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ME 497 Design Report

The Great Moonbuggy Global Team 2003

Team Members:

Chris Auth

Cliff Nurrenbern

Branden Horne

Matt Snodgrass

Valerie Stringer

Tommy Woods

4/23/03

Submitted to the Faculty of the Mechanical Engineering Program College of Engineering and Computer Science

University of Evansville Evansville, Indiana 47722

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Abstract: (Chris and Cliff) For the past 120 days we have attempted to build a moonbuggy capable of

winning the 10th Annual Great Moonbuggy Race held April 12 in Huntsville, Alabama.

To accomplish this feat we have designed a moonbuggy that exceeds all the competition

requirements and has innovative features that will help it tackle harsh terrain. We

designed a light weight aluminum frame that is hinged in the center to provide for quick

assembly. The frame was modeled in ProE and analyzed using FEA software to evaluate

performance and eliminate unnecessary material to reduce weight. An independent

suspension system has been designed that utilizes a dual a-arm construction combined

with shocks, rockers, uprights, and custom hubs to provide 8 inches of travel and superior

ride. The moonbuggy is maneuvered using a simple yet effective steering system

consisting of two steering arms, a butterfly, and connecting rods. The moonbuggy’s

drive train incorporates collapsible pedal supports to give the riders extra leg room.

Internal hubs have been incorporated into the design to provide the rider a range of

shifting speeds without the risk of chain-fall offs common on many derailleur systems.

Separately driven axles contain roller clutches that allow the wheels to operate at

different speeds during turning and don’t have the disadvantages and weight of regular

differentials. We modeled the entire moonbuggy in ProE to evaluate system interaction

and provide drawings needed for construction. The steps took in the design and

construction of our moonbuggy minimized problems and breakages that occurred during

testing and the competition.

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Acknowledgements: • University of Evansville -Aluminum tubing and use of shop equipment • O’Neal Steel - Discounted Aluminum • Bicycle World - Technical assistance on the internal hubs and shifters • Mr. Funk - Advice on drive train and general moonbuggy • Ottis Putler - General advice and CNC fabrication • SGA - Funds for trip down to Huntsville • ISC – Funds to purchase parts needed to build buggy Trent Zuehsow a freshmen team member was fatally injured in a car accident on February 8,

2003. Although his time with the project was very short, he showed great enthusiasm about

the project and is still considered a part of our team.

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Lists:

I. Abstract………………………………………………………………… 2 II. Acknowledgments………………………………………………………3 III. Lists……………………………………………………………………...4 IV. Introduction

A. Background………………………………………………………… 6 Figure 1 – Map of Moonbuggy Course……………………………..6

B. Project Objective and Scope of Work………………………………7 V. Design

A. Suspension i. Conceptual Design…………………………………………. 8

Figure 2 – Suspension Conceptual Design………………… 9 ii. Analytical Method…………………………………………..9

Figure 3 – Lower A-arm with Support…………………….. 10 Figure 4 – Rocker…………………………………………... 11 Figure 5 – Suspension Assembly…………………………... 13 Figure 5a – Front Hub………………………………….........15 Figure 5b – Rear Hub………………………………………..15 Figure 6 – Rear Suspension Assembly……………………...17 Figure 7 – View of Front Rocker and Shock Assembly….....18

iii. Prototype Design……………………………………………16 B. Steering

i. Conceptual Design…………………………………………. 18 Figure 8 – Conceptual Steering Design……………………. 18

ii. Analytical Method…………………………………………..19 Figure 9 – Steering Butterfly………………………………. 21

iii. Prototype Design…………………………………………… 22 Figure 9a – Steering Butterfly and Connecting Rods……… 23

Figure 9b– New Steering with Plate of Aluminum………... 23 C. Frame

i. Conceptual Design…………………………………………. 24 Figure 10– Conceptual Frame………………………………24 Figure 11– Conceptual Frame Collapsed………………….. 25 Figure 12– Conceptual Frame Front…..…………………… 25

ii. Analytical Method…………………………………………..25 Figure 13– Model of Front Frame…………………………..26 Figure 14 – Model of Rear Frame…………………………..27

iii. Prototype Design..………………………………………….. 28 Figure 15 – Moonbuggy Assembly…………………………28 Figure 15a – Moonbuggy Assembly……..…………………29 Figure 16 – Collapsed Moonbuggy…………………………29 Figure 16a – Collapsed Moonbuggy……………………….. 30

D. Drive Train i. Conceptual Design…………………………………………. 33

ii. Analytical Method…………………………………………..35 Figure 17– Crank Support Loading…………………………35

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Figure 18 – Keyway Loading……………………………… 39 Figure 19 – Gearing Spreadsheet………………………….. 40

iii. Prototype Design..………………………………………….. 41 Figure 20 – Model of Crank Assembly..……………………42 Figure 21 – Roller Clutch Assembly………………………. 43 Figure 22- Front Drive Axle Layout………………………. 44 Figure 23- Reverse Gearing………..………………………. 44 Figure 24- Rear Wheel Assembly with Brake System…….. 44

VI. Experimental Method……………………………………………………48 VII. Results…………………………………………………………..………49 VIII. Project Budget…………………………………………………………..51

Table 1 – Moonbuggy Budget………………………………51 IX. Project Schedule………………………………………………………...52 X. Discussion, Conclusions, and Recommendations………………………53 XI. References………………………………………………………………56 XII. Appendices

Appendix A – Competition Requirements Appendix B – Technical Design Drawings Appendix C – Finite Element Analysis Appendix D – Stress Strain Plot of A-arm Material Appendix E – Misc Suspension Calculations Appendix F – Misc Steering Calculations Appendix G – Misc Frame Calculations Appendix H– Misc Drive Train Calculations Appendix I – Task Breakdown and Gantt Chart Appendix J – Roll Analysis Appendix K – Detailed Budget Breakdown

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Introduction: (Chris) Background:

During the past four months we built and tested a two-person moonbuggy that

was raced in The Great Moonbuggy Race. The Great Moonbuggy Race is sponsored by

NASA and is held every year at the U.S. Space & Rocket Center in Huntsville, AL. The

moonbuggy is a human powered vehicle that is designed to overcome the same problems

engineers faced with the design of the original NASA moonbuggy. The moonbuggy

carries a male and female rider over a half-mile of simulated lunar terrain course

including “craters” , rocks, “ lava” ridges, inclines, and “ lunar” soil. In addition to

navigating the lunar terrain, the buggy also fulfilled dimensional requirements

Figure 1 - Map of Simulated Moonbuggy Course

as shown in Appendix A. The buggy was judged on how fast it maneuvered the

simulated obstacle course (Figure 1) as well how quickly was assembled from a collapsed

state.

The buggy is constructed mainly from aluminum. The male and female riders sit

in a back-to-back configuration to power the buggy. Each rider pedals bicycle cranks

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that run to a four gear internal hub via a chain. After the internal hub the power is

transferred to the drive axle via a chain. Each wheel is independent of the other due to

four roller clutch bearings. The suspension is provided by double A-arms on all four

wheels. A linkage attached to the bottom of the double arm with transfer the vertical

movement of the suspension to horizontal shocks pinned to a rocker welded on the frame.

To make the buggy fit into a four-foot cube it is hinged at the middle of the inverted

triangular frame. The front wheelbase is smaller than the rear to allow the buggy to be

folded in the middle. The buggy is directed by two steering arms at the front rider’s sides

and connected to a butterfly arrangement that connects to the hubs.

The design of the project used engineering knowledge from a broad area of

disciplines. This project incorporated the use of learned shop skills, mechanical

engineering curriculum, design techniques, software design tools, and team management

skills.

Project Objective and Scope of Work:

The top goal of the team was to secure the first place position at the Great

Moonbuggy Race in Huntsville, Alabama. Another goal for the team was to finish

construction of the moonbuggy by the end of February.

Scholastic goals for the team are to earn an A in our senior design class by

designing a capable moonbuggy while gaining usable skills for our future careers.

