Car # 102

Sri Ramakrishna Engineering College Baja SAE India 2013Final Design Report

T.AAKARSH DEEPTeam captain

M.B. LOOKMAN UL HAKEEMSuspension head

Copyright © 2009 SAE International

ABSTRACT

Baja SAE India, is an inter collegiate designcompetition for aspiring engineers to create aminiature off-road vehicle, through which theobjective of simulating real world engineeringdesign challenges are met. The students mustdesign, build and market a single seater all terrainvehicle with adherence to the Baja SAE Indiarulebook. The vehicle should aspire to marketleading performance in terms of speed, ride,handling and ruggedness over rough terrain and offroad conditions. An aspect of this competition is tocompose a design documentation package thatcreates an overview of the vehicle’s constructionelements. The Phoenix Racing team has createdthis report to describe their design.

INTRODUCTION

In an effort to establish Phoenix Racing baja teamamongst the elite who compete in the annual SAE(Society Of Automotive Engineers) collegiate bajadesign series competitions, a thorough inspectionof our previous Baja was conducted and comparedto other highly competitive Baja teams in priorcompetitions. Considering the top five teams in alldynamic events over the past several years, arough outline of all strong design characteristicswere compared to Phoenix Racing’s lastcompetitive design. This process immediatelyexposed our current weaknesses in need ofimprovement

To accomplish goals involving increasedoverall vehicle performance, half of our team hasfocused on designing a new and revamped drivetrain, while the other half focused on improvements

to both the current frame and suspension. Both ofour teams went through three major phases duringthe overall design process. The first phase of theprocess was the research phase in which valuablebackground information was learnt, troublesomeissues were determined, and brainstormed newdesign ideas to overcome these issues. Next, thedesign phase began which included the modelingof our frame, suspension, and gearbox incomputer-aided programs such as lotus suspensionanalysis (shark), KISSsoft, Pro Engineer andsolidworks in order to establish and perfect theoverall design packaging. Lastly, an analysis phasetook place to determine the structural integrity andaerodynamic efficiency in our design. The designphase encompassed finite element analysis toolslike ANSYS and solidworks software andComputational Fluid Dynamic tools like ANSA andANSYS fluent to ensure that our designs werestrong and efficient to endure all SAE Baja events.

TECHNICAL SPECIFICATIONSSPECIFICATIONS

Target Achieved

DimensionsLXWXH (mm)

2062X1574X1549 2286X1574X1560

Front TrackWidth

1372 mm 1372 mm

Rear trackWidth

1270 mm 1270 mm

Wheel base 1626 mm 1626 mm

CG height 423 mm fromground

381 mm

Table1:Vehicle Specifications

3-D MODEL:

Fig 1: ISOMETRIC VIEW

Fig 2:3-D MODEL

ROLL CAGE DESIGN

DESIGN CONSIDERATIONS

The primary objective of the frame is to provide a 3-dimensional protected space around the driver thatwill keep the driver safe. Its secondary objectivesare to provide reliable mounting locations forcomponents, be appealing, low in cost, and low inweight. We met these objectives by choosing aframe material that exceeds the SAE strengthrequirements, but still gives us an advantage inweight reduction. We provided a low cost framethrough material selection and incorporating morecontinuous members with bends rather than acollection of members welded together to reducemanufacturing costs.

Engine Drive

Max.Acceleration

2.5m/s2 2.19 m/s2

Max. Speed 55 Kmph 52.7 Kmph

Max.Gradeability

35o 33.8 o

Max.Traction 2548N 2487 N

30m dash time 5.4s 5.3s

Suspension and Steering

Groundclearance

Front and Rear:330 mm

Front:330mmRear:279 mm

Static stabilityfactor

1.44 1.51

Roll Gradient 1.5deg/g 1.6deg/gFront stiffness 26.10N/mm 27 N/mm

Rear stiffness 37.320N/mm 39 N/mmSteering Lock tolock

1 turn(360 deg.) 1.2 turns(432deg.)

Turning Radius 3.05m 3.187m

Brakes

Brakes type Hydraulic Rounddisc(front:2/Rear:1)

HydraulicRound disc(front:2/Rear:1)

Disc diameter Front:190mm/Rear : 240mm

Front:180mm/Rear : 240mm

Brake bias 50:50 50:50

Stoppingdistance

13.4 m (from60kmph)

15 m (from60kmph)

Decelaration 8m/s2 7.2 m/s2

Weight

Kerb Weight 235 kg 252 kg

Sprung mass 175kg 185kg

Unsprung mass 65kg 67kg

DESIGN METHODOLOGY:Rule compliance Preliminary CAD design PVCmock-upProvisions setup CADModellingFEA model refinement final model

