Design Project
Formula SAE Drivetrain
Submitted to
Neal Birchfield
MME 213: Computational Methods for Engineering
Miami University
by
Marissa Shaffer
James Elias
Bob Mueller
Kalene Kelly
May 15, 2014
In order to begin thinking about a design that will transfer torque from the differential to
the wheel of the car, it is important to understand the overall drivetrain of a FSAE car. The basic
components include a differential, axles, and hubs as shown in the Figure 1 below:
Figure 1: This figure shows the various components of a FSAE drivetrain (Honeychuck)
The drivetrain of the FSAE car is responsible for transferring the torque produced by the
engine to the wheels in order to make the car go.The torque that is produced by the engine is
transferred to the differential. The purpose of the differential is to transfer the available power to
the slowest spinning axel. This allows the car to gain better traction. The differential is purely a
mechanical design which responds extremely quickly to slippage with minimal energy loss.
From the axles, the power is then transferred to the wheel (Schwartz). This report will discuss
potential design options for the drivetrain along with some ideas for optimizing the performance,
strength and angle of torque projection. Additionally, one of the most important goals for the
final design is to make sure that the drivetrain system transfers the torque from the engine to the
wheels with minimal energy loss.
Drivetrain mechanical failure background
While designing the drivetrain, it is extremely important for the system to be as light as
possible in order to maximize the power to weight ratio along with maximizing the strength of the
material to prevent failure. By increasing the stiffness of the shaft, the angle of torsion will
decrease, which will increase the direct flow of power to the wheel with minimal energy loss. By
making the system more rigid, the shear stress experienced by the driveshaft will increase,
however the performance of the car will increase. Minimizing the impulse thats being delivered
from the transmission to the differential will reduce shear stress within the differential thus
reducing the shear stress within the driveshaft. This can be done by adding a damping pad
along a linear shaft. Typically this dampening pad will be located between the transmission and
differential if the shaft is horizontal with the two systems. Failure within the driveshaft joint is
directly related with the angle of projection. As the angle increases you increase shear stress
which subjects it to failure. Failure typically occurs within the sleeve of the ball joints, which is
the thinnest part of the design (Common Driveline Issues). Figure 2 shows a Universal Joint
that has cracked due to large amounts of internal shear stress. Most designs include a sleeve
socket for the joint to fit in. The purpose of the sleeve is to allow the drive shaft axial
displacement to change as the wheel vertically changes. Failure could also occur if the joint
slips out of the sleeve, however this rarely happens unless a collision occurs.
Figure 2: Figure 2 shows a U-joint that has failed due to high shear stress (Cracked Axle Ear)
Suggested Designs
After research was conducted, it was found that the most important aspect of a drivetrain
design is finding an efficient way to transfer the torque from the engine to the tires of the car. It
is also imperative to limit the shear stress in the drivetrain, especially in the joints that are
extremely susceptible to the effects of high shear stress. In order to reduce the stresses
experienced by the joint and to protect the performance of the joint, different designs are
presented in order to achieve the desired output. Therefore, all of the proposed design ideas will
include a dampening unit to reduce shear stress that will be located behind the differential and
transmission. The differential will ensure that both axles are always rotating at the same speed
relative to one another. The dampening unit will reduce the impulse which will reduce the shear
forces felt by the differential and along with the remaining system.
Ball Bearing Joint
A joint that allows the drivetrain to transfer torque from the engine to the wheels while
still allowing the wheels a certain degree of vertical movement . An optimal design of a CV joint
is lightweight and simple.
The first joint design proposed is a ball bearing type CV joint. This joint consists of an
outer socket and an inner joint connected by a series of ball bearings. The ball bearings allow
for the inner socket to be able to translate slightly in all directions as well sliding in the sleeve
along the horizontal direction to allow for this motion. Figure 3 shows a ball bearing joint.
Figure 3: Figure 3 shows a ball bearing type CV joint (2 J’s Automotive)
CV Tripod-type joint
A second possible solution to the problem we face is a tripod-type CV joint. This joint
acts much like the ball bearing CV joint, but instead of ball bearings transferring the torsional
force between the outer socket and inner joint, there are 3 bigger bearings located at 120
degrees from each other. This design is more practical and durable than the ball-bearing type
CV joint, however it gives a less angle of projection. Figure 4 describes a cut out view of a
tripod joint.
