
-
Level Component
Requires
12
V Power
Requires
5
V Power
Requires Unknown Power
The Sensor Bus is the pathway by
which information will be transferred
from the sensors to both the Motor
Controller as well as the Recording
Setup
.
A
/
D conversion may be required
at points we have yet to determine
.
The Data Recorder records
information from selected sensors on
the vehicle and stores them in array
form in our storage system
.
This
data is available for download via the
PC serial connection for proper
graphing and display
.
The Motor Controller is the means by
which we control and time injector
pulse width
.
Feedback A
/
F
,
cam
angle
,
and Manifold Pressure weigh
in strongly along with a lookup table
built from other sensor data to
determine injected fuel volume
.
Description
The Intake
,
Motor
,
and Exhaust are
all part of a Briggs
and Stratton
single
-
cylinder
4
-
cycle motor used
in SMV
competition
.
EFI Design
Team
:
Jordan Crafts
Mike Flanagan
Dorinda Garcia
Jim MacDonald
Andy McCaron




















University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

27

Appendix B


Table 1

Body Team Expenses

Part

Cost

Bike Tire Pump

$
51.15

Stop Watches

$
34.80

Defogger

$
15.21

Heat gun

$
32.99

Frame Material

$
271.09

Windshi
eld Material

$
87.61

Steering Rod Material

$
33.32

Brake Lights

$
47.23

Fiberglass and Hardener

$
572.66

Aluminum Tubing

$
114.22

Hardware

$
87.34

Ball Joints

$
27.50

Sanding Material

$
118.47

Total

$
1493.59


Table 2

Engine Team Expenses

Part

Cost

USB

Cable

$
41.99

Steel Rod

$
26.11

Flywheel & Elec. Starter

$
134.62

Starter Gears

$
18.69

Tools

$98.38

Thread Hand Tapper

$
26.63

Engine Bearings

$
82.45

Water Dyno Repair

$
478.05

Engine Gears

$
41.61

Bearing Pinion

$
11.14

Engine Parts

$
26.76

Roller Lif
ter

$
29.51

Fuel Pump

$
70.97

Belts

$
31.19

Starter Drive

$
15.69

Steel Tube

$
31.91

Bearings & Parts

$
147.68

Cutting Tools

$
40.28

Dual Valve Springs

$
28.56

Gear Pinion

$
10.61

Rods & Studs

$
94.01

Ignition Magnets

$
47

Copper Sheet

$
71.61

Piston Rings

$
62.12

Iso
-
Octane

$
42.11

Electric Dyno Costs

$
1598.32

Sensors

$
180.87

Total

$
3488.87


University of Massachusetts


Amherst


Department of Mechanical &

Industrial Engineering

28


Table 3

Electronic Fuel Injection Expenses

Part

Cost

Fuel Pressure Regulator

$
140

Fuel Pump

$
94

DigiKey Parts

$
225

Temperature Sensor

$
172

Voltage Regulat
or Chips

$
30

Total

$661







































Table 4

Donations





Company

Item/Service

Amount/Approx.

Bergfelt Racing

Flywheel, Carb, Parts

$150

Competitive Edge

Bike Handle Grips

$35

Commonwealth College

Money

$3,900

JE Pistons

Cast Aluminum

$100

JE Pistons

Pro
mo Stickers/Banner

$70

SAE

Money

$1,000

WR Sharples

Water Jet for Foam Cut

$750

NE Foam & Sealed Air

Foam Material

$300

Texas Instruments

Money

$6,000

General Dynamics

Money

$250

American Sensor Tech.

Pressure Sensor

$18
0

Allegro Micro

Hall S
ensors

$9

Arnold Magnets

Polarized Magnets

$18

Misc.

Boards and Chips

$461

Total


$13,073

Table 5 Miscellaneous Expenses

Part

Cost

SAE Registration

$
250

Sponsor Team Photos

$
22.72

Sponsor Frames & Packages

$
32.02

Total

$304.74