MATERIAL SELECTION

The materials used in the cage must meet certainrequirements of geometry and minimum strengthrequirements found in SAE competition rules. Sincethe frame is being used in a racing vehicle ratherthan a recreation vehicle, weight is a very largefactor in the shape and size of the frame. Theproper balance of strength and weight is crucial forthe team’s overall success. We felt that one of thekey design decisions of our frame that wouldgreatly increase safety, reliability and performanceis material selection. To ensure that we chose theoptimal material, we did extensive research andcompared materials in multiple categories. Our keycategories for comparison were strength,availability, machinability, weight, and cost. Designconsiderations aside, the driving factor behindchassis material selection were the SAEcompetition vehicle regulations. The ruling left theteam with only few options. The comparison chartof various materials that we analyzed under therules specified by BAJA SAE are as follows:

MATERIAL CHROMOLY4130

ASTMA252

AISI1018

IS3074

ASTMA106B

Carboncontent

( % )

0.32 0.18 0.18 0.20 0.30

Yieldstrength(MPa)

395 350 365 372 383

Tensilestrength(MPa)

560 455 450 473 466

Elongation(%)

25 20 20 5 20

Cost inrupees /m

550 650 600 800 500

Availability Easy Medium

Easy Difficult

Easy

Table 2: Comparison Of Materials

FRAME DESIGN AND TESTING:Solid Modelling and Analysis:The vehicle was completely modeledand analyzedusing Solid Works. A complete vehicle assemblywas created in Solid Works. This completeassembly included the chassis, suspension,steering linkages, brakes, wheels and drive train.Frame Dimensions:

As we began to visualize and dimension our frame,we kept in mind strength, aesthetics, and lowmanufacturing cost. We adjusted the dimensionsand placement of key members to make sure thatthey would fit well with other components. We alsodesigned our frame members to incorporate manybends so that we would decrease the amount ofwelding that would need to be done. This helps tokeep the strength and integrity of our framemembers as well as decrease the manufacturingtime and costs. Section 31 in the SAE rules hasstrict guidelines that must be followed for thedimensions of the roll cage. These guidelines alongwith the fact that our team members range inheights of 5’ 8” to 6’ 5” guaranteed that wedesigned a vehicle that will fit almost any size adult.To accommodate our drivers we made a cockpitthat is 29” wide. This allows the tallest drivers tostick their legs to the side of the enclosed volumewhen driving and still be safely encompassed bythe roll cage. At the same time the short drivers areable to keep their legs stretched straight and stillcomfortably reach the pedals.When the entire powertrain was modeled, theengine bay area was found to be sufficient. Thepowertrain and other features were analyzed andthe engine bay was resized. Serviceability was alsotaken into consideration with changes made to rollcage design.

WEIGHT:The performance of the vehicle depends upon theweight of the vehicle. As the weight reduces theperformance or the efficiency gets increased.

Length of Material used In Roll Cage = 32 mDensity of Steel= 7833.4 Kg/m*3

Primary Member Outside Diameter=33.4 mm Length = 12 m Wall thickness = 3.38 mm Inner Diameter = 33.4 - (2 x 3.38) =26.64

mm Area = /4[D*2 – d*2] Area = 318.47 mm*2 Volume = Area x Length=318.47 x 12000 Volume = 3821640 mm*3 Volume = 0.003822 m*3 Total Weight = Density x Volume = 30.05Kg

Secondary Member Outside Diameter = 31.75 mm

Length = 20 m Wall thickness = 2.10 mm Inner Diameter = 31.75 - (2 x 2.10) = 27.55 Area =195.5 mm*2 Volume = Area x Length= 3910000 mm*3 Volume = 0.00391 m*3 Total Weight = Density x Volume =30.62 Kg

Total Weight of Roll Cage = 60.67 Kg

DRIVER ERGONOMICS:Ergonomics is the science of equipment designintended to maximize productivity byreducing driverfatigue and discomfort. One of the major designcriteria that was used in the design of the chassis isthe idea of driver ergonomics. It is an essential partof the car that each member of the team is able tosafely and comfortably operate the vehicle.Although driver comfort sometimes is oversteppedor thought of last, if we are to meet our goal ofperforming well and completing the endurance racein good standing, than ergonomics must beconsidered. Good ergonomics will allow the driverto drive quickly and comfortably avoidingunnecessary stoppage throughout the endurancerace.

To allow for proper movement of the drivers feetto control the gas and brake pedal, the foot boxwas designed to be as small as possible whileallowing the driver the proper amount of space tooperate the controls safely. The size and shape ofthe foot box also allowed for an optimum placementof the brake pedal and assembly low and betweenthe lower frame members maintaining a low Centreof gravity.

