2014 University of Cincinnati Baja SAE Drivetrain

A Baccalaureate thesis submitted to the Department of Mechanical and Materials Engineering

College of Engineering and Applied Science University of Cincinnati

in partial fulfillment of the

requirements for the degree of

Bachelor of Science

in Mechanical Engineering Technology

by:

Dakota Kirtland

April 2014

Thesis Advisor:

Professor Allen Arthur

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2014 University of Cincinnati Baja SAE Drivetrain

Dakota Kirtland 2014 Bearcats Baja Drivetrain Lead

ABSTRACT

The 2013 Bearcats Baja drivetrain can be significantly improved through more optimized gearing, better

packaging for improved vehicle handling, and reduced weight. The SAE mandated engine, previous drivetrain,

and possible configurations were all analyzed in order to eliminate alternative configurations to ultimately end

up with a belt diven CVT and double reduction gearbox concept. The components of the system were selected

based on the design goals of the system. Worst case failure modes were determined, and analyses were

performed based on these modes to ensure funtionality. At its current state, the project exceeds the weight

reduction and packaging goals set forth from the beginning. Further testing and competition performance will

prove or disprove the effectiveness of gearing optimization.

INTRODUCTION

Every year the Society of Automotive Engineers (SAE) hosts an intercollegiate design competition where

teams must design, build, and race a single seat off-road vehicle powered by a 10 horsepower Briggs and

Stratton engine. For the past three years, senior Mechanical Engineering Technology students at the

University of Cincinnati have competed in the Baja SAE competition to fulfill their senior design capstone

requirement. In order to fulfill senior design requirements, the task of designing, engineering, and fabricating

the drivetrain for the 2014 University of Cincinnati Baja SAE vehicle was chosen. The project was performed in

tandem with a partner, Marcus Schmidt, whose responsibility was to focus mainly on the detail design of a

custom gearbox. The development of this gearbox is contained within a separate report.

PROBLEM STATEMENT – Design an improved overall drivetrain system for the 2014 Baja SAE vehicle.

Previous designs proved unreliable due to chain drive design, but the 2013 vehicle achieved high reliability at

the cost of using a heavy, off the shelf gear box designed for much larger side by side vehicles. The new

design must have reduced weight, equivalent reliability, and more optimized gear reduction when

benchmarked against the 2013 drivetrain system. A lower center of gravity can also be achieved by lowering

the mount position of the engine in order to improve vehicle handling. The target weight delta is 15 lbs less

than the 2013 system.

PRE-DEVELOPMENT

SAE ENGINE RESTRICTION – The competition has strict rules and restrictions for safety and maintaining

competitiveness. The main rule pertaining to the drivetrain is the mandated use of a specific engine, the Briggs

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and Stratton Intek 305cc. This engine is a simple overhead valve, four stroke, air cooled engine. The SAE rules

also state that it must remain completely stock, and that the governor must be set at a maximum speed of

3800 RPM. With this restriction in place, this engine produces a maximum of 14.5 ft-lbs of torque at 2800 RPM

and 9.7 horsepower at 3800 RPM. These values were generated using supplied torque data from Briggs and

Stratton; various data points were taken from the supplied torque chart shown in Figure 1 of Appendix A. The

power was then calculated using the following euation:

𝑃 =𝑇 × 𝑁

5252

Where:

T is torque in ft-lbs,

N is engine speed in RPM,

And P is power in horsepower

These data points were used to create the power and torque outputs shown below in Figure 2.

Figure 2: Briggs and Stratton Engine Output Chart

It is important to note that this engine’s operating RPM is relatively small. With the governor set at 3600 RPM

and the idle set at the factory recommended 1750 RPM, the resulting RPM range is only 1850 RPM.

PAST EXPERIENCE – As a member of the University of Cincinnati Bearcats Baja team since late 2011, the

author has witnessed the progression of the various drivetrain systems over the years. The continuously

variable transmission (CVT) to chain drive systems from the 2011 and 2012 cars proved unreliable due to

stretching of the chain and the resulting chain tension issues. The 2013 team’s solution was to use a CVT to a

double reduction Dana H-12 gearbox sourced from a Gravely Trekker side-by-side. This gearbox featured

different reduction ratio options and forward, neutral, and reverse (FNR) capability, a Bearcats Baja first. The

system was very successful, and the functionality of reverse proved invaluable during the competition. The

main problem was the obvious weight penalties that accumulated by retro-fitting the gearbox from a larger,

more powerful vehicle.

