Atv Design Report

TEAM ID60000 BAJA SAE INDIA 2012 DESIGN REPORT Copyright 2009 SAE International TEAM THE CONRODS AUTHOR K. SUBHASH BABU. CO-AUTHOR KARN JAIN. ABSTRACT The intents of the mini-Baja arguing be to design and manufacture a fun to drive, versatile, safe, durable, and high performance sullen road fomite. Team members must chink that the fomite satisfies the limits of restore rules, while also to generating financial support for the project, and managing their educational responsibilities. This vehicle must be capable of negotiating the nearly extreme terrain with confidence and ease.The 2012 SRM UNIVERSITY Mini-Baja Team, THE CONRODS met these objectives by dividing the vehicle into its major comp whizznt subsystems. By examining the 2011 entry, the team was able improve on many design features to better meet the stated requirements. Function Diagram (QFD) to determine which parameters were the almost critical. These key parameters ranging from most critical to least critical a re safe, reliability, first wagon train be, ease of ope symmetryn and maintenance, and overall performance. TECHNICAL SPECIFICATIONSENGINE Type Displacement contraction proportionality Max Power Max Torque DRIVE TRAIN Transmission Gear Shift Mechanism SUSPENSION Front Suspension overturn Suspension Ground Clearance Shocks and Springs Front Susp. Travel Rear Susp. Travel WHEELS Front Tyres Rear Tyres BRAKES Wor pansy Fluid Type Pedal Ratio M C Bore Dia W C Bore Dia Brake Disc Dia STEERING Type Mechanism Steering Ratio Lock to manoeuver move 4 Stroke, OHV,B&S 304 cc 81 7. 5 KW 3600 rpm 18. 5 Nm 2600 rpm Mahindra Champion Alfa (forward Orientation) Sequential repeat Wishbone Double Wishbone 11. ines Customized 5 marches 6. 5 inches 22*8-10 22*8-10 Dot-3 Oil All turn over Disc 41 0. 8 inch 1. 6 inch 6 inch Ackermann Rack and Pinion 10. 71 400 INTRODUCTION&CONSUMER INFLUENCES THE CONRODS BAJA SAEINDIA vehicle is designed as a figure of speech for manufacture by an revea ldoor recreation firm. The paragon vehicle is safe, simple and inexpensive. Additionally, the vehicle is attractive to potential buyers in both(prenominal) its visual appearance and performance. These characteristics are considered in design of the following major vehicle subsystems frame, disruption, steering, and braking.Before any design could begin, we had to understand exactly who our customers are and their needs. To gain this understanding, we did extensive research that acceptd food market survey and interviewing both professional and nonprofessional local off-road enthusiasts. With this research, we determined that our customers are the BAJA SAEINDIA event and non-professional weekend off-road enthusiasts. We felt it necessary to distinguish between the cardinal to ensure that we followed all rules set by SAEINDIA INDIA and to accommodate the weekend off-road enthusiasts in a safe manner within the SAEINDIA rules.With all necessary design parameters determined for ea ch(prenominal) customer base, we were able to combine them for an overall list of design specifications that met all SAEINDIA requirements. We used these parameters to create a Qualitative 1P ag e Turning universal gas constant 2. 7 meters CHASSIS/OVERALL DIMENSIONS Chassis Material IS 3074 CDS1 Tubular Frame Overall Length 2100 mm Wheel Base 1490 mm Wheel Track 1143. 2 mm Height of fomite 1520. 0 mm WEIGHTS Front Wheel Assembly 10 Kg Rear Wheel Assembly 11. 8 Kg Engine(with engine oil) 23 Kg Transmission(with 17 Kg lubricant) Chassis 55 Kg Dampers 8 Kg Expected curbing Weight 260 KgTARGET SPECIFICATIONS Parameters Speed Stopping Distance Acceleration Gradability Turning circle dia. Ground Clearance Emissions Values 40 km/h 7m 11. 6 seconds 82. 2% 5. 4 m 11. 