The final ergonomic parameter that wasconsidered was the location and inclination of thedriver’s seat. The seat is crucial to supplyingenough support to the drivers back to allow him tostay upright with a clear view of the track ahead, toapply the proper forces to the gas and brakepedals, and support the shift in the driver’s weightwhile cornering or landing from a jump. To properlydetermine the inclination of the seat a great deal ofresearch was done both online and by a trial anderror approach with many members of the teambeing tested. In the search that was conductedelectronically, the typical answer was anywherefrom 10-20 degrees was the standard used in theautomobile industry with a great deal of emphasison adjustability being a very important feature fordriver comfort. With this basic information, aphysical test was conducted with a few members of

the team and it was found that roughly 20 degreesof inclination was comfortable for those tested andwas chosen.

Fig 3: Ergonomic check

Before we started any fabrication, we decided tobuild a life size mock-up frame out of PVC pipe.This gave us the opportunity to measure the framedimensions with our tallest and shortest driverssitting in the frame. It enabled us to check the driverergonomics and other design changes that can beimplemented.

STATIC AND DYNAMIC ANALYSIS: Refer theillustrations (figure 22)

SUSPENSION SYSTEM

The primary system that determinesperformance and car handling , while traveling offroad, is the suspension system. The purpose of thesuspension is to reduce shock loads that act on thecar while providing optimal wheel contact whenoperating under dynamic conditions. Thesuspension must provide enough wheel travel todampen the impacts imposed on the vehicle.

DESIGN CONSIDERATIONS

Three primary goals were established whiledesigning the suspension system. The first goalwas to design a suspension system with minimumunsprung mass, as the vehicle performanceincreases with a decrease in the unsprung mass.The second goal was to control and obtain desiredwheel geometry such as to obtain negative cambergain in bump and cornering, and to eliminate the

bump steer. To promote oversteer while eliminatingthe usage of anti-roll bar.

DESIGN METHODOLOGY: Vehicle specific leveltargets suspension type selection stiffness &shock selection suspension geometry in lotusFEA in ANSYS knuckle hard point generation knuckle design & analysis

TARGET DESIGN SPECIFICATIONS Ground Clearance: Front-13 ’’: Rear- 11” To achieve FLAT RIDE Camber Gain in Bump To promote Over steer Zero Bump Steer Lowest combined mass possible

Suspension Type Selection

Front Suspension

Double-wishbone suspension type waschosen for the front as it facilitates bettersuspension tuneability and provides better cambergain in jounce than its counterpart MacphersonStrut .It also holds good for resisting bending forceand distortions. The material was chosen to beSAE 1026. In the 2011-2012 season ,Doublewishbone suspension type was used at both frontand rear. But the performance of the vehicle waslargely affected by the suspension failure in termsof low spring stiffness and limited travel . This yearof air shocks are used that not only acts as aprogressive spring but also provides good travel.

Fig 4: Lotus Suspension Analyzer

Fig 5: FEA of lower wishbone

Rear suspension

Good serviceability and easy packagingplayed a major role in choosing Trailing arm withcamber links as the rear suspension system for thisyear’s buggy, owing to the problem of limitedaccessabilty and serviceability of the engine baycaused by using double wishbone suspensionsystem at the rear,in the previous year’s car. Thetrailing arm suspension system is a simple andreliable design that is not only durable but alsoprovides good vertical wheel travel than most othersystem. Apart from the above characteristics theshocks can be mounted anywhere on the armwhich allows complete utilization of the shocks

Fig 6: Trailing arm with camber links

Frontsuspension

Rear Suspension

TypeDouble SLA

wishbonesuspension

Trailing arm withcamber links

ShockAbsorber Fox Float R EVOL Fox Float R EVOL

Shock travel 6 inch 6 inch

Shocksstiffness 26.10 N/mm 37.320 N/mm

Motion Ratio 1.83 1.83

Material SAE 1026 SAE 1026

Roll centreheight 234mm 274mm

Lateral loadtransfer

distribution35% 65%

Roll ratedistribution 37% 63%

Camber -1.5degree -1degree

Castor 5 degree -

Bump Steer Nil Nil

Table 3: Suspension specifications

SUSPENSION STIFFNESS CALCULATION

Ride frequency CalculationA ride frequency is the undamped natural

frequency of the body in ride. Anything between 60to 90 cpm (1-1.5 Hz) is perceived to be desirablefor comfort ride. Lower frequencies produce asofter suspension with more mechanical grip,however the response will be slower in transient.

Flat ride:The flat ride is achieved when the

rear suspension frequency is greater than that offront suspension. This is because ,the higher rearfrequency catches up with the front suspensionwhen the car pitches and reduces the pitchingaction of the vehicle,thereby provides comfort.

Fig 7:Flat Ride Condition

Flat ride condition

Frequency is given by

Assuming the undamped frequency as 1 Hz..