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DESIGN CONFIGURATIONS – Several configurations were considered as viable options to design upon:

Sequential 5 speed transmission to single reduction gearbox

Belt driven CVT to FNR planetary

Belt driven CVT to double reduction gearbox

Sequential 5 Speed to Single Reduction – This option was considered mainly because of the efficiency of a

sequential gearbox when compared to the more popular choice of a CVT. Gear drives are typically 95 - 98%

efficient where a CVT is anywhere from 70 - 80% efficient during low ratio engagement. In the Baja application,

the engine is severely power limited so this efficiency gain would be quite noticeable. Also, the added aspect of

being able to manually clutch and shift would add another element of vehicle control. The idea behind Baja

application of this configuration is to remove the gearset internals out of an ATV or dirt bike, and then create a

custom case that could be coupled to the Briggs and Stratton engine. The internals of a viable gearbox are

shown below in Figure 3.

Figure 3: Internal Gearset from a 1994 Kawasaki Bayou

After studying the power curve of the Briggs and Stratton Intek 305cc engine, it becomes painfully obvious that

this engine is not very well suited for a sequential gearbox. The usable RPM range is so small that it would

require instant shifting every time the vehicle encounters an obstacle such as a turn or hill. The vehicle would

likely spend more time with the clutch disengaged than it would actually propelling the vehicle around a

maneuverability course. Also, The Briggs and Stratton Intek has a tendency to stall very easily when shock

loaded. The author has witnessed Northeastern, a team that consistently uses a sequential gearbox, stall their

car during dynamic events such as the hill climb. This results in that run being disqualified and no points being

scored. In short, the engine just does not make enough power over a wide enough RPM range to be suitable

for a sequential gearbox in an off-road scenario.

Belt Driven CVT to FNR Planetary – This configuration would employ a belt driven CVT to a planetary driven

rear axle similar to the design shown below in Figure 4. This option would require an additional reduction stage

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before or after the planetary; this could be a duplicate planetary box coupled in line with the first or a simple

chain reduction as shown below. The main appeal to a planetary gearbox is that reverse functionality can be

achieved by locking different parts of gear train such as the planet carrier. This accomplishes reverse capability

with reduced weight when compared to a standard double reduction gearbox.

Figure 4: A Solid Model of Wichita State’s Planetary Gearbox Layout

This configuration choice was abandoned after considering manufacturing of the internal outer ring gears

required. During the configuration selection phase, access to a Wire EDM machine had not yet been achieved

so typical broaching techniques would not be practical for an internal gear of that size.

Belt Driven CVT to Double Deduction Gearbox – This configuration would employ a belt driven CVT coupled to

a double reduction gearbox similar to the Dana H-12 (shown below in Figure 5) used in the 2013 car.

Figure 5: The Dana H-12 Gearbox used in 2013

The benefit of a CVT when compared to the sequential gearbox is that it eliminates the shifting lag time that

results from changing gears. This also eliminates the possibility of stalling the engine during dynamic events as

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well. As stated above, the Briggs and Stratton engine does not have a very broad range of usable engine

speed. A CVT system is forgiving in this aspect because it holds engine RPM constant as it varies the gear

ratio while the vehicle accelerates, making it much more suitable than the sequential gearbox for Baja

applications. Because CVT drives naturally incorporate a centrifugal clutch, they also promote weight reduction

and simplicity when compared to the sequential five speed gearbox; instead of 10 gears, three shafts, a clutch,

case, and gear oil, there is only a driver sheave (primary), a driven sheave (secondary), and a belt for the

continuously variable transmission. CVT systems applicable to engines of similar torque output and engine

speed as the Briggs and Stratton generally only weigh 15 – 25 lbs.

The manufacturing of this style of gearbox also offers increased feasibility when compared to the planetary

drive because of the gears and shafts can be manufactured by any of the gear manufacturers within Cincinnati.

The gearbox case can then be CNC machined by a team member, Randy Knepp, who has access to CNC

equipment through his co-op employer. If the luxury of CNC access was not available, the case would have to

be machined manually on a vertical mill, and subsequently, holding the necessary tolerances for bearing race

diameters and shaft center to center distances could be very problematic.

In the end, this combination of a lighter, more efficient belt driven CVT and a custom double-reduction FNR

gearbox offered the best design and manufacturing feasibility given the time constraints of the project while still

maintaining feasibility to hit the target weight reduction. The system also allowed for reverse capability within

the gearbox, which the team decided was mandatory for the final design.

DEVELOPMENT

GEARING OPTIMIZATION – In order to move forward with the design of the drivetrain, the optimal overall

reduction ratio had to be determined in order to calculate the worst case torsional loads that each component

would have to withstand. The 2013 team had the luxury of changing gearing because of the modular design of

the Dana H-12. Their gearbox reduction ratio could be chosen as 10.15:1, 12.58:1 or 13.25:1. When used in

series with their CVT’s low ratio of 3.0:1, created overall reduction ratios 30.5:1 and 39.8:1 for the 10.15:1 and

13.25:1 gearing, respectively. When coupled to the chosen tire diameter of 23 inches and using the CVT high

ratio of 0.5:1 (as quoted from Benjamin Steele’s 2013 senior design report), this would create engine governor

(3800 RPM) limited theoretical max speeds of 51 mph with 10.15:1 gears and 39 mph for 13.25:1 gears.