Figure 4: Figure 4 describes a Tripod type CV joint (Nasioc Forums)
This design includes the common sleeve where the tripod bearings are able to adjust
along the axial alignment depending on the position of the wheel. The purpose for this design is
to include the projection of angle in all directions. The components within this consist of the
spider, the spherical roller, retainer, needle roller and the rings. This is commonly used for front
wheel drive cars where you're applying torque while the wheel’s direction is not aligned with the
direction of the car.
Figure 5: Figure 5 shows the internal pieces of three types of Tripod joints (CV Joint Blog)
U-Joint
This design may be the most optimal design for the driveline. Since the wheels of the car
are only moving in the vertical direction while staying aligned with the car, there is no rotational
movement about the vertical axis. The join consists of a sleeve for displacement change just like
the previous joints however this joint is more durable and simple. Failure does typically occur at
the thinnest point of this joint which we can reduce the change by upsizing the joint, and by
decreasing the impulse through the system.
Figure 6: This is a U-joint typical of FSAE drivetrains (CV Joint Blog)
Decision Matrix
The criteria below will be put into a selection matrix along with the three re-designs in
order to help choose a design that will be the best for the project. Each section will be weighted
from least importance to most importance to help aid in the search of a best design. Without
choosing exact number rankings for the criteria, the performance will most likely be weighted
the highest along with strength/durability. Another important aspect to consider is the mass.
Next should come cost and the least important thing to consider should most likely be the
appearance of the design. Reasoning for the ranking can be determined by the definitions
provided.
Mass - the mass of the car should be as light as possible. Looking at the equation for the
conservation of energy, kinetic energy depends on the weight of the object in motion. The lower
the mass, the more acceleration there will be with a given amount of force coming from the
engine.
Strength of the material used/ Durability - the material used needs to be strong and capable of
withstanding the shear stress that will be experienced during the race. The material should also
be long lasting and able to withstand multiple races. With a weak material, the drive train could
fall apart and cause a rippleffect of destruction.
Cost - in any real-life scenario, the least amount of money spent is best. A design team may be
on a certain budget, or has the goal of spending as little of their companies' money as possible.
Aesthetics - the physical appearance may not hold as important in the selection matrix, but it
does hold some level of importance. The first impression of a design is judged by what a person
can see. It is possible to have a great performing design that is aesthetically not pleasing. This
may ruin the credibility of your design to others.
Performance - the way a design performs will really say it all. A design may appear to be well
designed, but you don't know how great it is until it is put to the test. The performance of the car
will determine how it stands out against the rest of the cars in the race.
Analysis Plan
The main point of failure the design needs to focus on solving is reducing the impulse
torsional forces felt by the components of the drivetrain. Sudden braking, acceleration, and
collision all send impulse torsional forces through our drivetrain that can damage components
and affect the performance of the race car. Because the causes of torsional forces cannot be
avoided in racing, the final design must include a plan to reduce their effects on the drivetrain to
avoid component failure. One way this can be done is installing a dampening unit along the
drivetrain to absorb some of the shear impulses and reduce their effects on the drivetrain
components. Another failure can be avoided is choosing a material for the driveshaft that can
absorb some of this impulse while still being strong and rigid enough to efficiently transfer the
torque from the engine to the wheels. The CV joint used can be upsized to increase their overall
strength by using more material. Also, the joint strength is a function of its bending angle,
therefore if the angle of projection can be reduced, the shear forces felt within our CV joint can
be reduced. All of these things will reduce the efficiency of the torque transfer to the wheels and
add weight to the car, but this is a necessary tradeoff to ensure that the final design does not fail
during operation.
