The Great Moonbuggy Global Team 2003

Transcription

1 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

2 Abstract: (Chris and Cliff) For the past 120 days we have attempted to build a moonbuggy capable of winning the 10 th 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. 2

3 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, 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. 3

4 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 Figure 5 Suspension Assembly 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 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 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 4

5 Figure 18 Keyway Loading 39 Figure 19 Gearing Spreadsheet.. 40 iii. Prototype Design Figure 20 Model of Crank Assembly.. 42 Figure 21 Roller Clutch Assembly. 43 Figure 22- Front Drive Axle Layout. 44 Figure 23- Reverse Gearing 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. XII. References 56 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 5

6 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 6

7 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. 7

8 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. 8

9 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 9

10 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. upright _ travel shock _ travel = = 3.3 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 10

11 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 11

12 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 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. ( 1in. )( 500lbs. ) ( 9in )( ) Σ M =. Eqn. 2 1 F x x = Eqn. 3 F y y 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 5 1 offered a in. O. D. with in. wall thickness of 6061 aluminum. The buckling equation 8 8 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. F x P cr 2 Cπ EI = Eqn. 4 2 l 4 4 where I is I ( D D ) = π Eqn. 5 o i 64 12

13 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 13

14 P σ = Eqn. 6 A 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 3 were decided to be manufactured from in. sheet metal found in the shop at the 8 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. where I is Mc σ = Eqn. 7 I 1 bh 12 I = 3 Eqn. 8 I = ΣI + Eqn. 9 2 x Ad y 14

15 ΣyA ΣA 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 15

16 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 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 $ 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 16

17 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 17

18 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 18

19 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 19

20 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. 20

21 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 21

22 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 $ Bushings were purchased for the steering arms which totaled to $8.40. Together the total cost of the steering was $ A detailed cost breakdown of the steering can be seen in Table 1. 22

23 Figure 9a Steering Butterfly and Connecting Rods Figure 9b New Steering Arms w/ Plate of Aluminum 23

24 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 24

25 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 25

26 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 26

27 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. F Fy x y = Eqn. 12 y 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 27

28 center of the frame towards its respective bracket. The analysis yielded a max stress of 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 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 28

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 29

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 30

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 heavyduty 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. 31

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. 32

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 33

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 34

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 35

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 R Cx 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: S y Mc M max ro n = where σ bending = = Eqn 13 & 14 σ bending I π 4 4 ( ro ri ) 4 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: n = S π y ( ro ri 4M max r o 4 ) (60000 psi * π ( = 4(4500in * lb)(.5in) ) = 1.31 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 P cr 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: x APi + y AP j + z APk (10.5i + 15 j k λ AP = = =.565i.807 j k Eqn x + y + z AP AP AP The force acting along AP is a product of the unit vector and the applied force magnitude: F = Fλ = 100lb(.565i.807 j k) = 56.5i 80.7 j 325k( lb) Eqn 16 AP AP + 36

37 F CP and F BP 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: 2 6 π 4 2 π (11*10 psi)(.(50in) ) π EI P 4 cr = = = 13700lb 2 2 ( KL) (.7(23.0in)) 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 F CP is much smaller than P cr, 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 inches. Cross supports where added to stiffen the pedal supports and reduced the maximum deflection on the middle supports to an acceptable 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 kpsi occurring at the constrained inner surfaces. 37

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: Tρ Tro τ = = Eqn 18 J π 4 4 ( ro ri ) 4 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 700inlb *.625in torsion stress τ = = 6183psi which much lower than the yield π 4 4 ((.625in) (..500in) ) 4 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. S y n = Eqn 19 τ The minimum solid steel shaft diameter to resist torsion can be found by rearranging equation 18 and substituting equation 19 to obtain the form: 4Tn 4(700inlb)(1) d o = 2ro = 23 = 23 = 0. 25in πs π (55000 psi) y 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 38

39 V V τ = = into A π 2 4 r 4V r =. Eqn 21 πτ By assuming yield strength of 55 kpsi, the stub axle must have a diameter larger than to rest shear. 4V 4(500lb) r = = =. 108in πτ π (55000 psi) 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: 39

40 2Fn t = Eqn 22 ls y T where F =, T being the applied torque, and r is the shaft radius, n is the factor of r 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 m B and the number of teeth 40

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 and 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. 41

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 /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. 42

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 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 43

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 takeapart 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 44

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. 45

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 46

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 47

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 48