Because these top speeds are based on engine speed alone, the car is not guarenteed to achieve the

designed top speed due to both rolling and aerodynamic drag. Even with the lower 13.25:1 gears in the Dana

H-12, the 2013 team reported that they never fully shifted out to achieve the top speed of 39 mph due to power

limitations, because of this they recommended lowering the top end gearing of the vehicle.

The 2013 test reports claimed that their 400 lb car had achieved 35 mph during testing on level ground. Even

so, the endurance race drivers reported that the majority of each endurance course they participated in did not

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allow enough level ground or straight section to achieve even this speed. Furthermore, research into team

history proves that acceleration could greatly be improved as shown below in table Table 1.

Table 1: University of Cincinnati Baja Acceleration Results

Year Rank/100 Event

2011 67 Kansas

2012 59 Auburn

2013 31 Tenn. Tech

2013 65 Rochester

The acceleration results of previous cars have been subpar. This was enough evidence to support the fact that

gearing the car to the theoretical power limited top speed is not the best method for overall performance at

competition.

Because there was never any data to analyze other than the opinion of the previous drivers and the evidence

that the governor limited top speed was never achieved, the methodology chosen for selecting the 2014

gearing was to set the governor limited top speed to a reasonable value that would increase acceleration while

maintaining a competitive top speed. Based on this logic, a desired top speed of 32 – 35 mph would be

acceptable. This governor limited top speed is calculated throgh the following equation:

𝑉 =𝑁

(𝐻 × 𝐺)(𝜋 ∙ 𝑑) ×

1 𝑓𝑡

12 𝑖𝑛×

1 𝑚𝑖

5280 𝑓𝑡× 60

𝑚𝑖𝑛

ℎ𝑟

Where:

𝑉 is vehicle speed in mph

N is engine speed in RPM

H is high reduction ratio of the CVT

G is the reduction ratio of the gearbox

d is tire diameter in inches

Using the above equation and rearranging to solve for overall reduction ratio, a 32.5 mph top speed would

yield:

(𝐻 × 𝐺) =𝑁

𝑉(𝜋 ∙ 𝑑) ×

60

63360=

3800 𝑟𝑒𝑣/𝑚𝑖𝑛

32.5 𝑚𝑝ℎ(𝜋 ∙ 23 𝑖𝑛) ×

60

63360= 8.00

This means that the overall reduction ratio of the gearbox multiplied with with the high ratio of the CVT would

ideally be 8.0:1 in order to achieve a top speed of 32.5 mph with a 23 inch tire.

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CVT SELECTION – Because the decision to design a custom gearbox where the gear ratios could be

controlled, the next step was to select a CVT so the low and high ratios could be known and gearbox layout

could begin. Severall different CVTs were considered including the Comet 500 Series, several CVTech

models, and the Gaged GX9 Engineering Baja series. In the end, the Gaged GX9 series was selected based

on its all 6061-T6 aluminum construction and its roller helix secondary design. This setup features an 6 inch

primary and an 8 inch secondary with a resulting low ratio of 3.9:1 and a high ratio of 0.9:1. The manufacturer’s

recommended center to center distance is 10 inches. While other models such as the CVTech use primary fly

weights and secondary shift ramps that slide along one another, the Gaged setup utilizes roller bearings in

these high friction areas that increases the overall efficiency of the system. Also, the weight of the Gaged setup

is signinifcantly less with a total of 12.2 lbs for the primary, secondary and belt shown in Figure 6; a total of 2.3

lbs less than than the CVTech system used in 2013.

Figure 6: The Gaged Engineering GX9 CVT

GEARBOX RATIO – With the CVT low and high ratios known, the desired gearbox reduction ratio could be

calculated based on the CVT high ratio of 0.9:1 and the desired overall drivetrain reduction ratio 8.00:1 as

shown below:

𝐺 =(𝐻 × 𝐺)

𝐻=

8.00

0.9= 8.89

With this ideal reduction ratio of 8.89:1, Marcus Schmidt could begin the initial layout and design of the

gearbox internals. Because of packaging and actual gear selection, the final overall ratio of the gearbox

resulted in 9.04:1 which results in a governor limited top speed of 32.0 mph. As previously stated, the detailed

design of the gearbox internals can be found in Mr. Schmidt’s senior design report.