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Design:

Suspension (Cliff)

Conceptual Design:

A suspension was deemed necessary for the buggy for several reasons. First, with

a suspension the frame would be twisted less so it would not have to be as strong and

heavy. Second, with a suspension the wheels should be in contact with the ground at all

times. This makes sure that all the drive power goes to the ground and is not wasted. A

suspension would also make the buggy easier to maneuver while it is traversing the

course since all four wheels will remain on the ground. Thirdly, a suspension would also

make the ride more comfortable for the rider. With a more comfortable ride the rider can

more easily concentrate on pedaling.

A double A-arm suspension was chosen because it has proven performance, it’s

widely used, fits the design of our frame well, and is well within our ability to construct.

The buggy was designed to have eight inches of suspension travel. Two to three inches is

then taken up once the buggy is loaded. From the hub, an upright is connected the hub to

the rocker arm. The rocker pivots about the top member of the frame. The rocker uses a

mechanical advantage to get a large suspension movement out of a smaller shock travel.

The rocker is also connected to a bicycle shock mounted inside the frame. The

conceptual suspension assembly can be seen in Figure 2.

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Figure 2 – Suspension Conceptual Design

Analytical Method:

To start the suspension design the location and dimensions of the a-arms,

uprights, and hubs were determined. How the suspension was going to be mounted to

the frame was also determined. To design the length and location of all the a-arms a trial

and error approach was used. The design was constrained by the maximum width of the

buggy and the frame’s geometries. The key variables were the vertical locations of the

hub and where the a-arm attached to the frame. The hubs and the locations on the frame

were then modified until the proper camber angles were attained. The camber was set so

that as the suspension was flexed that there would be negative camber in the wheels.

This would keep the wheels vertical as the buggy rolled into curves thus keeping with the

maximum strength of the wheels.

Rocker Design

To design the rocker the location of the upright had to be known. The location of

where the upright was attached to the lower a-arms was determined by putting it into

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ProE and seeing the best location for it, so it and the upper a-arm would not interfere with

each other. The upright support location on the lower a-arm can be seen in Figure 3.

Figure 3 – Lower A-Arm w/ Support

With the location of the upright the entire side of the moonbuggy was drawn and the

wheel was rotated up eight inches along it’s axis of travel. When the wheel was at eight

inches the height of the upright was also recorded. Then a ratio of the movement of the

suspension to the shock was established as shown in Eqn. 1. This ratio was then

multiplied by the distance from the rocker pivot to the shock pivot on the rocker. This

yielded the distance from the rocker pivot to the upright pivot on the rocker. The rocker

geometry can be seen in Figure 4.

3.35.1

5

_

_ ==travelshock

travelupright Eqn. 1

To start out the suspension stress analysis the weight that was going to be

transferred to each wheel had to be determined. This was accomplished by adding the

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total weight of the buggy and the weight of the two passengers. Then that weight was

divided by four symbolizing the weight that was distributed to each wheel. After the

weight going to each wheel was established the weight was multiplied by a factor of

safety of two and a roll factor of two. The roll factor takes into account that as the

moonbuggy goes around a corner it will roll to the outside putting more weight to the

outside of the moonbuggy. A ProE model of the rocker can be seen in Figure 4. The

dimensions for the rocker can be seen in Appendix B.

Figure 4 –Rocker

After the force going to each wheel was established, the forces going to the a-

arms and the uprights had to be determined. The analysis was run only on the front

suspension because it had the longest members. It was reasoned that since the rear

members were shorter that the stresses would inherently be smaller so if a certain

diameter tube could be used in the front suspension that the same tube could be used in

the rear. To determine the axial loads in the suspension it was simplified into Figure 2

with the hub, a-arm, and upright joint being point one and the upper joint being point

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two. First the vertical component in the upright was assumed to be exactly that of the

force being applied at the wheel. Then moments around point one were summed as

shown in Eqn. 2 to get the axial load in the top a-arm. The horizontal force component of

the upright was determined by Eqn. 3 Forces in the x-direction were then summed to

get the axial force in the lower a-arm (Fx). The results were a 355.6 lbs. tension load on

the bottom a-arm, 60 lbs. compression load on the top a-arm and a 464 lbs. compression

load on the bottom a-arm.

( )( ) ( )( )xFinlbsinM .9.500.11 −=Σ Eqn. 2

y

x

F

F

y

x = Eqn. 3

Upright Design

Since the upright was loaded in compression the main concern is that the column

would buckle. Several different geometries were tried after calling around to local steel

suppliers and seeing what they were willing to offer at a student discount. O”Neal Steel

offered a ...8

5DOin with .

8

1in wall thickness of 6061 aluminum. The buckling equation

shown in Eqn. 4 was applied with an end constant of .25 and determined that the

aluminum was ok to use since it could support 870lbs. before buckling. A sketch of the

upright can be seen in figure 4 in the suspension assembly. The upright dimensions can

be seen in Appendix B.

2

2

l

EICPcr

π= Eqn. 4

where I is ( )44

64 ioDDI −= π

Eqn. 5

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A-arm Design

For the A-arm’s, tubular aluminum was attained in the shop at the University of

Evansville. However the dimensions and material properties of the material were not

known. Caliper’s determined the diameter to be .733in. A tension test was also done on

the aluminum. This determined the materials yield strength to be 27,000 psi. A stress

strain diagram of the aluminum can be seen in Appendix D.

For the upper a-arm buckling was the most likely failure for the member due to

the loading characteristics so Eqn. 4 was applied to the aluminum found in the shop. This

found that the critical load was 4790 lbs. which is much greater than the load applied.

For the bottom a-arm Eqn. 6 was applied to determine the stress on it. The stress was

determined to be 842 psi, which showed that it would not yield under the loading. As

shown above the a-arms are way over designed but we feel that the increase in weight is

justifiable due to the fact that the aluminum is free. A sketch of the suspension assembly

with the a-arms in their proper locations can be seen in Figure 5.

Figure 5 – Suspension Assembly

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A

P=σ Eqn. 6

Bolts

The bolts were analyzed assuming they were in double shear as they connect the

tie rod ends to the supports. Eqn. 6 was again used to yield a stress σ of 2640 psi on a

3/8 in. bolt on the upright where the most forces are applied, which is well within the

limits of any bolt. Where in Eqn. 6 P is applied load and A is twice the cross-sectional

area of the bolt due to the fact it is in double shear.

Hubs

Along with the a-arm’s yielding it is possible that the hubs might give before the

a-arms, so they must be made strong enough to withstand any loads applied. The hubs

were decided to be manufactured from .8

3in sheet metal found in the shop at the

University of Evansville. A bending analysis was run on the hubs using Eqn. 7. In the

end of the hub a rib was run up the center of it for added strength against bending at the

end since the mount for the rod end is cantilevered out. However this added another

equation, Eqn. 9 to find the moment of inertia of that part of the hub. Also the location of

the maximum stress, c was not easily obtained. Eqn. 10 had to be implemented to find it.

I

Mc=σ Eqn. 7

where I is

3

12

1bhI = Eqn. 8

2yx AdII +Σ= Eqn. 9

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A

Ay

ΣΣ

Eqn. 10

Along the front hub a steering arm extends towards the rear of the buggy that be

connects to the connecting rods going to the butterfly. The rear hub will also have a

similar rod extending from it but with a different purpose. It is used as a stabilization

technique so that the rear wheel will not toe in or out. The rod is connected to the top

rear a-arm via a connecting rod. A model of the hubs with flange bearings and stub

shafts can be seen in Figures 5a and 5b.

Figure 5a and 5b – Front and Rear Hubs (Respectively)

Along with doing paper calculation a finite element analysis was run on the front

bottom a-arm since it is the longest a-arm and has a compressive load applied at the

upright support. The resultant displacement and Von Mises stress FEA plots can be seen

in Appendix C, Figures C1 and C2 respectively. The analysis was performed on the ProE

model by meshing it in ProMechanica and exporting in to Cosmos/M. The thread hole

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surfaces (where rod ends are inserted) were constrained in the x, y, and z direction and

allowed to rotate about the frame support axis and the vertical and horizontal hub axis’

(x, y, z axis in the analysis). A 500lb load was applied at a 75 degrees to the horizontal to

simulate the maximum force applies through the upright (2” of compression for springs

with a k=250lb/in). The maximum Von Mises stress of 9290 psi was shown near the

support edges and the hub attachment end which is below the yield strength of aluminum.