Proposed Designs:
Design 1
This design involves a hollow spline shaft made from heat treated precipitate
strengthened aluminum. This shaft will connect a steel U-joint and a steel tripod joint to a hub
interface plate that will match the bolt pattern to the wheel and differential. Starting at the
differential of the driveline, a 1/4’’ aluminum interface that is part of the U-joint is bolted to the
inboard spline shaft. The length of the U-joint will be 50.8 mm. The other end of this U-joint is
splined and the teeth of this joint are then intermeshed with the teeth of the hollow aluminum
spline shaft. Figure 7 shows how the spline shafts included in the design will be able to transfer
torque through the components. This specific aluminum grade has a relatively low shear
modulus compared to steel, which means it cannot absorb more shear without plastically
deforming (MatWeb); however, the weight difference is great. The shaft design was able to be
hollow and maintain the required allowable shear stress which is still less than the maximum
allowable shear stress given from the material. The twist that occurs in the shaft is very minimal
therefore the power will be transferred quickly, improving the performance. The right side drive
shaft length is 393.7 mm with a radius of 50.8 mm and a thickness of 10.3 mm. We included a
factor of safety of 2 to the thickness. These values were calculated from the equations in
Appendix B. The left side drive shaft’s radius and thickness is the same as the right shaft,
however the length will be 292.1 mm. The outer end of this aluminum shaft is splined as well
and is connected with a steel tripod joint. This tripod joint will allow for vertical movement while
also allowing for the horizontal movement of +/- .25 inches as required. The length of this tripod
joint will be 76.2 mm. The outgoing shaft of this tripod joint will be splined and connect to the
hub interface, which is made from 1/4’’ thick aluminum. The other end of the tripod joint will be
splined and fit into the hollow aluminum splined shaft. The purpose of this hub interface is to
connect the Tripod joint to the hub by utilizing the given bolt pattern .The hub interface will be
directly bolted to the hub and splined to the tripod joint in order to transfer the required torque to
the wheels. The purpose for the tripod joint is to utilize its additional function of the sleeve.
Figure 7: This figure shows how two spline shafts can be interconnected by the meshing
of the teeth of each shaft. This is how torque can be transferred along the driveline. (Automotive
Engines)
Design 2
Design 2 also requires the use of the aluminum spline shaft with the radius and
thickness matching the shaft in design 1, but with a different length. This design uses two steel
U-joints and aluminum hub interfaces. Figure 8 shows a system of two U-joints and the sleeve
design between the shaft and one U joint. Design 2 will allow telescopic motion on one end of
the hollow shaft and fixed on the other. Starting at the differential, a 1/4’’ thick aluminum
interface that will match the bolt design will be connected to a U-joint. This U-joint will have a
length of 50.8 mm a radius of 25.4 mm and will be made of steel since most failures occur at the
joint. From this U-joint, the aluminum shaft will be tightly splined to the other U-joint, allowing no
movement. After calculating the applied shear stress on the drive shaft from the maximum
torque from the wheels, the shaft’s material is going to be aluminum which has a reduced
weight compared to steel. The right side drive shaft length will be 419.1 mm with a radius of
25.4 mm and a thickness of 10.3 mm. The left side drive shaft length will be 317.5 mm with the
same radius and thickness.
The thickness was developed from the shear stress and moment of inertia equations which can
be found in Appendix B. Since this shaft is longer than the design 1 drive shaft, it will be able to
absorb more shear stress since the angle of twist increased slightly; however, there will be a
delay of transferring power which will dampen the acceleration. The outer end of this aluminum
shaft will be splined as well and interconnected with a two inch long steel U-joint. This U joints
splined shaft that will be connected to the drive shaft will hold a looser fit than the other side in
order to allow a change of length as the wheel changes position. The U- joints will be
experiencing a lot of shear stress, therefore the material they are fabricated out of must be
strong and durable to prevent failure. The second U-joint is in turn connected to a 1/4’’
aluminum hub interface which is bolted to the hub, as described in Design 1.This horizontal
movement is provided by the splines that connect each component of the driveline. The
interlocking teeth will allow each part to move horizontally as needed, as long as they remain
well lubricated and within its own range.
Figure 8: This figure shows a system of two U-joints connected by a spline shaft (Supplier &
Distributor)
Design 3
The final design is composed of two steel tripod joint connected to the differential
by an aluminum interface. Tripod joints allow for both horizontal and vertical motion,
which will meet the constraints of the driveline. Connected to the tripod joint is a long
splined aluminum shaft that is hollow. The right drive shaft length is 368.3 mm and the
left side’s length is 266.7 mm. The thickness will be 10.3 mm and the radius is 25.4mm.