1Hz * 4 *∏2 *330 = K

Spring stiffness , K = 13027 N/m

Spring rate for front and rear suspension:

Front Wheel travel= 11 inch; Spring travel = 5.8 inch msm = 190 kg fr = 1 Hz MR = 11/5.8 =1.89 Ks = 4*∏2 *12 * 190 * 1.892

front spring stiffness , Ks = 26193 N/m

Rear spring stiffness , Kr = 4 *∏2 fr2 * msm * MR2

Kr = 4 * ^2 * 1.352 * 190 * (11/6)2

Rear spring stiffness , Kr = 37973.30 N/m

Front and rear roll rate:

Wheel rate at frontWheel rate, KLF = KRF = spring rate / (motion ratio)2

KLF = KRF = 26193/1.89 = 7332.66 N/mTF = 54 ” = 1.3716 m

KΦF = ∏ * (1.3716)2 * 7332.66 *7332.66

180 * (2 * 7332.66)

KΦF = 120.3 (Nm/deg roll)

Rear roll rate:

KLR = KRR = 37973.30/1.832 = 14701 N/m

KΦR = ∏ * (1.27)2 * 14701 *14701

180 * (2 * 14701)

KΦR = 206.91 Nm/ deg roll

Roll gradient :

W = 330*9.81 H = 6.5* .0254 m KΦF = 120.3 Nm/deg roll KΦR = 206.94 Nm / deg roll

ΦR = -330 * 9.81 * 6.5 * .0254AY

120.3 + 206.94

ΦR = -1.6 deg / rollAY

The Suspension Characteristics

Suspension characteristics such as camber,toe in, toe out, scrub radius plays a pre-dominantrole in determining the performance and handlingof the vehicle. For any off road vehicle, maximumtire contact on the road is very vital whileencountering corners or bumps for providingmechanical grip and negative camber gain of thewheel ensures maximum tire contact patch andhence better mechanical grip. Bump Steer is oneof the undesirable characteristic that has to betaken care of while designing a suspension system.Zero bump steer was achieved by working inaccordance with the steering team.

Fig 8:Bump Vs Suspension Characteristics

-6

-4

-2

0

2

4

6

-100 0 100 200

Bump Travel

Bump Vs SuspensionCharacteristics

Front camberchange

Rear camberchange

Bumpsteer(toechange)

castor angle

Shocks: FOX Float R :

As previously stated the performance of thelast year car was greatly hindered by lowsuspension stiffness and very limited travel. So Airshox was selected for this year’s buggy as it hasseveral prominent advantages over coil-oversprings. Air Shox not only reduces the unsprungmass of the vehicle but also increases the ridehandling characteristics of the vehicle. It acts as aprogressive spring whose stiffness increases whenthe load acting on the vehicle increases. The foxfloat R EVOL was selected since the force vsvelocity curve of our suspension design matchedwith that of Float R EVOL.

Fig 9: Force Vs Velocity Curve

STEERING SYSTEM

Design Considerations

The main objectives of the steering systemis to handle stress for safely directing the vehiclethrough any type of terrain and to provide the driverwith an accurate, predictable, and reliable methodof driving along with good ergonomics.

TARGETED DESIGN SPECIFICATIONS 100% Ackermann geometry for directional

control and easy cornering of the wheels. Minimum turning radius to increase the

chance of winning the maneuverability andendurance race.

Minimum steering ratio to decrease driver’seffort for maneuvering the vehicle.

DESIGN SPECIFICATIONS ACHIEVED:

In order to meet the targeted designspecifications 14” centralized rack and pinion waschosen from desert karts. This rack and pinionsetup has a rack travel of 4.25” (107.95mm) andsteering wheel turn equal to 1.5 turns.

This setup is used to achieve theAckermann geometry using iterative methodsresulting in achievement of 85% Ackermanngeometry which was verified using LOTUSsoftware.

Fig 10 :Ackermann Geometry

Fig 11 :Ackermann Percentage

Outer Steering angle (δ o) = 24.261 deg =0.4234 rad

Inner Steering angle (δ i) = 34.177 deg =0.5965 rad

δ = (δ o + δ i) / 2 = (24.261 + 34.177) / 2 =29.219 deg = 0.5099 rad

δ = 29.219 deg

0200400600800

10001200

0 20 40 60

Force

(

lbs)

Velocity (in/s)

Force Vs velocity Curve

Float R

Obtained ForceVS velocity curve

δ = L / R Radius of turn (R) = L / δ = 1625.6 / 0.5099

= 3187.668 mm

R = 3.187 m.