CV JOINT SELECTION – In order to tranfser power from the gearbox to the rear wheels of a vehicle with

independent rear suspension, axle shafts need to be able to move and plunge with the rear suspension. To

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accomplish this, constant velocity (CV) joints are used. CV joints are typically made of an outer cup with

internal grooves for ball bearings to ride. These ball bearings ride in a bearing cage that is connected to the

end of a shaft; the ball bearings transmit torque while allowing the shafts to vary angle and plunge in and out of

the bearing cup. A section veiw of the solid model of the TRX420 inboard CV Joint is shown below in Figure 7,

and a picture of the actual CV joint is shown next to it in Figure 8.

Figure 7: Section View of TRX420 Inboard CV Joint Figure 8: Actual TRX420 Inboard CV Joint

Because of the extrememly tight tolerances within CV joint parts and the very extreme manufactuing

thechniques involved, original equipment driveshaft and CV joint assemblies were purchased. In this case,

Honda TRX420 ATV front driveshafts and joint assemblies were selected, and because the TRX420 engine

outputs 26.2 horsepower there is no concern of the CV Joints or axles failing due to torsion. Using a set of

calipers and a digital angle finder, plunge and maximum angle specifications were measured for both the 2013

and the 2014 CV joints and are displayed belore in Table 2.

Table 2: CV Joint Measurements

2013

Inboard Plunge 1.125"

Inboard Angle 30°

Outboard Angle 45°

2014

Inboard Plunge 1.375"

Inboard Angle 35°

Outboard Angle 49°

It can is easily be seen that Honda TRX420 components offer increased performance when compared to the

Ariens and Polaris components used in 2013.

AXLE SHAFTS – In order for the CV and axle shaft assembly to function throughout the travel of the rear

suspension, the distance from the center point of the inboard CV to the centerpoint of the outboard CV must be

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customized. In this case the stock TRX420 shafts were not long enough and required increased length. To

achieve this, the shafts were cut and a sleeve was designed that the shaft ends would be press fit into. The

shaft is made of AISI 4340 steel and has 0.5 inch and 1.0 inch diameters inside and out, respectively. The

ends of the shoulder on the sleeve were then fillet welded to the shaft using the GTAW process. For increased

torque capacity, 0.375 inch holes were added to the sleeve for rosette welds. The sleeved axle and CV

assembly is shown below in Figure 9, and detail drawing of the sleeves can be found in Appendix C.

Figure 9: Sleeved Axle and CV Assembly

SUBFRAME – Mounting the drivetrain to the frame and holding the drivetrain components together is also a

responsibility of this project. The center to center dimension of the output shaft of the engine to the input shaft

of the gearbox is extremely crucial for correct CVT operation. If the dimension is too far over the nominal

distance, the belt will be too tight and will experience premature wear. If the distance between shafts is too

short, there will be a significant efficiency loss in low ratio engagement. Also, for packaging and CVT sheave

alignment purposes, the engine was shifted 0.5 inches to the right side of the vehicle centerline. For the

purpose of lowering the vehicle’s center of gravity, the engine was lowered by 7 inches when compared to

2013. In order to hold these critical dimensions, a subframe fixture was created and can be seen below in

Figure 10.

Figure 10: The Solid Model of the Drivetrain Subframe

The fixture is composed of normalized 4130 sheet steel and will eventually become part of the frame

weldment. The main body plates are 0.125 inches thick and the center gussets are 0.800 inch thick. The fixture

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is skeletonized for weight reduction, and it was designed specifically to be cut on the CNC Plasma CAM in the

University of Cincinnati’s Victory Parkway weld shop. The center gussets have tabs that are designed to slide

into the slots cut on the side plates and the engine mounting plate. This fixes the side plates and engine mount

plate relative to one another in all three directions, and because the plates are cut with CNC accuracy, the

aforementioned critical dimensions will easily be held. The fixture will be clamped together and joined using the

gas tungsten arc welding (GTAW) process. The fixture is shown holding the drivetrain components below in

Figure 11.

Figure 11: Solid Model of the Subframe Holding the Drivetrain Components

FAILURE ANALYSIS – Several modes of failure were considered for the drivetrain as a whole. The most

severe were determined to be overtravel of allowable CV Joint plunge and angle, gear case failure, and axle

shaft torsional failure.

CV Joint Plunge and Angle Analysis – Because the CV joints have limited plunge and angle travel, it is

important to make sure that this travel is not exceeded when the rear suspension reaches full bump or full

droop. If the plunge is exceeded with enough force when the suspension is in droop, the inboard CV joint can

be pulled out of the cup that holds the ball bearings. Conversely, if the plunge is exceeded in bump, the axle

shafts will bottom out in the bearing cups and will be absorbing the load meant for the shock absorbers.

Furthermore, if the maximum angles of the CV joints are exceeded, the joints will see rapid wear and will

ultimately fail prematurely.