The hollow shaft is extremely lightweight and takes less rotational energy to make it
spin, maximizing the torque transfer through the shaft. The hollow shaft is then
connected to another steel tripod joint. The outboard shaft of this joint is connected to
an aluminum hub interface, which can be bolted onto the hub to complete the driveline.
The Selection Matrix
Taking the criteria described above, a weight for each category was added to ensure the
optimal design was chosen. Aspects of the design that improved the functionality, such as
weight, strength and performance were weighed very heavily. Other features such as cost and
aesthetics were not important to the overall performance of the driveline and received a low
weight rating. Each design received a score for each category of ranking, three being the best
and one being the lowest possible score. Additionally, each of these categories carried a weight
of importance, with all of the weights summing to 1. The scores of each design received for
each category was then multiplied by that categories weight, and then all the scores for a given
design were summed to give a final score. The most optimal design from the three proposals
will be the one with the highest score. Table 1 below shows the decision matrix as well as the
scores received by each design. When analyzing the data it is inferred that design 2 is the
optimal design based off of the listed criteria.
Criteria
(weight)
Design Mass (.2)
1
2
2
3
3
1
Cost (.08)
Aesthetics (.02)
2
3
1
Strength/Durability Performance
(.3)
(.4)
Overall Score
2
2
3
2.4
1
3
2
2.56
3
1
1
1.04
Table 1: The table above shows the scores received by each design for every category
considered as well as the Overall Score each design achieved.
Analysis Scoring of Designs
All values and calculations for the analysis of the selection criteria can be found in Appendix C.
Mass:
For each design the total mass was calculated by finding the mass of each part included
in the design. The mass of the U-joints and tripod joints were found through research (Supplier
and Distributor). The mass of each shaft along with the interfaces was found by multiplying the
density of the material with the volume. Once the mass of each individual piece was found, all
the parts were added together. The total mass for design one is 5.28kg, design two is 4.99kg
and design three is 5.57kg.
Cost:
The cost of each design was determined through research. The U-joints and tripod joints
cost were found (Supplier and Distributor) and then the cost for an aluminum hollow rod with the
desired diameter and thickness was found (MetalsDepot) with a length of the sum of all the
shafts. A proportion was then used to find the cost of the shaft for each design. The total cost for
design one is $145.84, design two is $99.24 and design three is $192.45.
Aesthetics:
Aesthetics is best defined as how visually appealing the design is to someone. Tripod
joints are more contained than U-joints and therefore were deemed to be more aesthetically
pleasing. Because of this, design 3 scored highest in aesthetics because it contained two tripod
joints. Design 2 scored the lowest in this category because it was made up of two “ugly” looking
U-joints.
Strength/Durability:
In order to rate the strength/durability the shear stress on the shafts and joints was
calculated using τallowable equation found in appendix B. Since the only difference between the
shafts is the length the shear stresses on all the shafts were the same. This shear stress value
was compared to the maximum shear stress value to make sure that it did not exceed it. The
allowable shear stress in the U-joints was much greater than the allowable shear stress of the
tripod joint therefore the design with two U-joints got the highest rating.
Performance:
Performance was evaluated based off of two criteria, the angle of twist that the shaft and
joints would be subjected to as well as the ease of horizontal movement of the driveline, as
required in the constraints. It was found that U-joints have less of an angle of twist when
subjected to torque, increasing the performance of the driveline. However, tripod joints allowed
for horizontal movement more easily. A greater emphasis was placed on the angle of twist,
because maximum torque transfer is the overall goal of the design. For these reasons, design 1
was ranked the highest due to its low angle of twist as well as its allowance of horizontal motion.
Design 3 was ranked lowest because although it allowed for easy horizontal movement, its high
angle of twist would greatly decrease the maximum torque transfer for this design.
Summary of Decision Matrix
It was found that the design composed of a hollow aluminum shaft, two aluminum
interfaces, and two U-joints was the optimal design for this application based off of the criteria of
mass, cost, aesthetics, strength, and performance. The low mass of this design along with its
high durability and strength outweighed its visual flaws as well as its lacking in performance.