6 inches BS III the planes created by the stray cage and the drivers helmet. SAEINDIA also require a 3 inch envelope when a straight-edge is applied to any two tubing. Emphasis was placed on creating an easily manufactured roll cage with few parts, minimal conjoin and yet is still both light and strong, hence the numbers of bends were unploughed to a minimum.Roll hoop Overhead members and anterior Bracing Members are one continuous bent tube. Lower Frame slope tubes are straight and are bent inwards to connect to the search foramen mounts. The Side relate Member is a single tube with a single bend that encompass the car from the Rear Roll Hoop forward. The foot box of the vehicle is shaped by the LFS, SIM and straight tubes welded to the upper side concern tube forming a hexagonal front bulkhead taking into consideration the time out design and drop-off in dead space based on experience from the 2011 entry. A 3-D view of the car is shown beneath FRAME DESIGNOBJECTIVE & FRAME CONFIGURATION The objective of the haoma is to encapsulate all Components of the car including a driver efficiently and safely. With a limited amount of power, the focus is primarily on the power to pack ratio of the vehicle. The only means to improve this critical parameter is to reduce the overall vehicle fish. Great care is taken in laying out the chassis. SAEINDIA requires each vehicle conform to a 95percentile male for all ergonomic evaluations of the design. The pertinent information is taken from Body-space Anthropometry, Ergonomics and Design by Stephen Pheasant.Several key safety factors in the design process dictate chassis roll cage layout and foot box design. For the roll cage, SAEINDIA requires 6 inches of clearance measured from the inside of star aspects of the chassis focused on during the design and implementation included driver safety, suspension and drive-train integration, structural locatedity, weight, and operator ergonomics. The number one priority in the chassis design was driver Page 2 safety. With the help of the 2012 Baja SAEINDIA Competition Rules and Finite Element Analysis (FEA), design assurance was able to take place.Rear Impact succeeding(a) nominate, imp act analysis was done while assuming 15,000N as the impact force. melodic line SMX-172. 22 N/mm2 FOS2. 43 MATERIAL SELECTION Two existents were considered for the eddy of the chassis AISI 4130 and IS 3074 CDS 1. IS 3074 CDS 1 steel with an OD of 25. 4 mm and a wall thickness of 3 mm was chosen because it exceeds the bending cruelty and loudness requirements of SAEINDIAINDIA which gives increased protection to driver. PROPERTY Tensile strength(N/sq. mm) Yield strength(N/sq. mm) Elongation on 50 mm G. L Density (g/cc) IS 3074 438 376 32% 7. 872 AISI 4130 760 460 27% 7. 5 Side Impact The next step in the analysis was to analyse a side impact with a 5000N load. As a side impact is most likely to occur with the vehicle being hit by another MiniBaja vehicle it was assumed that neither vehicle would be a fixed object. STRESS 237. 49 N/mm2 FOS 1. 77 It was found out that the bending stiffness and bending strength of IS 3074 CDS are greater than those of 1018 steel having a circular fa ll guy section of 25. 4 mm and 3 mm thickness LOADING ANALYSIS To properly dear the loading that the vehicle will encounter, an analysis of the impact loading seen in the various types of impact scenarios was required.To properly model the impact force, the deceleration of the vehicle after impact is generally assumed to be zero. To approximate the worst case scenario that the vehicle will see, research into the forces the human body can endure was completed. It was assumed that this worst case collision would be seen when the vehicle runs into stationary, rigid object. Front Impact The first analysis to be completed was that of a front collision with a stationary object. In this case a deceleration of 20,000N was the assumed loading. STRESS SMX-177. 81 N/mm2 FOS 2. 36Rollover Impact The Final step in the analysis was to analyse the try on on the roll cage caused by rollover with a 5000N load on the cage. The Loading was applied to the two upper forward corner of the perimeter ho op with a combination sender sideways and downward. The load was chosen to be on two corners as this would be a worst case scenario rollover. STRESS 267 N/mm2 FOS1. 