In order to insure the joints will not exceed their allowable travels, a dynamic model of the rear suspension had

to be created in order to simulate the suspension as it travels from each of its extremes. With this model, the

optimized shaft length was determined by altering the length value incrementally and then re-running the

simulation model. A still image of this model (showing bump, static ride height, and droop) and the various

travel results is shown below in Figure 12.

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Figure 12: The CV Plunge and Angle Simulation

Gear Case Strength Analysis – The most obvious mode of failure for the gear case is from the bearing loads

applied to the bearing races of the case. Marcus Schmidt’s aforementioned Matlab code was used to resolve

the vector components of the support loads on the shafts at each bearing location for the worst case torque in

both forward and reverse scenarios. The directions of these loads were then reversed because the loads on

the case would be equal and opposite of those which were exerted on the shafts. The vector components (in

the vehicle coordinate system) of these loads, which can be seen below in Table 3, were then applied to the

bearing races of a finite element model of the gear case.

Table 3: Shaft Support Load Components

Forces on Bearing & Shafts

Forces on Case

Forward (lbs) Forward (lbs)

Left Right Shaft Direction Left Right

Input Y 1075 674

Input Y -1075 -674

Z 896 5.1 Z -896 -5.1

Inter Y -3967 -1374

Inter Y 3967 1374

Z 938 802 Z -938 -802

Output Y 2873 694

Output Y -2873 -694

Z -1406 -258 Z 1406 258

Reverse (lbs) Reverse (lbs)

Left Right Shaft Direction Left Right

Input Y -88 -135

Input Y 88 135

Z 897 1374 Z -897 -1374

Inter Y 2914 1582

Inter Y -2914 -1582

Z -1586 -1617 Z 1586 1617

Output Y -2677 -647

Output Y 2677 647

Z 1248 302 Z -1248 -302

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Figure 13: Gear Case FEA Images

The resulting Catia finite element analysis (FEA) results can also be seen above in Figure 13. In the image, the

worst case stress distribution on the exterior of the case is shown to the left. To the right a section view of the

finite element model is shown with the highest local stresses shown. The scale of the image is set from 0 –

8,500 psi with a maximum global stress of 7,925 psi. The case is machined from 6061-T6 aluminum with a

heat lot yield strength certification of 70,000 psi; the resulting safety factor is 8.8. While this may seem high,

the Baja SAE rules mandate that all rotating parts spinning at a rate faster than the driveshafts must be

guarded. These guards must be made of “at least AISI 1010 strength steel at least 1.5 mm (0.06 in) thick” or

“at least 6061-T6 strength aluminum at least 3.0 mm (0.12 in) thick.” The standard wall thickness of the case

was already 0.125 inches thick, so reducing this thickness would violate the competition rules. For this reason,

the case was not optimized further.

Axle Shaft Strength Analysis – Because the axle shafts were designed to be cut, sleeved, and welded, it was

necessary to analyze the welded connection between the shaft and sleeve as well as the entire sleeve itself

under torsion. A problem arises when one considers which loads to apply to the shafts in an analysis; the axle

shafts will most certainly undergo torsional shock loading. As a worst case scenario, one could imagine the

vehicle landing after being air born with the throttle wide open. This energy input could be calculated

analytically, but without real torsional strain data of previous designs to verify this load, it would more than

likely end up being speculation.

Because the 2013 team used a sleeved shaft design similar to the design described above on a vehicle of

greater weight and equal power without any failure, it is acceptable to analyze the two designs side by side and

compare the results. This is only valid given that the 2013 shafts have been proven, which they are considering

the three competitions and numerous hours of testing they have undergone without any failures.

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In order to compare the two designs directly, the driveshaft models were each converted to a finite element

model in Catia using the same mesh size and global sag constraints. The torsional loads to apply were derived

by multiplying the maximum output torque of the engine by the respective product of the low CVT reduction

ratio, gearbox reduction ratio, and the assumed CVT low ratio efficiency. The calculations are shown below:

𝑇 = 𝑇𝑒𝑛𝑔𝑖𝑛𝑒 × (𝐿 ∙ 𝐺 ∙ η)

Where:

T is the respective shaft torque to apply, in ft-lbs

Tengine is the maximum engine torque, in ft-lbs

L is the low ratio of the CVT

G is the overall reduction ratio of the gearbox

η is the assumed low ratio CVT efficiency

𝐹𝑜𝑟 𝑡ℎ𝑒 2013 𝑠ℎ𝑎𝑓𝑡 𝑎𝑛𝑎𝑙𝑦𝑠𝑖𝑠: 𝑇 = 14.5 𝑓𝑡 ∙ 𝑙𝑏𝑠 (3.0 ∙ 13.25 ∙ 0.80) = 461 𝑓𝑡 ∙ 𝑙𝑏𝑠