Due to this designs light weight, it will take less kinetic and rotational energy to move and rotate,
which increases the amount of energy it can transfer to the wheels of the car. The high durability
of this design means that the driveline can withstand high amounts of shear stress without
plastically deforming or breaking.
In the future, the design chosen has to be more fully analyzed. The calculations
performed during this phase were mostly based off of gross estimations and simplifications that
will not be adequate enough to base the final design off of. A three dimensional drawing will be
created for all parts of this design using a drafting software. From this model, full analysis can
be run to seen what areas will concentrate the most stress during operation and what parts, if
any, need to be altered to handle the stresses the part will experience.
Final Design
The final design that was chosen consists of 2 U-joints, an aluminum outboard hub
interface, an aluminum inboard hub interface, and a hollow, splined aluminum drive shaft. The
splining used was 2 mm deep and 6 degrees apart for a total of 30 splines. The U-Joints are
constructed from 1018 cold rolled steel and consists of two splined brackets connected by a
solid steel cross piece. One bracket has a splined section that is 25.00 mm long that is splined
to the hub interface on the outboard and inboard sides. The other U-Joint bracket has a splined
section that is 35.00 mm long, and this connects inside of the hollow aluminum shaft. Starting
from the left side of the drivetrain, there is a 6.35 mm thick outboard hub interface that is bolted
to the wheel hub using 5/16 24 3/4’’ bolts (Home Depot). The hub interface is then connected to
a U-joint by a splined extrusion of 25.4 mm. The splined extrusion of the hub interface is
assembled with the splined extrusion of the U-joint. The splining allows for the hub and the joint
to easily fit together and maintain a connection to transfer torque. This connection is essential to
the durability and performance of the car. Connected to the same U-joint is another splined
extrusion that is assembled to a 368.30 mm splined hollow aluminum shaft constructed from
6061-T6 Aluminum. The right end of the aluminum shaft is is splined to another U-joint. the
other end of the U-joint is splined to a 6.35 mm inboard hub interface with 25.4 mm splined
extrusion. The inboard hub interface is connected to the wheel hub using the 5/16 24 3/4’’ bolts.
Detailed engineering drawings of all parts of the driveline can be found in appendix E. The left
side of the driveline is made up of the same components, except a 266.7 mm aluminum shaft
connects the two U-Joints.This systems of two U-joints allow for the torque to be efficiently
transferred from the engine to the wheels with minimal energy loss as well as allowing for the
vertical mobility required by the restraints of the cars suspension. The splining of the shafts
allow for horizontal translation within the system by the overlap of the splining. The final length
of the right side was 532.8 mm and the left side was 430.6 mm. It was calculated that the left
driveline weighed 2.4786 kg and the right driveline weighed 2.8276 kg. It was decided that all
parts would meet a 1.4 factor of safety to ensure that no components of the driveline would fail
during operation.
Finite Element Analysis
After the design was finalized, the next step was to run finite element analysis through
Abaqus. Each piece was analyzed separately and different constraints, boundary conditions and
loads were applied.
U-Joint
For the U-Joint analysis from figure 9 and figure 10, a plane was created to divide the
holes into two sections in order to apply a static general type surface traction. The pressure on
each side that created a couple force is equal to (½)*Force. Applying the torque equation from
Appendix B and additionally adding the factor of safety of 1.4, the applied force was calculated
to be 1556.8 N. The below figures show the maximum shear stress that is occurring on both UJoints with a maximum deflection multiplied by 7000. The maximum shear stress for figure 9
was 1.73*104 Pa and 1.252*104 Pa in figure 10.
Figure 9: U-joint
Figure 10: Extended U-joint
Cross Joint
The cross joint from figure 11 was simulated in Abaqus similar to the U joints from
figure 9 and figure 10. A plane was created in order to create two independent sections on the
applied surface area. The pressure on each side created a couple force that is equal to
(½)*Force. Applying the torque equation from Appendix B and additionally adding the factor of
safety of 1.4, the applied force was calculated to be 1556.8 N. Figure 11 is showing the
maximum shear stress along with a maximum deflection multiplied by 7000. The actual
deflection experienced by this part was almost negligible, but it was present. The maximum
shear stress that is occurring is 4.19*104 Pa, which is well under the maximum shear strength of
the material according to Appendix A.