Steering Ratio = degrees of steering wheelangle : Front wheel angle = 432 : 58.438

Steering Ratio = 7.39 : 1

Type Centralized Rack andPinion

Steering RatioTurningRadius 7.4 : 1 (20:1)

3.187 meters(3.82)

Rack Travel Lock toLock

3.38 inch/4.25inch 1.2 turns (3)

Over Steer Gradient (K) - 1.74 deg / g

Steering Torque 71.57 N-m

Table 4: Steering Specifications

Drive train:

Due to the restrictions regarding the engine,optimal performance of the vehicle in terms ofacceleration and top speed must be achieved byreducing the vehicle mass, reducing frictional

forces, and optimizing the drive train. The drivetrain implemented on the 2011 – 12 season wasfour speed Mahindra alfa transaxle, which washighly durable but the overall weight and the highinertial mass of the rotating components reducedthe acceleration performance plus the open slipdifferential in the transaxle failed to propel thevehicle through muddy terrains. The shifting alsoconsumed valuable time and driver effort

The first step in the design process is todetermine the type of reduction needed to completethe power transfer. The most important designconstraints that drove the team’s decisions are: thedrive train must be light in weight and compact dueto the small size of the vehicle, it must transfer thecalculated amount of power to the drive shaftallowing it to complete both high and low gearapplications that will be necessary to complete allaspects of the endurance portion of competitionwithout failure, should take very less time and effortfor shifting lastly complete the previous tasksaround a single supplied engine and a tire size of25 inches.

In order to meet the above goals we had todetermine if we were going to use a manual orautomatic transmission. For an automatictransmission we would have to utilize aContinuously Variable Transmission (CVT). We didextensive research to compare the manualtransmission (MT) and the CVT.

Table 5 : CVT vs MT decision matrix

Most MT that would fit our application arefound on motorcycles and four wheelers. Thesevehicles have a very high rpm range. A MT onthese vehicles is beneficial because the operatorcan shift into a higher gear with the rpm at a highvalue. Since our engine has such a small rangebetween 1750 and 3800 rpm, the performance gainby incorporating a MT is minimal. We felt that sincethe CVT allows our engine to constantly run near itsmaximum torque, it would give us the ability to getmax power from the engine in both the high andlow ranges. Also, operating the CVT is easier for

Parameter CVT MTWeight 1 0Performance 1 1Drivability 0 1Reliability 0 1Tuning 1 0Simplicity 1 0Ease of Installation 1 0Total 5 3

Fig 12:Rendered view of the drive train subassembly

the driver since the driver doesn’t have toconstantly shift gears. The performance gain of theMT happens only if the operator shifts gears at theoptimal RPM, but if the operator does not thenthere is a significant loss of performance. By usingthe CVT we eliminate this possibility of error whichgreatly improves the performance and reliability ofthe vehicle especially when it comes to enduranceevents. Also, this ease of operation will attractconsumers to buy our product.CVT: CVTech PWD 50/LP2

Belt length (c/c) = 299 mmMax. Ratio = 0.43:1Min. Ratio = 3.00:1Engagement speed = 2300 rpmShift speed = 3200 rpm

Fig 13: Engine speed vs vehicle speed graph

To connect the CVT to the axles we had theoption of using a chain driven system or using agear box. We wanted to keep the efficiency andreliability of our power-train system high. Gears arethe most efficient way of transmitting power. Havinga chain in the system provides another area forfailure with the possibility of the chain beingknocked off of the sprockets. It also increases themaintenance of the system by having to constantlymaintain the tightness of the chain as it stretches

Table 6: Achieved performance using the current drive trainsetup

and having to replace the chain after extensive use.The decision was made to make a custom FNRgear box for secondary reduction as the existingOEM gearboxes were either heavy or doesnot havethe suitable reduction ratios. The gears, shafts andbearings were designed with the help of KISSsoftmachine element design software as per AGMA2101-D04 (Metric Edition) regulations.The gearboxwas modeled using Pro ENGINEER CAD software.

Fig 14: CAD model of the custom FNR gearbox

FNR GEARBOX: CUSTOM MAKE

Forward gear ratio = 13.8:1Reverse gear ratio = 20.5:1Type : constant mesh, two-stage reductionSolid spool driveDetent and ball shifterLubrication : oil bath typeLubrication oil : Castrol opti-gear synthetic A 220Lubricant Oil Viscosity : 220 Centistokes (at 40̊̊̊ C )Ventilation, drain plug, dip stick provided in casingCasing material : Al 6061Shaft Material : EN 24Gear Calculation Method : AGMA 2101-D04 (MetricEdition) [KISS SOFT SOFTWARE]Max. Shaft Deflection : 8 µmGearbox Dimensions : 389x92x193 (l x b x h)

2012 modelPerformance

2013 targetperformance

Achievedperformance

max speed 45 Km/h 50 Km/h 52.7 Km/h

max.acceleration

1.37 m/s2 2.5 m/s2 2.19 m/s2

max.gradeability

30° 35° 33.812°

0 – 30m 9 sec 5sec 5.4 sec

Fig 16: deflection analysis of intermediate shaftusing KISSsoft sftware

Power unit mountings:The engine was mountedover four rubber engine beds of jeep to dampen thevibrations transmitted to the roll cage. The gear boxwas mounted through delrin bushes for the samepurpose mentioned above.( see illustration no: )