𝐹𝑜𝑟 𝑡ℎ𝑒 2014 𝑠ℎ𝑎𝑓𝑡 𝑎𝑛𝑎𝑙𝑦𝑠𝑖𝑠: 𝑇 = 14.5 𝑓𝑡 ∙ 𝑙𝑏𝑠 (3.9 ∙ 9.04 ∙ 0.80) = 409 𝑓𝑡 ∙ 𝑙𝑏𝑠

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Figure 14: Shaft Analysis Section View Comparison

As shown above Figure 14, the area of concern on both shafts is where the weld holding the sleeve to the

shaft terminates. The local weld maximum stress values were 171,000 psi and 148,000 psi for the 2013 and

2014 shafts, respectively. This results in a 14.4% reduction in critical weld stress. The stresses within the

actual sleeves are also noticeably lower in the 2014 model. Admittedly, the lower stresses in the 2014 shafts

were most likely due in part to the less severe overall reduction ratio, but the bottom line is that the stresses in

the areas of concern, the weld and sleeve areas, are globally lower than the benchmark. Because of this, it is

fair to declare the new shaft design safe.

Once again, it should be clear that the magnitudes of these stresses are not what should be examined; rather

the relative comparison is what is to be taken away from this. Also, the stresses within the original OE shafts

can be ignored given the much higher torque and power requirements of the original applications of these

shafts. Furthermore, the torque capacity of the CVT belts in the low ratio engagement can be offered as an

explanation to why the shafts will most likely never see torsional loads this severe.

As for a clarification of the image above, the 2013 shafts had a great majority of the original shaft left inside the

sleeve. In Figure 14, it does appear to be a solid shaft, but the opposite shaft ends are actually separate solid

2013 Shaft Analysis

2014 Shaft Analysis

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bodies in both the model and on the real life parts, hence the absence of stress in the original shaft shown in

the analysis.

MANUFACTURING

AXLE SHAFTS – The axle shafts and CV assemblies began life as stock Honda OE parts. The shafts were cut

down the center in preparation for their press fit assembly with the sleeves, and a slight chamfer was applied

using a bench grinder. The sleeves started out as 1.0 inch solid AISI 4340 steel rod. The rod was first band

saw rough cut, and then a 0.5 inch through hole was drilled through each of the resulting rough cut rods. Next,

the the sleeve blanks were faced on the lathe to the desired 6.5 inch overall length. After this, a 0.625 inch hole

was drilled on each side of the shaft to the specified minumum depth of 1.5 inches. The final step in creating

the sleeve was to bore out the 0.625 inch hole to 0.717 inches using a boring bar. An image of this operation

can be seen below in Figure 15 in Appendix A.

The next step was to press the cut OE shafts into the shaft sleeves. Because this is an interferance fit, a

significant amount of force was necessary. Because the outboard CV joint is designed to never be

disassembled and a direct compressive force could damage the joint, a special press jig was fabricated. The

jig clamps around the shaft above the boot band shoulder to allow the press load to be transferred into the

shaft. Also, because the distance between shaft referance points overall length was larger than alailable

calipers was measured, a special press ruler was milled to the appropriate length. The asssembled press jig

and the press ruler can be seen below in Figure 16.

Figure 16: The Assembled Press Jig, Press Ruler, and a Pressed Shaft Assembly

SUBFRAME – As mentioned above, the subframe was designed specifically for use with a Plasma CAM

machine. Because of this, the subframe came together rather easily. The sheet bodies were converted to DXF

wireframe models and then imported to the Plasma CAM machine which quickly cut them out. After this, the

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crucial mount holes were cleaned up on a drill press. The pieces of the subframe were then assembled by

placing the tabs of the center gussets into their proper slots on the side plates and top plate. The assembly

was clamped together using a variety of C-clamps and quick grips, and it was finally tack welded together

using the GTAW process with careful attention to not warp the plates. Various finish welds were then applied

with the same discretion. Images of this process can be viewed in Figures 17 and 18 of Appendix A.

TESTING

ENGINE DYNOMOMETER – As a Bearcats Baja first, the competition engine was broken in and tested using

a dynomometer. This was achieved by removing the standard 6 horsepower Briggs and Stratton engine from

the Otto Cycle Thermodynamics Lab apparatus located in the machine shop of the University of Cincinnati’s

Victory Parkway campus. The Briggs and Stratton Intek 305cc shared the same vertical shaft offset, so the

only additional steps to couple the engine were new mounting holes and a different coupler for the larger

diameter output shaft of the Intek 305cc. This dynomometer can be seen below in Figure 19.