Figure 11: Cross joint. This piece connects the U-joint and extended U-joint
Hub Interfaces
For the hub interfaces a reference point was put in the center of the part and the inner
splining was constrained to the reference point. The back of the interface was fixed and a
moment of 62.27N/m was applied at the reference point. The moment was found by applying a
factor of safety of 1.4 to the given maximum torque of 44.48 N/m. The figures below show the
maximum shear stress in the interfaces, figure 12 is 2.698*10-2 Pa and figure 13 is 2.003*10-2
Pa. According to appendix A these values are well below the maximum shear strength of the
material.
Figure 12: Outboard Interface
Figure 13: Inboard Interface
Shafts
For the shafts analysis a plane was added 35 mm in and a partition was made. The
partition was added in order to apply a moment on the specific region of the contact area. Again
a reference point was put at the center of the part and the inner splining was constrained to the
reference point and the moment of 62.27 N/m was applied. The other end of the shaft was fixed.
The figures below show the maximum shear stress in the interfaces, figure 14 is 5.628*10-3 Pa
and figure 15 is 6.996*10-3 Pa. According to Appendix A these values are well below the
maximum shear strength of the material.
Figure 14: Left side shaft
Figure 15: Right side shaft
Fatigue Analysis
One concern with the final design is the use of aluminum for the shaft material. Although
this type of aluminum is extremely strong and lightweight, it is also susceptible to the effects of
fatigue. Figure 16 shows a graph of the fatigue stress experienced by an aluminum alloy.
Although this is not the exact alloy used in the final design, it accurately shows how aluminum
alloys are vulnerable to the effects of fatigue. After about 105 cycles, the maximum stress the
alloy can handle drops by nearly half, which would leave the shaft open to failure. The shaft is
subject to the most impulse torques and by far is loaded and unloaded the most during the
course of a race. Using Abaqus, a fatigue analysis was run and the hollow shaft was loaded and
unloaded many times. Figure 17 shows the results of the Abaqus analysis.
Figure 16: This figure shows a fatigue analysis of an Aluminum Alloy
Figure 17: This figure shows the deformation that occurs in the shaft after 36034 loading cycles
From these results, it can be concluded that the aluminum shaft will be able to be
subjected to the loading and unloading required by the race without having its maximum shear
stress lowering by a significant amount. Although this shaft will not be able to be used for years
to come, it is strong enough to resist the effects of fatigue for these races.
Appendix A
Material Properties
For the design, it was necessary to choose materials that would perform the best as well
as be strong and durable enough to withstand the shear stress involved in multiple uses. Two
materials were chosen to be implemented in the designs, a mild, low carbon steel and a
precipitate strengthened aluminum alloy. The relevant material properties for each alloy can be
found in the table below. (matweb)
Material
1018 Steel
Aluminum 6061-T6
Shear Modulus
205 GPa
26.9 GPa
Possions Ratio
.29
.33
Density
7.86109x10^-6 kg/mm^3
2.82336x10^-6 kg/mm^3
Shear Strength
237.5 MPa
207 MPa
Appendix B
Stress Analysis
Each equation was used in hand calculations in order to determine allowable shear
stress, the twisting which occurred in each part, the polar moment of inertia which was
developed due to the spinning of the material about each axis. Additional variables were derived
due to specific constraints.
τ𝑎𝑝𝑝𝑙𝑖𝑒𝑑 =
𝑇𝑐
𝐽
𝐽𝑐𝑦𝑙𝑖𝑛𝑑𝑒𝑟 =
𝜋 4
∆𝑟
2
∞
φ=∑
𝑖=1
𝐼𝑥 =
𝜋(𝑅 4 − 𝑟 4 )
64
1
1
𝐸 = 𝑚𝑣 2 + 𝐼𝜔2
2
2
𝐹. 𝑆. =
𝑇𝑖 𝐿𝑖
𝐺𝑖 𝐽𝑖
𝑚𝑎𝑡𝑒𝑟𝑖𝑎𝑙 𝑠𝑡𝑟𝑒𝑛𝑔𝑡ℎ
𝑑𝑒𝑠𝑖𝑔𝑛 𝑙𝑜𝑎𝑑
Appendix C
Data and Calculations
See the attached spreadsheet appendix for information about calculations and the final
calculated results for each design.