CV joint plunge and articulation calculation:Extensive discussion was made with thesuspension team to suit the position and length ofthe camber links in the rear suspension, to reduceCV plunge and articulation required duringsuspension travel. The camber links were broughtcloser and parallel to each other such that the CVlied in the vertical midpoint of the camber linksArticulation needed: 16 degree during full droopand 5 degree during full bumpPlunge needed; 4 mm for the whole suspensiontravel(See illustration no: )

Weight:Engine: 24 kgsCVT: 5.5 kgsGear box: 15.2 kgsDrive shaft: 11 kgs (5.5 each)Hub: 1.9 kgsWheel: 18.8 kgs (two wheels in fully inflatedcondition)Total: 76.4 kgs

Braking System:The purpose of the braking system is to

increase the maneuverability by locking all the fourwheels at a time. It is required to statically anddynamically lock all four tires on both hard andloose surfaces. Although the previous designworked effectively it was decided that modification

Table 7: Gear Data

of our previous year’s braking system is necessaryin order to reduce the unsprung mass of the vehicleand to achieve 50-50 biasing. This year our aim isto reduce the weight of wheel assembly so afterstudying various available options to get effectivebraking we have decided to use hydraulic discbrakes on all the four wheels with a single in boarddisc at the rear. Honda Aviator’s disc and caliperhas been chosen for the front and pulsar disc andcaliper has been chosen for the rear. Hydraulic discbrakes have high performance, easy replacement,and comparatively less weight than other availableoptions. have the common circuit of brake fluid forall the four wheels. So if any one chamber fails wecan still apply the brakes to all the four wheels .Toreduce the unsprung mass, it was decided inboardbrakes at the rear would be used.

Fig 16: Front/rear Hydraulic Split

SPECIFICATIONS

PINION GEAR

NO. OFTEETH

19 (1st

stage)19 (2nd

stage)

78 (1st stage)64 (2nd stage)

MATERIAL EN 36 casehardeningsteel

EN 24

FACEWIDTH

15 mm ( 1st

stage)25 mm (2nd

stage)

15mm (1st

stage)25 mm( 2nd

stage)

FOS(TOOTHBENDING)

1.74 1.85

FOS(CONTACTSTRESS)

1.15 1.07

DIAMETERAL PITCH

10 (for first stage) & 8( forsecond stage)

Fig 16: deflection analysis of intermediate shaftusing KISSsoft sftware

Power unit mountings:The engine was mountedover four rubber engine beds of jeep to dampen thevibrations transmitted to the roll cage. The gear boxwas mounted through delrin bushes for the samepurpose mentioned above.( see illustration no: )

CV joint plunge and articulation calculation:Extensive discussion was made with thesuspension team to suit the position and length ofthe camber links in the rear suspension, to reduceCV plunge and articulation required duringsuspension travel. The camber links were broughtcloser and parallel to each other such that the CVlied in the vertical midpoint of the camber linksArticulation needed: 16 degree during full droopand 5 degree during full bumpPlunge needed; 4 mm for the whole suspensiontravel(See illustration no: )

Weight:Engine: 24 kgsCVT: 5.5 kgsGear box: 15.2 kgsDrive shaft: 11 kgs (5.5 each)Hub: 1.9 kgsWheel: 18.8 kgs (two wheels in fully inflatedcondition)Total: 76.4 kgs

Braking System:The purpose of the braking system is to

increase the maneuverability by locking all the fourwheels at a time. It is required to statically anddynamically lock all four tires on both hard andloose surfaces. Although the previous designworked effectively it was decided that modification

Table 7: Gear Data

of our previous year’s braking system is necessaryin order to reduce the unsprung mass of the vehicleand to achieve 50-50 biasing. This year our aim isto reduce the weight of wheel assembly so afterstudying various available options to get effectivebraking we have decided to use hydraulic discbrakes on all the four wheels with a single in boarddisc at the rear. Honda Aviator’s disc and caliperhas been chosen for the front and pulsar disc andcaliper has been chosen for the rear. Hydraulic discbrakes have high performance, easy replacement,and comparatively less weight than other availableoptions. have the common circuit of brake fluid forall the four wheels. So if any one chamber fails wecan still apply the brakes to all the four wheels .Toreduce the unsprung mass, it was decided inboardbrakes at the rear would be used.