57 forum To maximize the geometrical consistency of the fabricated chassis, all fixturing and measurements were based on a single fixed coordinate system relative to a rigid table on which the chassis and all components were bolted.Through the use of this table and good fixturing practices, the team was able to best assure that the chassis geometry, especially Page 3 in critical sections such as the suspension pickup points, correlated closely with the design specifications. In addition, measuring from aFixed location minimized tolerance stack-up referable to measurement error and component movement results. We stimulate decided to fabricate the second hub since it has minimum weight and optimized FOS. *Material Used to manufacture the hubs- postgraduate Carbon vane *Hardening Process Done-Cyaniding SPACE IN DRIVER COMPARTMENTDRIVER EROGONOMICS Driver ergonomics has been our major concern during design of the frame and also during positioning of various systems in drivers cabin. Cabin is made spacious for safe and reliefable. All the cables and wires are routed properly so that they would not interfere with driver legs or hands. all the routings are done in design stage itself and ROH is raised to a suitable height so that it would give proper vision to the driver DRIVERS VISION window SUSPENSION fair game A Mini-Baja suspension system must satisfy the following design requirements.Control movement at the wheels during vertical suspension travel and steering, both of which shape handling and stability. Provide sufficient sprung mass vibration isolation to maintain satisfactory jaw quality, while maintaining high tire-ground contact rate and low tire vertical load fluctuation rate to improve road holding and handling. Improve jumping performance by limiting sprung mass sales talk displac ement while the vehicle is airborne. Limit chassis roll during cornering to keep roll-over, decrease roll camber, and therefore, decrease steering reaction time and slip fee induced drag forces.Prevent excessively high jacking forces by managing static roll center location and roll center migration. Limit lateral tire scrubbing to maintain straight line stability and minimize horsepower losses at the base suspension. Control lateral load transfer distribution to influence both steady state and limit of adhesion over steer/under steer handling characteristics. The non-professional weekend off road enthusiast requires a vehicle which exhibits both safe, stable, responsive handling and a soft, comfortable ride . DRIVERS VIEW OF THE CABIN Alternatives consideredSeveral different types of suspension system were considered before selecting the mugwump unequal arm double wishbone suspension system for both front and rear. Unequal double A-arm In the design, suspension is supported by t riangulated Aarm at the top and bottom of the knuckle. Advantages *Improved ride quality *Good road holding *Rigid links *More control over geometry *Wheel control is precise *Negative camber gain during vertical suspension travel. Page 4 FRONT SUSPENSION Setting static roll Centre A two dimensional sketch was made after estimating the Centre of mass of the vehicle on paper.Various references were taken to make a 2D sketch these include ? ? ? ? Track width of vehicle Front hub king pin axis inclination, king pin length, ball joint dimension Rim off set(for king pin positioning) Wishbone mounting point lengths rebound. Since we could not find springs that were less stiff than this we decided to go for the Auto springs as it satisfied our ride comfort requirements. A stiffer spring was required in the rear to achieve the coupling effect of suspension so as to convert the pitching action into a bouncing motion. fundament SUSPENSIONThe primary concern in conception the rear suspensi on was to abide the maximum possible travel (jounce and rebound) such that the rear driving wheels were always kept in contact with the ground. The camber metamorphose in the rear wheels should be such that there is not much appreciable change in camber throughout the travel of the wheel. The other factor taken into account was that we were having things with the rear suspension in last years design as it was observed that the drive shaft coupling was coming in contact with the lower wishbone in the rebound condition and this issue has been addressed and rectified in this years design.The rear suspension wheel rate was fixed such that the inborn frequency of the rear suspension is 20% greater than the front suspension thus providing a flat ride over bumps by converting the pitching motion of the vehicle to be born-again into bouncing motion. DAMPER SELECTION Method for selecting springs The process began by selecting an appropriate wheel rate for the front axle. A typical road frequency of 3. 