The displacement analysis (Figure C2) yielded a maximum resultant displacement of

0.011” which well within acceptable levels. Sample calculations and other miscellaneous

calculations can be seen in Appendix E.

Prototype Design:

After all the parts were designed, they were assembled in ProE. See figure 5 for a

sketch of the moonbuggy. It was discovered that the original shock design positions

couldn’ t be fitted within the frame due to the drive train. So the shocks were then moved

to the top section of the frame where they will not inhibit the movement of anything.

Also the arms are separated by 12 in. at their centers on the frame. This was determined

to be the optimum separation as to use a minimum of aluminum and yet still be

structurally sound laterally. As for the cost of the suspension all the a-arms and the hubs

are free. The aluminum for the uprights was free since we obtained that from the shop.

Thirty two tie rod ends will be needed to make the suspension and at $5.30 a piece that

makes $169.60. Also the aluminum plating on top of the lower a-arms that the upright

mount will be welded to is also in the shop at the University of Evansville so is also free.

$15.33 was attributed to the various grade 8 nuts, bolts, and washers that were needed to

make the suspension work. Replacement shocks cost $91.80 and a few other

miscellaneous items were purchased. This makes the entire suspension cost $401. A

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detailed cost breakdown of the suspension can be seen in Table 1. A drawing of the a-

arms can be seen in Appendix A. A drawing of the hubs can be seen in Appendix B. A

drawing of the upright can be seen in Appendix B.

Because the reverse gearing went up higher than originally thought the shocks had

to be mounted on top of the frame instead of inside the frame. This caused the distance

between the frame and the rocker to increase. This allowed the shock to move past the

horizontal poison into the frame, thus locking the suspension inside the frame and

damaging our shocks. To remedy the situation new rockers were manufactured to reduce

the distance between the frame and the rocker. These are shown in Figures 6 and 7.

Even though this solved the problem it allowed for only 5 inches of suspension travel

rather than the 8 inches that was originally designed.

Another design change occurred due to the fact that our current hubs bent after

they moonbuggy was loaded and was tested. After they bent the rear hubs were CNC

milled out of a 1inch block of aluminum and extensions were welded on to keep the

overall dimensions the same except thickness. The front hubs had a support member

welded on one side to strengthen it. The rear modified hub can be seen below in figure 6.

Figure 6 – Rear Suspension Assembly

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Figure 7 – View of Front Rocker and Shock Assembly

Steering: (Cliff)

Conceptual Design:

Figure 8 - Conceptual Steering Design

The steering system contains three major components. First there are steering

arms at each side of the rider which the driver pushes or pulls to make the buggy go in

the desired direction then connecting rods connect the handle bars to the butterfly. The

butterfly changes the lateral motion to transverse motion to the wheels via another

connecting. The arrangement is made from machined aluminum for weight savings. The

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connecting rods are arranged so that one pushes while the other pulls. This makes

steering easier for the driver. The steering arms are located next to the front rider’s

sides. A pivot is placed such that the driver can get a mechanical advantage. There are

connecting rods at the ends of the steering arms to transfer motion to the butterfly. The

butterfly converts the longitudinal motion into transverse motion to the wheel hubs. This

arrangement can be seen in figure 8.

Analytical Method:

To start out a turning radius had to be decided upon. For the competition the

buggy must turn within a 20 ft. turning radius. However we felt we could better compete

if our turning radius was sharper than that. So we decided to do the maximum that our

universal joints would allow. After several tries a turning radius of 14 ft was decided

upon. This is taking into account that the inside wheel turns 29 degrees which was under

the 35 degrees that the universal joints would turn without tearing themselves up. After

the turning radius was decided upon, the different turning angles had to be determined.

As a vehicle turns in a corner the inside wheel must turn more due to the fact that it has to

travel a shorter distance. These angles are known as the Ackerman angles. To

determine the Ackerman angles, the different angles that the wheels must turn in the

same radius of the turning curve, the frame of the buggy is drawn from a top view where

the radius of the circle is at the middle of the rear axle. Then lines are drawn from the

center of the turning radius circle to the ground contact point on the wheels. Then lines

are drawn perpendicular to those just drawn from the center of the circle. The angle

between the latest drawn lines and the vertical position of the wheels are the Ackerman

angles. Next, the location of the pivot point where the connecting rods coming from the

butterfly to the hub are located. This is done by drawing a top view of moonbuggy and

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drawing two lines from the center of the rear axle to the ground contact points of the front

wheels. The connection must lie somewhere on that line to keep the Ackerman angle.

Steering calculations can be seen in Appendix F.

Steering arms:

The steering arms are positioned so that they would give the most ease to the

rider. They are placed at the rider’s sides and just in front of the hips. The steering

arms pivot just below the seat on the frame and will continue down another three inches

till a tie rod end connects it to a connecting rod and then ultimately connecting to the

butterfly. The scrap aluminum that is used on the a-arms is also used on the connecting

rods, steering arms, and part of the butterfly. Using equation 7 with a max force of 100

lbs and a factor of safety of three the stress was 14667psi which was well under the yield

stress of the aluminum. Again this aluminum is oversized but the price of the aluminum

could not be beat. The dimensions of the steering arms can be seen in Appendix B.

Iterations for the steering and other miscellaneous calculations can be seen in Appendix

F.

Connecting Rods to Butterfly:

The connecting rods connect the steering arms to the butterfly. One is in tension

while the other is in compression. In designing the connecting rods the most likely mode

of failure is due to buckling as aluminum will buckle before it will yield due to tension.

Thus the rod is modeled using Eqn. 4 as having free ends and no supports thus giving it

an end constant of .25. Solving for the critical load yielded 6171 lbs. which is well under

the 1200lbs. that is applied as figured above.

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Butterfly:

The butterfly design and location was accomplished using a trial and error

method. Various widths and lengths of the butterfly were tried along with various

distances from the axle. It was determined that the pivot point must be 4.5 in. from the

axle and the other dimensions of the butterfly can be seen in figure 9 and Appendix F.

The butterfly was CNC milled from .5in. aluminum that is in the shop at the University of

Evansville.

Figure 9 –Steering Butterfly

Connecting Rods to Hubs:

A second set of connecting rods connects the butterfly to the steering arms

coming from the hubs. These connecting rods oppose each other in forces. One is in

tension while the other is in compression. Again the free aluminum from the shop is used

in its fabrication. To determine the amount of load it could handle Eqn. 4 was utilized.

The rods were modeled as having free ends and having no supports thus giving them an

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end coefficient of .25. Eqn. 4 yielded a critical load of 6921 lbs. which was well under

the applied load of 400 lbs.

Prototype Design:

Steering arms were chosen because they are the least restrictive as they are at the

rider’s side there is no way they can inhibit the rider from pedaling. Actually the steering

arms are supposed to provide support for the rider. After the steering arms, the power is

transferred to the butterfly via connecting rods which turn the lateral motion of the

steering arms and connecting rods into horizontal motion that turns the wheels. The

butterfly is the only major conceptual change from the original conceptual design. The

double butterfly design was abandoned due to the fact that to keep it from interfering

with the drive train it would have to be placed outside the buggy frame which was an

undesirable characteristic. However during building the steering was again modified.

The first steering was found to not be stable enough. The second steering had torsion

resisting arms on the steering arms so they wouldn’ t torque out. For the butterfly a

bushing was lathed out so it was more stable. After the new stability equipment was put

on the steering worked beautifully. The new butterfly and steering equipment can be

seen in Figure 9a. However during testing it was discovered that the thin amount of

aluminum that was one either side of the bushing was not enough. So the lower half of

the steering arms were cut off and replaced by a plate of aluminum 1.5 in. wide. This can

be seen in Figure 9b below. There were several costs for the steering. There were eight

tie rod ends of which four were purchased which added up to $21.20. Bushings were

purchased for the steering arms which totaled to $8.40. Together the total cost of the

steering was $58.09. A detailed cost breakdown of the steering can be seen in Table 1.