Figure 19: The B&S Intek 305cc Coupled to the Dynamometer

BENCHMARK TESTING – At the time this is being published, complete benchmark testing of the 2014 vehicle

against the 2013 vehicle has not been achieved, but complete testing and validation will be achieved before

the first competition of May 22, 2014.

RESULTS

WEIGHT TARGET – The weight results are displayed below in Table 4. The results were pleasing as the new

system weighed in under the target by 2.3 lbs. A majority of this weight was removed through the design of the

new gearbox, but collectively the incremental optmization of the axle shafts and subframe as well as the CVT

selecton also made a large impact to system weight. The more detailed list of part weights from 2013 and 2014

can be found Table 5 of Appendix B.

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Table 4: Weight Results

2013 System Weight 75.7

2014 System Weight 58.3

Weight Delta 17.3

Reduction % 26.0%

MONETARY BUDGET – The entire system, including the gearbox, was allotted a total budget of $4,868. With

several sponsorship donations, the total cost of the system was lowered to $1,595.88. This leaves $3,272.12

remaining. This large remainder is due to the generosity of our major sponsors: Cincinnati Gearing Systems,

Kaiser Aluminum, and Honda of Fairfield. A detailed cost sheet is displayed in Table 6 of Appendix B.

CONCLUSION

The purpose of this project was to develop a complete drivetrain that offered more optimized gearing for Baja

SAE competition, a lower overall center of gravity, and reduced weight by at least 15 lbs. While the

optimization of the overall gearing cannot be confirmed until the first competition, the 26% weight loss and

achieving packaging that lowered the engine by 7 inches can immediately be recognized as fulfillment of the

original goals.

ACKNOWLEDGEMENTS

A big thanks goes out to all of the Bearcats Baja teamates past and present for keeping the program alive. The

author would like to express the deepest gratitude to Thomas Miller, Kyle Grefer, and Maeve McGoff of

Cincinnati Gearing Systems for providing the time and resources the drivetrain team needed to accomplish

their goals. Thanks and credit are owed to Marcus Schmidt for being an excellent teammate and design

partner throughout this project. In addition, thanks must be given to Randy Knepp for helping out with the gear

case design and, more importantly, the CNC machining of the case. Finally, the team’s faculty advisor, Dean

Allen Athur, must be acknowledged for devoting his time and patience to the team during this development

and build.

Also, many thanks are owed to all of the sposors that made this project happen through their generous

donation of money, products, and services: Cincinnati Gearing Systems, Gallatin Steel, Kaiser Aluminum,

Honda of Fairfield, Cincinnati Steel Treating, Toyota, Wilwood Engineering, General Cable, GE Aviation,

Grainger, Timken, Discount Tire, VR3 Engineering, and Solidworks Corp.

18

REFERENCES

"2014_baja_rules_8-2103." Rules & Documents. SAE Collegiate Design Series, 2014. Web. 22 Apr. 2014.

Aaen, Olav. Olav Aaen's Clutch Tuning Handbook. Racine, WI: AAEN Performance, 1999. Print.

"BajaSAE Forums." BajaSAE Forums. N.p., n.d. Web. 22 Apr. 2014.

McCausland, Michael, Michael Watkins, Ian Masterson, and Andrew Sommer. "SAE Baja: Final Drive

Gearbox." Thesis. California Polytechnic State University, 2010. Web. 23 Apr. 2014.

Mott, Robert L. Machine Elements in Mechanical Design. Upper Saddle River, NJ: Pearson/Prentice Hall,

2004. Print.

Steele, Benjamin. "2013 University of Cincinnati Baja SAE Drivetrain." Thesis. University of Cincinnati, 2013.

Print.

Vamos, Bryan. "2013 Baja SAE Drivetrain." Thesis. University of Cincinnati, 2013. Print.