Appendix D
Bill of Materials
See the attached spreadsheet appendix for information regarding the materials
purchased and the manufacturing time and cost. The manufacturing estimates and costs were
obtained by an interview with Dave Kohler from Ultimate Machining Company.(Americas Metal
Superstore, Home Depot, Manufacturing Estimates)
Appendix E
Engineering Drawings
Left Driveline Assembly Drawing
Left Driveline Hollow Splined Shaft
Right Driveline Assembly Drawing
Right Driveline Hollow Spline Shaft
U-Joint:
Each U-Joint is made up of a cross piece as well as two splined brackets. One of the
brackets has an extended splined shaft component, this bracket goes into the aluminum shaft.
The bracket with the shorted splined component attaches to the inboard or outboard interface.
Cross Piece
U- Joint Splined Bracket with Extended Spline
U-Joint Splined Bracket
Inboard Hub Interface
Outboard Hub Interface
Standard Parts
References
"Americas Metal Superstore." Metals Depot. Metals Depot, n.d. Web. 30 Apr. 2014.
<https://www.metalsdepot.com/>.
"Area Moment of Inertia." The Engineering Toolbox. N.p., n.d. Web. 29 Apr. 2014.
"Automotive Engine Myths and Opinions." Home Built Airplanes. N.p., 22 Mar. 2012. Web. 29
Apr. 2014.
Bradford, Kip. "Carbon Fiber Driveshaft." Pb Works. N.p., 2009. Web. 29 Apr. 2014.
<http://engin1000.pbworks.com/w/page/18942689/Carbon%20Fiber%20Drive%20Shaft>.
"Common Drive Line Issues." Machine Service Inc. Machine Service Inc, 2009. Web. 14 May
2014.
<http://www.machineservice.com/technical-101/drive-shaft-failure-guides/>
Honeychuck, Michael. "Team FASE Powertrain." N.p., 13 Dec. 2010. Web. 21 Apr. 2014.
Kohler, Dave. "Manufacturing Estimates for a Driveline." Telephone interview. 13 May 2014.
Lionnet, Graham N. "Drive Line Design and Implementation for REV Formula SAE Electric
Vehicle." University of Western Australia, 28 Oct. 2013. Web. 29 Apr. 2014.
"Material Property Data." MatWeb. N.p., n.d. Web. 29 Apr. 2014. <http://www.matweb.com/>.
"Metals Depot - America's Metal Superstore!" Metals Depot®. ELCO INC, 2012. Web. 14 May
2014. <http://www.metalsdepot.com/hrsteel2.html>
"Metals Depot - America's Metal Superstore!" Metals Depot®. ELCO INC, 2012. Web. 14 May
2014.<http://www.metalsdepot.com/alum2.html>
"Military Standard Universal Joints." Apex Fasteners. Apex, n.d. Web. 29 Apr. 2014.
<http://apexfasteners.com/fasteners/images/UJCatalogue19.pdf>.
"Results For "5/16 24 3/4'' Bolts" (12 Products)." Search Results For 5/16 24 3/4'' Bolts at The
Home Depot. Home Depot, 2000. Web. 14 May 2014.
Schwartz, Paul. "Formula SAE Powertrain Phase 4: Performance Validation and Path Forward."
Formula SAE Powertrain. University of Delaware, 10 Dec. 2010. Web. 29 Apr. 2014.
<http://sites.udel.edu/meeg401/files/2011/08/FSAE_PowerTrain_Phase_4_NN.pdf>.
"SUPPLIER & DISTRIBUTOR OF UNIVERSAL JOINTS & DRIVESHAFT PRODUCTS."
The Rowland Company. Rowland Company, 2014. Web. 29 Apr. 2014.
Zhu, Y. "Torsion." NC State Archive. NC State, n.d. Web. 29 Apr. 2014.