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Figure 9a – Steering Butterfly and Connecting Rods

Figure 9b – New Steering Arms w/ Plate of Aluminum

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Frame:

Conceptual Design: (Chris)

In general, the frame was constructed using geometry and materials that minimize

weight, yet maintain sufficient structural integrity. The frame (Figure 10) will most

likely be constructed from chromyl steel or aluminum. The strength of steel will help

minimize the frame’s structural complexity and size of the frame such that the 4’ cubical

volume requirement is satisfied. By using steel instead of aluminum the beams can be

smaller and easier to weld. However if chromyl proves to be too expensive, and/or

aluminum is deemed strong enough by FEA analysis aluminum will be used. As viewed

from the front, the frame will appear as an inverted triangle (Figure 10). This design will

make the bottom member in tension and the top two in compression. Since steel has a

relatively high resistance to tension, only one member is needed on the bottom. The loss

of this member will help minimize the buggy’s weight. The frame will be hinged at the

middle to fold the buggy in half so that it can fit inside the four-foot cube.

The seats are mounted toward the frame’s center in a back-to-back configuration

with one seat on each half of the frame. The seat backs are foldable to minimize space

requirements. The pedals are mounted on the frame via a connecting rod that is also

foldable to help fit the buggy in the four-foot cube (see Figure 11).

Figure 10 – Conceptual Frame

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Figure 11 - Conceptual Frame Collapsed Figure 12 - Conceptual Frame Front

The frame is modeled in Pro-Engineer (ProE) and analyzed with finite element

analysis (FEA) software to optimize weight and strength. The FEA software will use

approximated applied loads (i.e. the passengers’ weight and suspension reaction forces)

and the frame’s geometry created in ProE to solve for stress. By looking at the key areas

of stress, the design will be improved by eliminating unnecessary material. The

numerical FEA solution will be compared to analytical results to measure their

consistency and accuracy.

Analytical Method: (Cliff and Chris)

To start out designing the frame the dimensions of the frame had to be determined

first before any force or stress calculations could be run on it. One of the rules of the race

is that the moon buggy must fit inside a 4 ft. cube. Thus the frame was designed to fit

into a 4 ft. cube while folded in half. To find the lengths of the frame a 4ft. box was

drawn and a 1 in. “cushion” layer was drawn around the inside of the box. The wheels

were then drawn in a far out as they could go. Then the frame was drawn in to the top of

the box. Then the axle to the middle of the buggy dimension was recorded and used as

the groundwork for all the later systems. The mounts for the suspension were put in so

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that there was ample room for the internal hub in front of the suspension and drive train.

Also the mounts were placed 12 in. apart at their centers as discussed in the suspension

analytical method. Also the rocker and shock mounts were placed in their respective

positions as the geometry dictated so they did not hit the upper a-arms. After the basic

outline of the frame was determined supports were put in to transfer the weight of the seat

and the a-arms.

The frame design can be seen in figures 13 and 14 below.

Figure 13 – Model of Front Frame

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Figure 14 – Model of Rear Frame

To do a force analysis all the a-arm forces and the seat forces were applied. Then

the forces in the support members had to be determined to find the actual shear and

moment diagrams for the frame. Eqn. 12 was used to find the respective forces in the

supports.

y

yx

F

F

y

22 += Eqn. 12

After applying Eqn. 14 the stress in the frame was determined to be 48,572 psi

which is under the 60,000 psi yield strength of the square aluminum tubing. As a

secondary measure a finite element analysis was run on the frame with the seat fully

constrained by constraining the 4 bolt holes where the seat will be bolted to the frame.

Then all the brackets that attach the suspension to the frame were loaded. The four

brackets that hold the a-arms were loaded with 550 lbs at and angle of 45 degrees

pointing downward and toward the frame. The rocker was loaded with 800lbs. pointing

up. The shock was loaded with 550 lbs. at 45 degrees pointing down and away from the

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center of the frame towards its respective bracket. The analysis yielded a max stress of

27050 psi just above the support for the seat. Thus through our analysis it was

determined that an aluminum frame would not yield and thus could be used instead of the

steel one that was proposed earlier. In Appendix C, Figures C3 and C4 show the Von

Mises stress distribution on the frame surface. Figure C5 shows the resultant

displacement of the deformed frame. The maximum resultant displacement was 0.011”

which is well within acceptable levels. The shear and moment diagram along with

sample calculations can be seen in Appendix G.

Prototype Design: (Matt and Tommy)

A complete dimensioned drawing of the frames is shown in Appendix B.

Through proper idealizations in the beginning and proper bracing and analysis it has been

determined that the frame that we have designed is light and strong enough to handle

anything that ourselves or the moonbuggy course can throw at it. Pictures of the

assembled moonbuggy with both frame halves can be seen in figures 15 and 15a.

Figure 15 – Moonbuggy Assembly

29

Figure 15a – Moonbuggy Assembly

A center hinge was designed to allow the moonbuggy to fold into the 4 foot cube. The

hinge is shown in figures 16 and 16a which displays the moonbuggy in its collapsed state.

Figure 16 – Collapsed Moonbuggy

30

Figure 16a – Collapsed Moonbuggy

The frame design requires that there be two latches that keep the front and rear

sections of the buggy connected and in constant contact during all portions of the race.

The latch was primarily used to keep the frame from it’s folded up position while riding it

around. Two latches were used to lessen torsion loads that would be applied to the center

hinge during riding. Taking in account minimizing the moonbuggy’s overall assembly

time, the latches were chosen based on how long they would take to close.

The design process of the hinge was one believed to be very crucial. It must be

able to withstand large torsion stress during the carrying of the collapsed and be durable

enough to withstand the tensile forces from the rugged terrain of the NASA obstacle

course. The initial design was to fabricate a simple hinge from steel tubing and plate,

although soon after fabrication was complete, it was clear to see that the given design was

not strong enough to prevent failure. The group ultimately agreed to purchase an

industrial strength hinge that was both large enough to support our triangular frame and

secondly to allow for fast and easy assembly from the 4 foot cube. The hinge that was

31

purchased was steel and was rolled around the pin. To prevent the steel from unrolling it

was welded together. The hinge was bolted to the frame in four different locations on

both front and back portions of the frame. To better support the hinge from torsion forces

“wings” were welded to the side of the frame, these doubled as extra points of contact for

the hinge to be attached to.

The NASA extras specifications were given in the rules for the race. The

complete list of rules is available in Appendix A. In short, NASA requires that teams add

specific extra parts that the real moonbuggy used in space. These extras include a

simulated TV camera (2”x3”x6”), a simulated antenna (diameter 2’), two simulated

batteries (4”x6”x8”), moon dust abatement devices, a US flag, and simulated control

panel (total combined size 1 cubic foot). The simulated control panel was broken up into

several different pieces to fit inside the frame.

The batteries, simulated TV camera, and antenna were constructed out of heavy-

duty poster board. These extras all took the shape of boxes that were then attached to the

frame with duct tape. They were constructed to be the exact dimensions NASA gave and

were painted black for cosmetic purposes. The total cost of producing the NASA extras

was $9.35 for the poster board, duct-tape, glue-gun, and glue sticks. The moon dust

abatement devices were simply fenders that could prevent dust and gravel from kicking

up from the course. These were placed over the front two wheels. They were made from

1/8” Lexan were approximately 2”x5” and were held to the hubs with wire. The tops

were then taped to cover the rough edges. The US flag was printed offline and was

laminated and taped to the back of the front seat. The simulated control panel was made

from Styrofoam and was broken up into three pieces that formed the required one cubic

foot. They were then spray painted black to match the other boxes.

32

The NASA extras were attached to open spaces in the frame. The batteries fit

under the front rider, just behind the steering arms. The two foot antenna was attached to

the back of the rear rider’s seat and was scored in the center to allow the antenna to fold

and fit in the four foot cube. The pieces of the control panel fit beneath the rear rider’s

seat and beneath the front rider’s seat. The NASA extras can be seen in figures 15a and

16a.