CONTACT

Dakota Kirtland

Drivetrain Lead

kirtladd@mail.uc.edu

19

APPENDIX A: FIGURES

Figure 1: Torque Curve from Briggs and Stratton

Figure 2: Horsepower and Torque curves of the Briggs and Stratton Intek 305cc

20

APPENDIX A: FIGURES

Figure 3: Internal Gearset from 1994 Kawasaki Bayou

Figure 4: A Solid Model of Wichita State’s Planetary Gearbox Layout

Figure 5: The Dana H-12 Gearbox used in 2013

21

APPENDIX A: FIGURES

Figure 6: The Gaged Engineering GX9 CVT

Figure 7: Section View of TRX420 Inboard CV Joint

Figure 8: Actual TRX420 Inboard CV Joint

22

APPENDIX A: FIGURES

Figure 9: Sleeved Axle and CV Assembly

Figure 10: The Solid Model of the Drivetrain Subframe

Figure 11: Solid Model of the Subframe Holding the Drivetrain Components

23

APPENDIX A: FIGURES

Figure 12: The CV Plunge and Angle Simulation

Figure 13: Gear Case FEA Images

24

APPENDIX A: FIGURES

Figure 14: Shaft Analysis Section View Comparison

Figure 15: Boring the Axle Sleeves on Lathe

25

APPENDIX A: FIGURES

Figure 16: The Assembled Press Jig, Press Ruler, and a Pressed Shaft Assembly

Figure 17: The Plasma CAM Cutting Out Subframe Plates

Figure 18: The Subframe during the GTAW Process

26

APPENDIX A: FIGURES

Figure 19: The B&S Intek 305cc Coupled to the Dynamometer

Figure 20: The Otto Cycle Lab Apparatus

27

APPENDIX B: TABLES

Table 1: University of Cincinnati Baja Acceleration Results

Year Rank/100 Event

2011 67 Kansas

2012 59 Auburn

2013 31 Tenn. Tech

2013 65 Rochester

Table 2: CV Joint Measurements

2013

Inboard Plunge 1.125"

Inboard Angle 30°

Outboard Angle 45°

2014

Inboard Plunge 1.375"

Inboard Angle 35°

Outboard Angle 49°

Table 3: Shaft Support Load Components

Forces on Bearing & Shafts

Forces on Case

Forward (lbs) Forward (lbs)

Left Right Shaft Direction Left Right

Input Y 1075 674

Input Y -1075 -674

Z 896 5.1 Z -896 -5.1

Inter Y -3967 -1374

Inter Y 3967 1374

Z 938 802 Z -938 -802

Output Y 2873 694

Output Y -2873 -694

Z -1406 -258 Z 1406 258

Reverse (lbs) Reverse (lbs)

Left Right Shaft Direction Left Right

Input Y -88 -135

Input Y 88 135

Z 897 1374 Z -897 -1374

Inter Y 2914 1582

Inter Y -2914 -1582

Z -1586 -1617 Z 1586 1617

Output Y -2677 -647

Output Y 2677 647

Z 1248 302 Z -1248 -302

Table 4: Weight Results

2013 System Weight 75.7

2014 System Weight 58.3

Weight Delta 17.3

Reduction % 26.0%

28

APPENDIX B: TABLES

Table 5: Total Weight Rollup

2013 Weights 2014 Weights Qty Part

Unit Weight

Sub-Total

Qty Part

Unit Weight

Sub-Total Notes

1 Engine/Gearbox Fixture 5.60 5.600 1

Engine/Gearbox Fixture 3.16 3.163 actual

1 CVT Drive Pulley 8.44 8.438 1 CVT Drive Pulley 7.23 7.230 actual

1 CVT Driven Pulley 5.17 5.169 1 CVT Driven Pulley 4.33 4.325 actual

1 CVT Belt 0.82 0.819 1 CVT Belt 0.63 0.625 actual

1 Dana H12 35.50 35.500 1 9.0:1 Gearbox 25.74 25.740 actual internal, case est. (dry)

2 Driveshafts (x2) 6.79 13.575 2 Driveshafts (x2) 6.44 12.888 actual

1 Mounting Hardware 0.000 1 Mounting Hardware 0.00 0.000 Acct'd for in GB

1 CVT Guarding 5.00 5.000 1 CVT Guarding 2.81 2.813 actual

1 Shifter Linkage & Hardware 0.56 0.563 1

Shifter Linkage & Hardware 0.56 0.563 Carryover

1 Splash Shield 1.00 1.000 1 Splash Shield 1.00 1.000 Actual

Total: 75.66 Total: 58.35

Target: 75.00 Target: 60.66

Note: All weghts in lbs DELTA: 2.32

29

APPENDIX B: TABLES

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Tabl

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Dri

vetr

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Cost

She

et

CVT,

Dri

vesh

aft,

and

Gea

rbox

Cos

t

30

APPENDIX C: MISC. ITEMS

31

APPENDIX C: MISC. ITEMS

2014 Bearcats Baja Drivetrain Specifications

Transmission CVT, Gaged 6" four arm drive, Gaged 8" roller helix driven

Final Drive Type Student built single speed FNR gearbox 6061-T6 AL

Final Drive Ratio 9.04

Differential Type Solid spool

Theoretical Top Speed (Power limited, Ratio Limited) 32 mph, 32 mph

Low CVT Ratio 3.9

High CVT Ratio 0.9

Half shaft size and material 25.4mm OD 4340 sleeves welded over 18.7mm Honda TRX420 front

axles

Joint type Honda TRX420 6 ball

Maximum grade capability Calculated estimate 55%

Predicted max axle torque 409 lbs-ft (555 N-m)

Peak driveline torque threshold, what component fails Belt failure before other component failure, untested to date