Fig 16: Front/rear Hydraulic Split

SPECIFICATIONS

PINION GEAR

NO. OFTEETH

19 (1st

stage)19 (2nd

stage)

78 (1st stage)64 (2nd stage)

MATERIAL EN 36 casehardeningsteel

EN 24

FACEWIDTH

15 mm ( 1st

stage)25 mm (2nd

stage)

15mm (1st

stage)25 mm( 2nd

stage)

FOS(TOOTHBENDING)

1.74 1.85

FOS(CONTACTSTRESS)

1.15 1.07

DIAMETERAL PITCH

10 (for first stage) & 8( forsecond stage)

Fig 16: deflection analysis of intermediate shaftusing KISSsoft sftware

Power unit mountings:The engine was mountedover four rubber engine beds of jeep to dampen thevibrations transmitted to the roll cage. The gear boxwas mounted through delrin bushes for the samepurpose mentioned above.( see illustration no: )

CV joint plunge and articulation calculation:Extensive discussion was made with thesuspension team to suit the position and length ofthe camber links in the rear suspension, to reduceCV plunge and articulation required duringsuspension travel. The camber links were broughtcloser and parallel to each other such that the CVlied in the vertical midpoint of the camber linksArticulation needed: 16 degree during full droopand 5 degree during full bumpPlunge needed; 4 mm for the whole suspensiontravel(See illustration no: )

Weight:Engine: 24 kgsCVT: 5.5 kgsGear box: 15.2 kgsDrive shaft: 11 kgs (5.5 each)Hub: 1.9 kgsWheel: 18.8 kgs (two wheels in fully inflatedcondition)Total: 76.4 kgs

Braking System:The purpose of the braking system is to

increase the maneuverability by locking all the fourwheels at a time. It is required to statically anddynamically lock all four tires on both hard andloose surfaces. Although the previous designworked effectively it was decided that modification

Table 7: Gear Data

of our previous year’s braking system is necessaryin order to reduce the unsprung mass of the vehicleand to achieve 50-50 biasing. This year our aim isto reduce the weight of wheel assembly so afterstudying various available options to get effectivebraking we have decided to use hydraulic discbrakes on all the four wheels with a single in boarddisc at the rear. Honda Aviator’s disc and caliperhas been chosen for the front and pulsar disc andcaliper has been chosen for the rear. Hydraulic discbrakes have high performance, easy replacement,and comparatively less weight than other availableoptions. have the common circuit of brake fluid forall the four wheels. So if any one chamber fails wecan still apply the brakes to all the four wheels .Toreduce the unsprung mass, it was decided inboardbrakes at the rear would be used.

Fig 16: Front/rear Hydraulic Split

SPECIFICATIONS

PINION GEAR

NO. OFTEETH

19 (1st

stage)19 (2nd

stage)

78 (1st stage)64 (2nd stage)

MATERIAL EN 36 casehardeningsteel

EN 24

FACEWIDTH

15 mm ( 1st

stage)25 mm (2nd

stage)

15mm (1st

stage)25 mm( 2nd

stage)

FOS(TOOTHBENDING)

1.74 1.85

FOS(CONTACTSTRESS)

1.15 1.07

DIAMETERAL PITCH

10 (for first stage) & 8( forsecond stage)

Table8: Braking System

BODY PANEL AND SAFETY

Body panels: The body panels that cover the areabetween the lower frame side member and the sideimpact member are made using a custom madecomposite material, rexine cloth sandwiched inbetween two sheets of 2mm thick acrylic plastic,where rexine gives rich look and plastic providesrigidity. This all new composite material was madeto reduce the weight of body panels and also toavoid the breakage or cracking of plastic duringimpact by foreign objects. The panels are made foreach section and are fastened to the roll cage toenable quick and cheap replacement.

Safety: The safety of the driver is given the utmostimportance by providing all the safety equipmentssuggested by the SAE BAJA 2013 Rulebook suchas the helmet, neck support, goggles, balaclava,five point seatbelt, arm restraint, head restraint,driver suit, fire resistant socks and shoes, gloves,fire extinguisher, roll cage padding, sufficient roomfor driver inside roll cage cockpit, kill switch, power-train guards, bright warning lights, sound buzzers,reflective stickers around the vehicle, etc.In addition to the conventional techniques forensuring the safety of the driver, the specialinnovation in the buggy helps the driver to sendSOS signal across miles and SOS message to thenearest emergency response unit on the press of asingle button.

Safety FeaturesSOS :An SOS button is placed on the reach of the drivernear the kill switch so that the driver can press it tocall for instant help in case of any kind ofemergency situation for the driver or for the car.When the SOS button is pressed, the vehicle sendsan emergency service request SMS with thedriver's name, blood group and the GPScoordinates of the vehicle, twice to three predefinedmobile numbers.While the SMS is being sent, the lights and thebuzzers of the car sends SOS signals by flashingall the lights and buzzing all the buzzers in the carin morse code SOS pattern, so that anybody in andaround the area of the vehicle location can learnthat the driver is in an emergency.