Due to the uncertainty about the riders, the seats needed to be made adjustable to

fit each rider perfectly. Boat seats were purchased at Wal-Mart and were then bolted to

angled mounting brackets that housed a sliding track for the rider to adjust the seats.

These mounting brackets used a 20° angle to give the rider the proper angle to reach the

pedals. The brackets formed the shape of an H and were made out of steel that was 1/8”

thick. The sliding tracks were salvaged off a previous moonbuggy and each track had a

small lever that fit just underneath the seat so each rider could adjust the distance to the

pedals while seated on the buggy. The seat mounting brackets were bolted to the frame.

The seat on the front faced forward, and the seat on the rear faced backwards. This

created the back-to-back seating arrangement that was designed.

For the restraining devices, nylon towing straps were used. These were bolted to

either side of the seat on the metal bracket that allowed the seat backs to fold down. The

buckles for the seat restraints were salvaged off a previous moonbuggy. The seats and

seat belts can be seen in figures 15a and 16a.

33

Drive Train:

Conceptual Design: (Chris)

Two separate axles power the four-wheel moonbuggy. The front rider powers the

front axle and the back rider powers the rear axle. By separating the drive axles, the

moonbuggy’s front and rear wheels are independent and allow the buggy to be folded

about its center. Each driver pedals a set of bicycle cranks that are supported about the

internal hub axis. The driven chain runs from the crank sprocket to the primary internal

hub sprocket. The internal hub transfers the motion to a secondary sprocket attached to

the hub’s rim. Another chain runs from the secondary sprocket to the driven sprocket

attached to the aluminum housing of the roller clutch assemblies on the drive axle. The

roller clutch assemblies act as a one-way clutch (much like a differential) that consists of

two housings bolted together, two separated inner shafts, and four drawn cup roller clutch

bearings (2 on each side). The roller clutch bearings are pressed into each housing bore.

Each bearing bore is fitted with a precision harden inner ring that is attached to its

respective inner shaft.

The inner shafts of roller clutch assembly are supported by stamped steel pillow

block bearings. The pillow block bearings have collars with set screws that disallow

axial movement of the inner shafts. The inboard universal joints are pinned to the ends of

the inner shafts protruding out of the pillow block bearings. The opposite ends of these

universal joints are keyed along their length. The keyed end of the inboard universal

joints mate with their respective hollow half shaft that has an undersized key stock

welded within its bore. This sliding or spline connection is required since the centerline

of the a-arms isn’ t aligned with that of the universal joints. As a result, the axle has to

lengthen during suspension travel. The opposite ends of the hollow half shafts are pinned

34

to an outboard universal joint. These outboard universal joints are positioned such that

their centerline falls within the pivot points of the a-arms. This is really only required on

the front drive axle since it has to rotate about those pivots when turning. Attached to

the outboard universal joint is a stub shaft. A disc is welded to the stub shaft which is

bolted to the wheel’s hub to transmit torque from the drive axle.

The rear and front axles are identical with two exceptions. The first is that each

half shaft on the rear axle is 2.25” shorter allowing the rear wheels to fit within the front

wheels during the moonbuggy’s collapsed state. The other exception has to do with the

way the chain is guided from the secondary internal hub sprocket to the drive sprocket.

The rear drive chain is run through a series of pulleys that reverses the rotation of the

drive sprocket with respect to the secondary internal hub sprocket. This pulley

arrangement allows the rear-facing rider to power the buggy forward pedaling

counterclockwise. Without this arrangement pedaling would have been very awkward

for a rear rider use to the traditional pedaling direction incorporated on all manufactured

bicycles. A sketch of this arrangement can be seen in Appendix B.

Standard bicycle parts are used for the sprockets, cranks, internal hub, and chains

since they were easy to purchase given the design parameters. The buggy’s crank

assemblies are located on the ends of the buggy due to the chosen back-to-back rider

configuration. Each crank assembly pivots about an axis to allow folding in and away

from the frame. By being able to fold the crank assembly out, the drivers have more

room to pedal. This also helps to maintain the chain’s tension since the distance between

the chain ring and internal hub remains constant. To maintain constant chain tension

during shifting, a single sprocket chain ring and an internal hub are used on both crank

assemblies. Shimano Nexus brand internal hubs were used; they house four internal

35

planetary gears that can be changed or “shifted” to change the gear ratio and thus the

speed of the buggy. Since the gears are internal, the hub eliminates fall off due to

shifting that can occur in a standard derailleur system.

The NASA guidelines do not require brakes, but for practical and safety purposes

a brake was added. Due to the differential on our driveshaft the team decided for our

purposes not to use a disc brake on our actual drive shaft since this would put extra forces

on the bearings. Since the drive axle support bearings are self aligning it was feared that

the extra forces from braking would knock the axle out of alignment. The team

brainstormed and thought braking only one wheel would be sufficient enough to stop the

buggy due to its light weight. It was also decided a typical mountain bike brake would

work sufficiently. To mount the brake to the buggy a bracket was fabricated and welded

to one of the rear hubs of the buggy.

Analytical Design: (Chris)

Crank Support Design

The crank supports were designed to withstand a 100 lb force applied along the

moonbuggy at the crank’s centerline as seen in Figure 17. The 100 lb load is an estimate

Figure 17 – Crank support loading

36

of the maximum force applied to the cranks while the buggy is in motion. The main

areas of concern were bending stress and buckling. To size the members to withstand

bending loads, moments were taken about the z-axis resulting in the reaction force RCx to

be 200lb and the moment Mc=4500inlb. By assuming that the maximum allowable

bending strength equals the yield strength Sy, the factor of safety n can be found by:

bending

ySn

σ= where

)(4

44

max

io

obending

rr

rM

I

Mc

−== πσ Eqn 13 & 14

We decided that CP6061 aluminum (Sy=60ksi) tubing would be used to construct the

support since it’ s a light, yet strong material. By substituting equation 13 into equation

14 and assuming 1” solid rod for CP, the following result for n is obtained:

31.1)5)(.*4500(4

)0500(.*60000(

4

)( 44

max

44

=−=−

=inlbin

psi

rM

rrSn

o

ioy ππ

The factor of safety for members AP and BP was found to equal 1.10 using .750” solid

rod. Although these factors of safety are acceptable, generally a factor of safety of two or

more is desired in design. We choose not to increase the diameters since cross supports

could be added later if needed to stiffen the assembly and lessen the bending stresses.

To size the members to withstand the critical buckling load Pcr a force analysis

was performed on the structure. The forces acting along the members was found using

unit vectors. For member AP the unit vector is:

kjikji

zyx

kzjyix

APAPAP

APAPAPAP 25.3807.565.

25.3155.10

25.3155.10(222222

+−−=++

++−=++

++=λ Eqn 15

The force acting along AP is a product of the unit vector and the applied force magnitude:

)(3257.805.56)25.3807.565.(100 lbkjikjilbFF APAP +−−=+−−== λ Eqn 16

37

FCP and FBP were found repeating this process and summing of the forces to zero in the x,

y, and z directions. The result found that that Fcp=-62lb (in compression) and

FAP=FBP=37.7 lb (in tension). The load in tension can be neglected due to the large

diameters required to resist bending.

The Euler critical buckling load can be calculated for member CP in the following

manner:

lbin

inpsi

KL

EIPcr 13700

))0.23(7(.

))50.(4

)(10*11(

)( 2

462

2

2

===

πππ Eqn 17

where I is the moment of inertia, E is Modulus of elasticity, L is the column length, and

K account for the end conditions (K=.7 since one end is pinned and other is fixed or

welded to the cranks). By observation FCP is much smaller than Pcr, therefore, the

buckling effects are negligible.

A FEA analysis was ran on the crank support’s ProE model in ProMechanica. The

resultant displacement and Von Mises stress plots Figures C5 and C6, respectively, are

shown in Appendix C. The support members were constrained as pin connections and a

100lb load was applied to the crank’s center. An initial analysis showed that the

maximum deflection on the middle two supports was an unacceptable 0.125 inches.