7 Hz may be encountered at the competition. This is based on a vehicle urge of 40Km/h and a road surface with bumps spaced 3m apart. The natural frequency of the suspension should be kept well below 3. Hz in order to avoid any unwanted excitation. A front suspension natural frequency of 1. 20 Hz was deemed to be suitable. The wheel rate required to obtain this natural frequency was established using the following equating (assuming sprung mass of 72kg/wheel) . 2 ? ? We need to calculate the damping ratios for the front and rear suspensions. The design process will commence by iteration only. First we find the ratio for sprung and unsprung with respect to the model. Sprung mass was found to be 71. 456kg the sprung weight was determined while the sprung mass was 288. 54kg. The ratio is 0. 247. The natural frequency of the front suspension is set at 1. 2Hz. Weight on each front wheel is 57. 71 kg. The max force of damping is given by Fcd =2*Msp*wn. Critical damping for ce for the front suspension system is 1085. 73 Ns/m. For the un-sprung mass natural frequency would be Wn=((Ks+Kt)/Ms)0. 5 The combined stiffness of tire and wheel is 53. 24N/mm. Amplitude ratios were calculated for a retch of damping ratios. These amplitude ratios represent the ratio of applied displacement and the displacement that actually reaches the sprung mass.Amplitude ratios were plotted against the ratio of applied frequency and natural frequency of the sprung mass. This graph shows the nonsuch damping ratio that should be used. This value as obtained from graph is 0. 7 which gives a damping co-efficient value of 760 Ns/m. In the similar manner the rear suspension has a ride rate of 1. 56Hz. The critical damping force is 1960 Ns/m. The graph of amplitude ratio vs frequency ratio shows an ideal damping ratio of 0. 7 the damping co-efficient is = 0. 7*1960=1372 Ns/m. ? fn ? k wheel msThe ideal wheel rate for the front suspension was calculated to be approximately 40N/mm. Th e relationship between wheel rate and motion ratio (MR) was used to generalise the location of the shock actuation point on the lower control arm. k wheel ? (MR) 2 ? k spring We need to set the motion ratio according to the wheel travel we require for our suspension. A travel of 50 mm was required and a list of springs were collected and measured for their stiffness characteristics. According to this calculation the motion ratio for auto spring As (Ks=58. 57N/mm) wheel rate (Kw=41N/mm) the motion ratio was 0. 8366. Travel of spring per unit wheel travel)The travel obtained by this spring was lesser than was required we could only obtain 26mm of travel in Page 5 STEERING DESIGN Objective of steering system in Baja vehicle ? ? ? To provide un problematic maneuverability of the vehicle over the undulating terrain. It must be durable to sustain the harsh offroad racing course. Less bump steer and return ability in steering Customer requirement DESIGN OF WHEEL HUBS Our wheel hubs hav e been designed and fabricated after an extensive research. Effort has been made for minimum scrub roentgen and obtains the best possible wheel geometry.Adams and Ansys have been used to take over and analyse the behavior of these hubs respectively. We have two major design concepts 1. 2. 3. 4. Optimum sensitivity Low turning radius Minimum feedback Low exist and easy maintenance Basis of our design We have decided to opt for a 400 degree lock to lock rack and pinion steering with Ackerman geometry. volute cut teeth will be used for the rack and pinion due to the following advantages over spur gears ? ? ? ? They take higher loads. They are quieter and smoother. HUB 1 SCRUB RADIUS FACTOR OF SAFETY HUB 2 8 mm 4. 6 1460gm. 15 mm 5. 2 2506gmRulebook Constraints All vehicles must be equipt with positive wheel lock? to? lock stops and adjustable tie rod ends must be constrained with a jam nut to prevent loosening Tie rod of vehicle should be secured by bumper in front or any other sa fety device in rear in order to avoid damage of tie rod during collision. WEIGHT Hence taking various factors in to consideration HUB 2 is considerd for fabrication and stress analysis is done on it. ALTERNATIVES CONSIDERED STRESS DEFORMATION Rack and Pinion Good High Low Light 1. Extermely Simple 2. Gives good driving step Recirculating ball screw Very High Low High Very High 1.Very Low free play 2. Non-selfreturn ability Worm and sector High Low Very High Comparatively Heavy 1. High free play 2. Non-selfreturn ability FRONT HUB Efficiency Compactness Cost Weight Comments REAR HUB Calculations Distance between King Pins (c) Using the formulatione = 1117. 