Fig 17:Electronics Flow Chart

InnovationObjective:

To reduce the wakeregion at the backandtoreduce the pressure drag.

To supply engine with high density of air. To remove the heat generated and to clean

thedust particles present in the engine bay.

Design consideration:The amount of heat generated by the

engine due to prolonged running condition and thelimited air circulation since it is a rear engine whichis placed just behind the fire wall is the mainconsideration according to that air duct is designedto cool the engine and to supply high density of airto the air filter. If there is not enough sufficient aircirculation in engine bay it leads to hightemperature at the engine bay and hence it willresult in low density of air at high temperature. Airducts will rectify this problem by deflecting air fromoutside vehicle dimension to the engine bay cooling

Front Aviator disc - Pulsar caliper-190mm

Rear Inboard - Pulsar disc & caliper-240mm

Caliper type 2 cylinder floating typePedal effort 400 NPedal ratio 5:1Req brakingtorque

front: 179.22 N-m/rear: 181.47N-m

load on wheels(static)

1294.9 N (front), 1942 N (rear)

Load on wheelsduring braking

2105.4 N (front), 1131.8 N (rear)

Biasing 50:50Stoppingdistance

15 m (60 kmph – 0 kmph)

DecelerationStopping time

9.27 m/s2 \1.79 s

Master cylinder Maruti 800 OEM -Bore dia 19.4mm

that area and hence forth density of air will be highat low temperature.

Inorder to delay the flow separation to occurpearl drop model along with body fit spoiler is usedat the downstream side of the vehicle.

STATIC AND DYNAMIC CFD RESULTS:

Fig 18: Distribution of static pressure in various parts.

CFD analysis is done using ANSA, TGrid andAnsys Fluent software’s. Surface mesh isgenerated using ANSA, Volume Mesh is generatedusing TGrid, and Flow Analysis is done in AnsysFluent. The inlet is assumed tobe velocity inlet withvelocity magnitude of 15.777m/swhereas the outlet is assumed to be atambientcondition. As geometry is pretty complexwe havecreated tetrahedron locally and structuralhexahedron inthe remaining area of the wind tunnelwhich will increasethe accuracy of the solution.Static pressure fig shows the high pressure area atupstream side of the vehicle.

Fig 19:Pearl Drop Model

Fig20: above shows the delayed flow separationoccurred due to pearl drop model along withreduced wake region.

Fig21: above shows the velocity vectors on theengine bay due to the use of air ducts on both sidesof the vehicle.

CONCLUSION

When undertaking any design project thereare several factors to be considered that arecommon to all engineering projects. A project musthave a proper scope with clearly defined goals. Thegoal of this year’s team was the redesign of therear suspension and drive train. The drive trainbecame much more accessible which was thebiggest problem with the vehicle in the previousyear. Second the frame could be triangulated. Thisnot only improved its strength but also it’sappearance. The car has a much more clean andsymmetry look. The team was also able to improveupon almost every other system in the car. Theseimprovements were made possible from theexperience and lessons learned during the previousyear of competition. It is hoped that this year’svehicle is able to outperform the previous year’svehicle in every way. The more responsivesuspension and aggressive steering will make for amore competitive vehicle. It is also hoped that theperformance and reliability of several other systemshas been improved such as the new pedalassembly. Finally the Phoenix racing Baja Teamhopes that the heightened appearance of thisyear’s car, along with its improved performance willmake it a more noticeable competitor at this year’scompetition.

ILLUSTRATIONS

STATIC AND DYNAMIC TESTINGS OF FRAME

Fig 22:FEA Analysis of Frame

ILLUSTRATIONS

STATIC AND DYNAMIC TESTINGS OF FRAME

Fig 22:FEA Analysis of Frame

ILLUSTRATIONS

STATIC AND DYNAMIC TESTINGS OF FRAME

Fig 22:FEA Analysis of Frame

KNUCKLES AND HUBS

Fig23: Front Knuckle Analysis

Fig24: Front Hub

Fig25: Rear Knuckle Analysis

DRIVE TRAIN ASSEMBLY:

Fig 27:Drive Train Assembly

CV PLUNGE CALCULATION

Fig 28:CV Plunge

REFERENCES

1. Fundamentals of vehicle dynamics by ThomasD. Gillespie

2. Race Car Vehicle Dynamics by Milliken andMilliken

3. Clutch Tuning Handbook by Olaav Aaeon4. Manual Gearbox Design by Alec Stokes5. Heat Treatment of Gears by A.K .Rahit6. Shigley’s Mechanical Engineering Design by

budynas – nisbett7. Tune to Win by Caroll Smith8. Kisssoft Manual9. A kinematic Analysis and Design of

Continuously Variable Transmission byChristopher Ryan Willis

10. Phoenix Racing 2012 mini baja design report