Cross supports where added to stiffen the pedal supports and reduced the maximum

deflection on the middle supports to an acceptable 0.032 inches. The pedal support

housing and bottom member showed a maximum deflection of ~0.063”, however, both of

these areas will strengthen due interior parts (i.e. internal hubs and mounting brackets)

that were not included in the analysis. The stresses in both models were similar with

maximum Von Mises stress of 124.8 kpsi occurring at the constrained inner surfaces.

38

This point was an assumed to be a meshing error since the average stress through the rest

of the support didn’ t exceed 27.0 kpsi.

Shaft Design The main load acting on all the drive train shafts with the exception of the stub

axle is torsion. The torsion stress τ action on a member can be found from:

)(4

44io

o

rr

Tr

J

T

−== π

ρτ Eqn 18

where T is the applied torque, ρ is the distance from the object’s center, and J is the

rotational moment of inertia. If the driver applies a maximum load of 100lb and the

crank length is 7” , the maximum torque acting on the drive shafts is 700in* lb. Assuming

the half shaft is hollow with inner and outer radii of .500” and .625” respectively the

torsion stress psiinin

ininlb6183

))500(..)625((.4

625.*700

44

=−

= πτ which much lower than the yield

strength of most metals. The factor of safety can be easily found by the ratio of the

material’ s yield strength and the torsion stress.

τyS

n = Eqn 19

The minimum solid steel shaft diameter to resist torsion can be found by rearranging

equation 18 and substituting equation 19 to obtain the form:

inpsi

inlb

S

Tnrd

yoo 25.0

)55000(

)1)(700(42

422 33 ====

ππ Eqn 20

The stub axles experience a shearing load V which could be as large as the weight

of a loaded buggy (Vmax=500lb) during normal operation. The minimum shaft radius

can be found by rearranging the shear stress formula

39

2

4r

V

A

Vπτ == into

πτV

r4= . Eqn 21

By assuming yield strength of 55 kpsi, the stub axle must have a diameter larger than

0.215” to rest shear.

inpsi

lbVr 108.

)55000(

)500(44 ===ππτ

The above calculation is shown in Appendix H. Combined alternating bending

and mean torque calculations for infinite life were also performed and are available in

Appendix H. Since we don’t really need infinite life they are to be considered as a very

extreme case.

Keyway Design

In the drive train design it is desirable to make the weakest points those that are

the easiest to replace. For the moonbuggy drive train the sliding keyway is a more

desirable weak point versus a more expensive and harder to replace shaft or universal

joint. While all shafts and universal joints were designed with factors of safety greater

than 2.5 the keyways were design to have a factor of safety of 1.5. The shearing forces as

Figure 18 – Keyway loading

shown in Figure 18 will generally crush the keyway before shearing occurs. To resist

crushing, the area of one-half the face of the keyway is used in which the keyway

thickness t can be found using the following formula:

40

ylS

Fnt

2= Eqn 22

wherer

TF = , T being the applied torque, and r is the shaft radius, n is the factor of

safety, l is the keyway contact length, and Sy is the yield strength. Assuming Sy=65ksi,

n=1.5, T=700 in* lb, and l=.375” the thickness t was found to be 0.23” or ¼”.

Gearing

Figure 19 – Gearing Spreadsheet

A gearing spreadsheet (Figure 19) was set-up to determine the sprocket sizes. A

full size copy is available in Appendix H. In this spreadsheet the secondary hub sprocket

B teeth, the crank sprocket C teeth, the drive shaft sprocket D teeth, and the drivers

pedaling speed are inputted and the moonbuggy’s velocity is calculated using known

variables and Equations 23 to 25. Equations 23 to 35 calculate the angular velocities of

each sprocket A, B, and C from gear ratio of the internal hub mB and the number of teeth

41

on each sprocket. The use of interchangeable sprockets with 30 teeth for B, C, and D

allowed the moonbuggy to operate a realistic velocity range of 14.2 mph to 26.1 mph

when the driver is pedaling at a speed of 90rpm.

Prototype Design:

Crank Supports (Chris)

From the analytical design and FEA results using 6061 aluminum the sizing of the

two top pedal support members was determine to be ¾” diameter rod. The required size

of the bottom pedal support member using the same material was determined to be 1”

diameter rod.

The crank support geometry was dictated by several parameters. The centerline

of the internal hubs had to be at least 2” above the frame and 3.5” from the front of the

frame to provide enough clearance for the sprockets and chains. The two top arms of the

support had to fit in between the internal hub axle and its support bracket. These arms

lengths were determined by assuming 17” of clearance for the cranks (distance between

the bottom of the rider’s shoe to crank centerline) and at least 42” of leg and back room

for each rider. The middle bottom member of the support had to be pinned to the frame

while in operating position and not surpass the 4 foot cube barrier when the buggy was

folded. Several dimensional analyses were done on graphical paper and ProE to obtain

the crank support lengths for the top members and bottom member to be 16.89” and

23.08” respectively. Inserts were designed to fit to be bolted to the top members so it can

be easily mounted on the internal hub axle. A drawing is shown in Appendix B.

Figure 20 shows a ProE model of the crank assembly. Its constructed prototypes

are shown in figures 15a and 16a. An additional support member was added to the

constructed prototypes to reduce the twisting deflection that occurred during testing.

42

Also, the bottom support member was shortened and mounted to the top front of the

frame to fit within the 4 foot cube. This was necessary since the purchased seats had

thick padding which didn’ t allow the pedal supports to rotate into frame as far as

originally expected. This design change did not adversely affect the pedal supports

performance. Actually, the reduced support length decreased the bending loads being

applied to the bottom member.

Figure 20 – Model of Crank Assembly

Internal Hub Bracket (Chris)

An internal hub bracket was designed to house each pedal support and internal

hub. Their simple design consists of a 4.0” x 2.25” 1/4” thick steel plate that is bolted in

two places to fit flush with the front interior surfaces of the frames. The 2.0”x2.75” 1/8”

thick plate was slotted along the center to position the internal hub 2” above the frame.

This piece was welded to the top of the ¼” thick plate. Once constructed, the edges of

the brackets had to be rounded to prevent interference with the bolts used to attach the

steel inserts to the pedal supports. Four brackets were used for the moonbuggy, two for

each internal hub. A dimensioned drawing is shown in Appendix B.

43

Roller Clutch and Drive Axle (Chris)

The primary factor that determined the geometry of the roller clutch assemblies

was the drawn cup roller clutch bearings. Each of these bearings is an assembly within

themselves consisting of an one-way roller clutch surrounding by two needle bearings

that resist radial loads. We chose to use two 1” bearings (rated at 412 in* lb each) for

each roller clutch assembly which cost $14.76 a piece. One 30mm drawn cup roller

clutch bearing assembly could have easily transmitted the estimated 700 in* lb of torque.

However, due to its larger size and unavailability, the 30mm bearings cost over twice as

much. The inner and outer shafts of the roller clutch assembly were designed

Figure 21 – Roller Clutch Assembly

around the roller clutch bearings. The housing was made out of aluminum to lighten the

axle’s weight. It was bored to 1.3125”Dia. and had an outer diameter of 2” to match the

housing specifications given by Torrington. The inner shaft to the roller clutch bearings

required tight tolerances to properly transmit torque. Instead of purchasing expensive

hardened precision shafts that are difficult to machine, the inner shaft design consisted of

a ¾” cold rolled shaft with two 1” diameter hardened inner rings. The hardened inner

rings fit tightly over the inner shafts and were held in place with bearing Locktite. The

44

inner rings were purchased at a cost of $5.65 each. Even with using a total of 8 inner

rings for both axles, this expense was well worth not having to machine a hardened 1”

shaft costing over $60.00 for both axles. The ¾” cold rolled shafts lengths were

designed to align the drive sprocket and fit outside of the drive axle support bearings.

Dimensioned drawings of the roller clutch and the drive axles are shown in Appendix B.

The drive axle assembly layout is shown in figure 22 and 23.

Figure 22 – Front Drive Axle Layout

Each drive axle contains two roller clutch assemblies, a 24 teeth sprocket, two stamped

steel pillow block bearing, four pin and block universal joints with boots, two hollow half

shafts, two stub shafts, and two wheel mount discs.