6mm FORMULAS FOR STEERING ANGLES ? ? ? cot O cot ? =c/b sin ? =(c-d)/2r sin(? + ? ) +sin(? O) =2sin ? Page 6 ? ? ? ? ? ? ? ? ? ? ? ? ? ? ? BOBLLIER CONSTRUCTION FOR bike POSITIONING turning radius = (track/2) + (wheelbase/sin(average steer angle) here O=? o =outer wheel angle ? =? I = inner wheel angle wherefore ? steering a rm angle r = length of the steering arm c= kingpin to kingpin distance d=length of the track rod b=wheelbase CALCULATIONS wheel base (b)=1532mm kingpin to kingpin distance(c) = 1117. 6 steering arm angle ? =30 degrees on substitution an comparing two results we get ? =40 degrees O=27 degres turning radius was calculated to be 2. 9m Clevis joint is used in rack to reduce the bump steer . The below picture shows the clevis joint used Rack and Pinion design Rack displacement calculation RACK ANALYSIS FOS 8. 5 deformation stress From to a higher place formula we get rack displacement =40+40=80mmThe picture of the complete rack assembly Page 7 Values No. 1. 2. 3. 4. 5. 6. 7. incident Symbol Formula Spur Gear 2 20 11 zm/2 + H zm D cos? 35 22 Db 20. 67 23 24 Rack Module Pressure angle Number of teeth Height of Pitch Line Centre Distance Pitch diameter Base Diameter M ? Z H Ax D Adams results CALCULATION OF FORCES ON RACK AND PINION R=steering wheel radius = 165mm r=pinion pitch-circle radius t=number of pinion teeth = 6 p=linear or circular pitch =22mm E= commentary steering-wheel effort = 2 * 20N W=output rack load If the pinion makes one revolution input steering wheel movement Xi = 2? Output rack movement Xo = 2? R = txp = 82. 86mm Therefore Movement ratio (MR) = Xi/Xo=2? R/2? r=2? R/tp=R/r= 165/11=15 15= W/E, w=600N force is to be applied on to the pinion to move the rack. ? ? ? ? ? ? ? ? Ft = Transmitted force Fn = Normal force. Fr = Resultant force ? = pressure angle Fn = Ft tan ? Fr = Ft/Cos ? Here ? =20 degrees therefore Fn=194. 95NFr=630N Opposite wheel travel Fig 3 Graph 1 camber angle vs wheel travel Graph 2 roll centre height vs wheel travel Graph 3 wheel rate vs wheel travel Fig1Shows the single wheel travel vs toe change and scrub radiusPage 8 POWERTRAIN DESIGN ? ? Engine and contagion are the loudest systems of the vehicle. Since the engine provided could not be touched in any way, the only hindrance lessening technique that could be adopted was through the use of proper mufflers. Various mufflers were tested on the engine but the stock muffler provided the least noise levels . It also provided the best fuel efficiency . So it was decided to use the stock mufflers considering the Go Green theme. The gear box and CV joints should always be kept properly lubricated to minimize noise due to friction.To reduce vibrations transferred to the chassis from the engine, it is mounted on rubber bushes. The drive shafts are welded properly so that they are inline and no vibrations occur during rotation. The gearbox is mounted firmly in such a way that there is a minimum contact between gearbox and chassis which means minimum transfer of vibration to chassis. The fuel tank capacity is 4 litres. ? ? ? Fig2 Shows roll steer vs wheel travel ? ? ? ? Driveline Power is familial from the engine to the wheels in the following way Engine Stub Axle Chain Drive Wheels Gearbox DriveshaftOpposite wheel travel fig 4 Graph 1roll centre vs roll angl e Graph 2 camber vs roll angle Graph 3 roll stiffness vs roll angle The Driveshaft consists of dowel pin on the gearbox side and rzeppa joint on wheel side . This design ensures transmission of power with minimal losses and allows transmission at longer wheel travel Page 9 Design Methodologies A customer expects the following things from the transmission system of a Baja vehicle Forward Orientation Gear Final Gear Ratio 31. 48 18. 70 11. 40 7. 35 55. 08 ? ? ? ? Max. Vehicle Speed (Km/hr) 12. 04 20. 27 33. 26 51. 59 6. 88 Max. Tractive Effort (N) 2240. 7 1348. 28 821. 93 485 3971 abolish Orientation Final Gear Ratio 55. 08 32. 72 19. 95 13. 40 31. 48 Max. Vehicle Speed (Km/hr) 8. 17 13. 