Universal Joints (Chris)

The universal joints were selected primarily by the manufacture’s torque rating,

the ability to connect them to their respect shaft, and the cost of the u-joints themselves.

It would have been very easy to choose a universal joint that cost $75.00 each, however,

since eight are needed this becomes cost prohibitive. Eight unbored 1”OD Curtis take-

apart pin and block universal joints were chose to transfer torque from the inner shafts of

the roller clutch. The lowest cost found for these universal joints was $33.04 each. Since

nine were purchased, one for a replacement, the total cost of the universal joints was

$297. These Curtis bearings allow for 35 degrees of misalignment instead of the standard

45

29 degrees offer with most pin and block u-joints. This allows for more unrestricted

movement in the suspension and a greater range of turning angle.

The take-apart feature of these universal joints was found to be most helped

during machining. The outboard universal joints were bored on one side to .625” ID to fit

over the stub shaft and left unbored on the side that fitted within the hollow half shafts.

The inboard universal joints were keyed ¼” x 1/8” along one side and bored to .625”ID

to fit over the inner shafts of the roller clutch assemblies. All connections, except for the

sliding keyed, where made using ¼” diameter roll pins. This allowed for easy assembly

and disassembly of the drivetrain. Boot covers were purchased for the outboard universal

joints to keep debris from entering the joint. The inboard universal joints were heavily

greased to ensure an easy sliding action with the hollow shaft key.

Half Shafts (Chris)

The hollow half shafts are made from aircraft grade chrome-moly tubing with

1” ID and 1.25”OD. One end of end half shaft has a ¼” x 1/4'” x 1.5” key welded into a

machined slot to fit the ¼” keyed pin and block universal joint. This connection was

designed to be long enough such that the half shaft could not fall off during normal

operation and allowed for lengthening required during suspension travel.

Stub Shaft (Chris)

The ¾” stub shafts were chosen to fit the purchased wheels, hub bearings, and

universal joints. The inner end of the stub shafts was turned down to 5/8” such that the

bored outboard universal joint could fit over the shaft and be pinned using a ¼” x 1”

stainless steel roll pin. This is larger than was required to resist shear, torsion, and

bending, however past moonbuggies have experienced problems in this area so caution

was warranted. The stub shaft drawings are shown with the hubs in Appendix B.

46

Wheels and Mounting Disc (Chris)

The mounting disc drawings are shown in Appendix B. They are designed to be

welded to the stub shaft and to be bolted to the wheel’s center hub. 26” utility tires rated

to withstand a 500lb radial load were chosen for the moonbuggy since they are of a

rugged design and fit our design parameters. They were purchased from Northern Tool

and Hydraulic at a cost of $28.00 each. These wheels are much cheaper than building

bicycle wheels to fit our design parameters which cost over $100 per wheel. The only

downfall with the chosen wheel design is that its spokes are in compression rather than

tension. Because of this, the spoke could bend rather than flex. If they bend the wheel

would be permanently weaken since the spokes couldn’ t return to their original position.

Shifters and Internal Hub (Chris)

Existing 4 speed Shimano shifters and internal hubs were salvaged off the old

moonbuggies. One shifter was attached to a front steering arm and another was attached

to the rear handlebar supports to allow each rider to easily change gears as the course

dictates. The installation of the shifting cord was found to be a tiresome process and had

to be done by trail and error. Also, the cord guide would become loose and allow the

cord to hit the frame when the pedal supports were folding in and out. The only way to

prevent this was to repeatedly tighten the bolt on the internal hub holding the cord guide

in between assemblies.

Gearing (Cliff and Chris)

The combination of gearing sprockets was changed from the original design of 30

tooth crank gear, 30 tooth secondary internal hub gear, and a 24 tooth drive gear. Since

the original design was deemed to require too much starting torque, the secondary

47

internal hub gear was reduced to 24 teeth. This made a large difference enabled the

riders to pedal sufficiently with a top speed around 24mph.

The reverse gearing arrangement for the rear drivetrain is shown below in figure

23.

Figure 23 – Reverse Gearing

Braking (Branden)

The team ran into a few problems while designing the brakes. The main problem

was to figure out how the brake bracket should be attached and still allow it to fold into

the foot cube. The bracket could not be attached at the top of the wheel or anywhere in

the front section of the wheel; it had to be attached on the back section so it would fold

correctly. The weight of the whole bracket concerned the team as well. Coincidentally,

the back hub it was going to be attached to was aluminum so the bracket could be welded

on using one inch square aluminum tubing. The actual brake would be bolted onto this,

and the rear driver would control the braking and with good communication between the

48

front and rear riders successful braking could be achieved. With the brake controlled by

the rear person it would allow the front driver to have more concentration on steering;

however. In the end the brake worked up to the team’s hopes and was sufficient enough

in stopping the buggy. The final picture of the brake can be seen in figure 24. There

were minimal expenses for the entire braking system since the aluminum for the bracket

and the brakes were all salvaged off of existing buggies.

Figure 24 – Rear Wheel Assembly With Brake System

Experimental Method: (Valerie)

Upon completion of the moon buggy, the team moved to the testing phase of the

buggy. The team was overall pleased with the testing results but came across a number

of small problems as well. The buggy achieved a speed of approximately twenty miles

per hour in each test run. At this speed the buggy was also able to overcome obstacles

with ease due to the suspension system that was designed for the vehicle. The riders

were also able to shift gears to increase their speed as well as brake to stop the buggy.

One of the most frequent setbacks during the testing phase was a problem with the

chain. During each test run the chain on either the front or the back half of the moon

49

buggy jumped off the sprocket that was attached to the shaft. This resulted in only one

rider powering the buggy. The chain seemed to jump due to a number of reasons. In the

rear frame, the chain was extremely loose, and the slack in the chain allowed the chain to

come off the sprocket. The pulleys and sprockets could also have been out of alignment.

These problems were fixed by correcting the alignment of the pulleys and shortening the

chain by removing a full length and replacing it with a half link.

The other major setback during the testing phase was the steering. While

traversing one of the obstacles in the parking lot where testing occurred, the left steering

arm snapped. This was due to the thin wall of aluminum that was left after inserting the

bushing to go around the rod that went through the frame. The force exerted on the arm

when the buggy went over the curb was too great for the thin wall to support and the arm

snapped. This problem was overcome by replacing the aluminum rod with a thick piece

of aluminum plate that had a bigger wall around the bushing.

With these slight modifications the buggy was in working order and ready for the

national competition at NASA.

Results: (Valerie)

On April 12, 2003, the team took the finished moon buggy to the national

competition at the Marshall Space Flight Center in Huntsville. The preliminary parts of

the competition included the assembly time and the preliminary weight test. The team

successfully fulfilled the weight requirement and the riders were able to carry the buggy

twenty feet without dropping it. In the assembly time competition, the riders were able to

assemble the buggy in a time of 33 seconds. This result also includes the time for the

riders to be seated and buckled into the seats. The actual time was practice at 15 seconds,

50

however, one of the riders had difficulty in latching the seat belt during competition. Still

this was one of the fastest assembly times.

At the national competition in Huntsville, Alabama, the buggy started the race

with high hopes but did not perform up to the expectations of the team. In the first heat

of the race, the riders began the race and were able to pass over the first two obstacles

with the speed that was necessary to finish the race. However, the front right wheel’s

spokes popped out from the rim on the third obstacle of the race, making the driver lose

control of the buggy and run aground. The collapse of the wheel forced the riders to

leave the course because the moon buggy was inoperable with only three working

wheels.

The team was able to replace the wheel with a spare that they had with them, but

the students in Huntsville soon realized that other damages had occurred because of the

wheel failure. When the wheel collapsed other problems arose that affected the buggy.