68 22. 19 32. 14 12. 75 Max. Tractive Effort (N) 2990 1776. 23 Reverse engine orientation resulted in problem with weight distribution and increased vehicle length. Using the transmission in forward helped to shift the center of gravity towards vehicles center. Due to decreased reduction it also re sults in increased vehicle speed. It also provides faster acceleration and higher top speed due to this reason we decided to use the transmission in forward orientation.To calculate vehicle speed at different engine speeds in different gears, we used the formula V= (2*3. 14*engine speed*radius of wheel/Gear ratio)*(60/1000) km/hr. The gear ratios obtained are Chain Drive gear ratio = 28/28 =1 1083 818. 36 1708. 91 The following graph is obtained Tractive effort is calculated by formula F=Engine torque*Gear efficiency/wheel radius The curves obtained are ratio*transmission First Gear Second Gear Third Gear Fourth Gear Reverse Gear High speed for acceleration and speed trials. High torque for towing and hill climbing events.It should be reliable and light weight. It should transmit power in any driving conditions. ? The gearbox operation should be smooth and easy for driving comfort. The engine used has low power to weight ratio, so its necessary to transmit power with minimal loss th rough drive train. It should be such that it can be easily couple with the engine. Alternatives considered We had three options while deciding the transmission system a) b) c) A cvt mated with Mahindra gearbox. A custom made manual gearbox. Use of Mahindra assistant gearbox coupled with chain drive. 3000 2000 1000 0 0 2000 4000 ractive effort in beginning(a) gear tractive effort in 2nd gear The maximum Tractive effort obtained is 2240N at 2600rpm in 1st gear. Providing an acceleration of 5. 6 m/s2. The variation of full throttle power with road speed is shown below with different gear ratio Our previous experience with cvt had problem of belt slipping at high torque conditions. Also it resulted in increased weight. So we decided against using this. As we already had 2 champion Alfa gearboxes, we decided on using this gearbox alongwith a chain drive due to the following reasons 1) 2) 3) 4) Reduced chassis width.Can be easily coupled with the engine. commensurate drive shaft length s increased ground clearance. Minimum rear overhang better vehicle dynamics. 60 2nd gear 40 1st gear 20 0 0 2000 4000 3rd gear We had 2 options for the orientation of gearbox A) Forward engine with engine in the front rear axle. B) Reverse engine orientation with engine behind of the rear axle. Total resistance of the vehicle at 3600rpm is found out by the formula R=k AW2+KW+WsinO. Where k= coefficient of air resistance N-m2. Page 10 A=frontal area of the car, m2. V= vehicle speed, km/hr. K=constant of rolling resistance.W= weight of car,N O=gradient angle, degrees. The value of resistance comes out to be R=442. 64+2452 sinO. We put this value in formula RV/3600nt=power of engine By solving the above equation for o, we get o=33 degree at 2600 rpm in 1st gear. Stopping Distance Braking Efficiency Parameters Master Cylinder Diameter Caliper Diameter Brake pad height Diameter of the disc Co-efficient of friction of the pasture brake pad Force generated by both the brake pads per whee l Braking Torque per wheel Weight of vehicle(with the driver) Wheelbase Height of COG self-propelling front axle load Dynamic rear axle load 0. 11 m 56% Magnitude/value 19. 05 mm 32 mm 27 mm 162 mm 0. 38 3431 N 1040 N 360 Kg 1397 mm 601. 3 mm 1780 N 1650 N 70 60 50 40 30 20 10 0 0 2000 4000 gradabilit y in 1st gear Gradabilit y in 2nd gear Gradabilit y in 3rd gear BRAKING DISTANCE VS revive This shows that the vehicle is capable of climbing a 30 degree slope in 1st gear. This is more than enough for heavy off-road conditions. BRAKES The criterion of designing the brake system, as stated by the rule book is that, all the wheels must lock simultaneously as the driver presses the brake pedal.Our ATV consists of disc in all the four wheels, as disc brakes are safer, reliable and more effective than drum brakes. Brake circuit used is Independent in order to ensure safety We are using rotors of the same diameter for all the four wheels. Special ATV rotors and wheel calipers have been im ported from Taiwan and tandem bicycle Master Cylinder of Maruti 800 is being used. Brass linings and Rubber (flexible) brake hoses are being used in the circuit. A Pro-E model of the brake circuit in the vehicle Brake specifications Force of the driver on the pedal Average circuit pressure Pedal ratio Deceleration 400 N 5. 16 N/sqmm 41 5. 5m/sqsec Page 11 BODY PANELS The criteria for selecting the material for body panels firewall and belly pan was as follows ? ? ? ? ? ? guard of the driver Rulebook constraints Weight of the panels Recyclability of the material used Cost of the material Serviceability of the vehicle INNOVATION Solenoid Operated Fire Extinguisher The body panels are divided into three parts Side panels, front bumper and rear panels. For increasing the serviceability of the vehicle, the panels and front bumper have been mounted using easily detachable clips.The materials used for the firewall and belly pan are 1. 5mm thick aluminium alloy sheets, which are both lig htweight and 100% recyclable. For body panels, 0. 2mm thick sheet metal is used. It is also 100% recyclable. We have decided to incorporate following safety features in our vehicle 1. All disc brakes with cross circuit. 2. Corrosion resistant stainless steel bolts with nylon lock nuts for all fastenings. 3. 2 fire extinguishers 4. First aid kit 5. Spill guard and splash shield for fuel tank 6. Four point harness seat belts. 7. enormous open throttle stop at the pedal. . Reverse alarm and brake lights. 9. Two 01-171 Ski-Doo decimate switches. 10. Steering stop at the wheels. 11. Rear view mirrors. 12. fervour switch for engine, apart from pull start. 13. Electronic operated fire extinguisher. 14. rump belt engine kill system 15. Driver emergency communication system This refreshed kind of fire extinguisher agreement is operated electronically through a solenoid valve. In case of fire the valve is opened by a manually operated sacking and a jet of CO2 is released in the engine compartment through various angles.This effectively extinguishes fire in the engine compartment and stops its further propagation. Seat Belt Engine Kill System This system is designed such that the driver will not be able to start the car until he engages his seat belt. The seat belt acts as a switch to operate the relay connected to the engine kill wire. When the seat belt is disconnected, the engine kill wire is grounded. Thus, the car cannot be started. As the seat belt is engaged, relay operates, and the engine kill wire circuit is now open enabling the driver to start the COMMUNICATION SYSTEMPURPOSE This is a two way communication system wherein messages and signals can be transmitted from the pit to the driver and vice versa. FEATURES The system uses two microcontroller based Arduino boards fitted with an ZIGbee communication module. Page 12 It is a transceiver. The signals are sent and received with the help of color coded Push Buttons and LEDs. The actual tested system a rrangement is shown in figure. BILL OF MATERIALS All the parts of the ATV are classified into eleven blocks and are given a unique ten digit part number.The cost of procurement of the part or the material is mentioned and all the machining operations are stated clearly. The spread sheet calculates the cost of machining also. Finally, the sub total of the procurement cost and the machining cost is obtained which helps in grand total of the costs. The BOM gives the level of hierarchy to each part. Sub-Division Engine Transmission Brakes Steering Suspension Wheels Electricals Body Chassis Fasteners Safety Grand Total Cost(in INR) 17000. 00 16800. 00 6928. 00 4457. 00 29954. 00 40308. 00 7940. 00 5340. 00 16240. 00 1346. 00 8272. 00 154585. 00 Page 13ACKNOWLEDGEMENT AND REFERENCES ? ? BAJA SAEINDIAINDIA Rulebook. ASIA 2010 Gillespie, Thomas D. , Fundamental of vehicle dynamics, SAEINDIAINDIA publication ? ? ? ? ? ? ? ? ? ? Body-space Anthropometry, Ergonomics and Design by Stephen Phe asant. Automotive Engineering Fundamentals by Richard scar and Jeffery K. Ball The Multi body Systems Approach to Vehicle Dynamics by Mike Blundell and Demian Harty Theory of Machines by S S Ratan Automobile Mechanics by N. K. Giri Machine Design by R. S. Khurmi Strength of Material by R. K. Rajput Google. com Howstuffworks. com Wikipedia. org Page 14