The steering system, which was located in the front frame, was adversely affected

because the butterfly bent and the tow was bent out of line. Also the drive train bent into

a V-shape with the point of the V pointing towards the pedal support. The students at the

competition attempted to fix the steering by welding a piece of aluminum on the frame to

stop the steering arms from turning too sharply. This prevented the buggy from

overturning when turning to the left. The tow was also corrected. Unfortunately, the

drive train was nearly impossible to repair. Tommy Woods used brute force to push the

front roller clutch assembly into the correct alignment, but there was a chance that the

suspension system would not be able to work as it was properly designed because of the

damage to the buggy. This fear was realized when the team began the second heat in the

afternoon at NASA. The buggy was able to continue to the fourth obstacle before the

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front left shock failed. This shock had been previously damaged when it was improperly

installed during testing causing the striping of threads that hold the upper portion of the

shock together. This shock was thought to be o.k. if was installed in the correct position.

The damage that occurred during the first run cause a misalignment which attributed to

the shock being pulled apart and getting logged within the frame. Once logged in the

frame the forces acting on the rockers from the uprights created a large bending moment

which spread the frame apart resulting in further damage to the left side of the

suspension.

This failure occurred in many areas, but each failure was related to the front right

wheel that had been damaged in the first heat and the defective shock. Primarily, the eye

and socket bolts that attached the uprights to the rockers and the top A-arm to the knuckle

sheared. This resulting damage made the buggy immobile after the fourth obstacle in the

second heat.

Because of these failures, the team was unable to finish either heat and did not

place in the Huntsville competition.

Project Budget: (Chris)

Moonbuggy Project Budget BUGGY COMPONENT COST FRAME $122 SUSPENSION $401 STEERING $58 DRIVELINE $732 CRANK & SHIFTING SYSTEM $177 WHEELS / HUBS / BRAKE $117 SEATS $72 NASA EXTRAS & SAFETY GEAR $79 SHIPPING $117

TOTAL $1,875 Table 1 - Moonbuggy Budget

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We had $2000 of funding from the Indiana Space Consortium (ISC) grant to

spend on parts, tools, labor, and shipping. Donated materials, available shop materials,

existing moonbuggy parts lowered the suspension, frame, and steering costs

considerably. The majority of $540 received from the Student Government (SGA) was

spent to cover travel and lodging expenses for the Huntsville, AL competition. We were

able to remain under budget spending a total of $1,875 as seen in Table 1. A detailed

breakdown of part cost for each component system in shown in Appendix F.

The majority of budget was spent on driveline components totaling $732. The

total drive train (driveline with crank and shifters) cost $909 almost half of the total

budget. The next largest area was the suspension system in which we spent a total of

$401. This was including an extra set of shocks purchased in case one of ours should

fail. The other main expenses were the steering system and shipping cost which were

$116 and $117, respectively.

Project Schedule: (Cliff)

The schedule for the project can be seen in Appendix I. A detailed task

breakdown (Figure I.1) and Gant chart (Figure I.2) lists the individual project tasks, their,

estimated start/completion dates, and estimated man hours.

During the design phase, the moonbuggy fell behind schedule. It took longer to

design the moonbuggy than expected. This pushed the fabrication stage two weeks

behind schedule before we even started it. Although extra effort was put in the following

months those two weeks were never gained back due to additional delays. Initially it was

discovered that welding aluminum was harder than expected and it took a while for Cliff

to get the hang of it. In getting used to welding he had to redo several pieces. The

steering took longer than thought because the first steering system was determined to be

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too unstable. The current system was then modified to become more stable. The

suspension took longer due to the initial modification of the shock mount location. This

caused us to make new rockers. In making the drivetrain, the roller clutch assembly was

found to be extremely difficult to manufacture due to the close tolerances of the

equipment. Another fact that caused us to be behind schedule is due to the fact that we

had limited shop time. We had to work around our classes and other agendas we might

have. Despite OD’s best effort to keep the shop open he could not keep up with our

demands.

Discussion, Conclusions, and Recommendations: (Branden and Valerie)

The dimensions for boring out the roller clutch housings to fit on the outer surface

of the roller clutch bearing have very small tolerances which were hard to attain with the

shop equipment here at the University of Evansville. Hiring an outside contractor to

machine out the housing’s required tolerances would save valuable time. While the

reverse gearing system worked as expected, it implementation required us to change

some suspension geometry which in turn reduce the amount of overall travel of our

suspension. The inner space in the rear frame needed to be increased to avoid this design

problem. Since the frame was the first thing built, design changes couldn’ t be

implemented in time.

A shock tower instead of the rocker would have been easier to design and

construct. A rocker was used because it was decided that using a rocker and upright

would allow us to use smaller a-arms. The 1.25” square 1/8” thick aluminum tubing

supplied from the school would more than likely been able to handle the bending

moments that are associated with the shock tower arrangement. Although a shock tower

design doesn’t has the inherent suspension travel obtained from a rocker, large

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motorcycle shocks could have been purchased to get the targeted 6 to 8 inches of

suspension travel. The motorcycle shocks would weighted more than the used bicycle

shocks, however, they would have been much cheaper to purchase or replace.

To prevent flipping the people were placed as near to the ground as possible.

When completely unloaded the lowest part of the seat will sit 20 in. off the ground and

when loaded it will be 17 in. off the ground. The two inches are cushion room as it

would be better to take a 30 sec. penalty than be completely disqualified from the race.

The center of gravity of the moon buggy was determined by use of Eqn. 23. The center

of gravity of the riders was approximated at 9 in. above the seat and the center of gravity

of the moonbuggy was given by ProE to be 2 in. below the top of the frame. The weight

of the moonbuggy was figured by using ProE to attain the volumes of any solid in the

model and then multiplying it by its respective density. This yielded a weight of 147 lbs.

for the buggy. After applying Eqn. 23 the center of gravity was determined to be 19.8 in.

above the ground and after drawing a model, shown in Appendix J, it was shown that the

moonbuggy would not flip over on the 30 degree incline and there was some room to

spare.

W

WyGC GC

ΣΣ

= .... Eqn. 23

Brackets and tie rod ends should not be used to mount the suspension to the

frame. Welding on the brackets to the frame uses up too much time. Also the tie rod

ends and the fasteners add a lot of unnecessary weight to the buggy.

Even though our buggy did not finish the race, I think our buggy was engineered

well and was very capable of placing high, we just had bad luck. Obviously the wheel

collapsed and we should have replaced it. The wheel collapsing caused all of our

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problems with the front drive train and suspension. I do think however a few

improvements in some areas would make it even better. I think stronger mountain bike

rims would work better and prevent the problems we had. Lower gears would also be

very nice because starting off was kind of hard, but it was not impossible. Our assembly

time was not horrible, but a spring loaded latch, as well as doing away with pins in the

pedal supports would make assembly time a lot better. I would also run longer brake lines

to the front driver so they could brake. I would also recommend redoing the front wheel

hubs. The final and most drastic change I would have to recommend would be flipping

the front and rear drivers’ roles somehow to make the steering be on the back so the

steering would not be subject to such impacts as our steering was. (Brandon)

This design was one of the most intricate designs I have seen in my two years of

experience in the NASA Great Moonbuggy Race. However, there were several flaws

that I think could improve this design and turn it into a winning moonbuggy for next

year. If the wheel had not collapsed at the outset of the first heat, this design would have

proved to be extremely efficient and reliable. However, because the collapse of the

wheel caused so much other damage to the buggy we were unable to finish the race. We

could design and fabricate a spring loaded hinge that would decrease the assembly time

for the preliminary competition. Also the suspension system will need to be improved by

changing the shocks and perhaps even changing the location of the shocks. If funds

allow, the wheels could be replaced for stronger rims and more tread on the tires. The

gearing system could also be redesigned to prevent the shifting wire from interfering with

the movement of the chain.

This project could have been improved by finishing the buggy in a more efficient

manner and using the extra time to test the buggy. If this had happened all of the flaws

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that became visible in Huntsville could have been noticed earlier and repaired correctly.

Overall, the final product produced by the Global Moonbuggy team had very few design

changes from the conceptual design produced in the ME 495 class.

References:

Hibbeler, R.C., Mechanics of Materials, Prentice Hall, Upper Saddle River, New Jersey, 2000 Shigley, Josheph E., Mechanical Engineering Design, McGraw Hill, Boston, 2001