Project Report on ATV prototype BAJA SAE INDIA

7/21/2019 Project Report on ATV prototype BAJA SAE INDIA

1/60

Design Anal ysis And Optimization Of Al l -ter r ain Vehicl e(ATV)

Page 1

CHAPTER N0. 1

INTRODUCTION

We approached our design by considering all possible alternatives for a

system & modeling them in CAD software like CATIA, AutoCAD etc. to obtain a

model with maximum geometric details. The models were then subjected to analysis

using Analysis Work Bench 14 software. Based on analysis results, the model was

modified and retested and a final design was frozen.

Dynamics analysis was done in Lotus suspension analysis software. The

aim was to optimize suspension variables to improve maneuverability. Theoretical

calculations of performance characteristics were also done. Extensive weight

reduction techniques were followed at every stage of the design to improve

performance without sacrificing structural integrity.


2/60


Page 2

CHAPTER N0. 2

DESIGN CRITERIA FOR THE VEHICLE & METHODOLOGY:

TABLE NO.- 2.1

As shown in above table, special considerations were given to safety of the

occupants, ease of manufacturing, cost, quality, weight, and overall attractiveness.

Other design factors included durability and maintainability of the frame.

NO. CRITERION PRIORITY

1 Reliability Essential

2 Ease of Design Essential

3 Performance High

4 Serviceability High

5 Manufacturability High

6 Health and Safety High

7 Lightweight High

8 Economic/Low

Cost

Desired

9 Easy Operation Desired

10 Aesthetically

Pleasing

Desired

REQUIRMENTS :-

Low Weight Vehicle.

Better Economy.

Better Comfort And Durability.

DESIGN AND CAD WORK :-

Collection Of Data And Calculation.

CAD And CEA Work of the

Subsystems.

REVIEW AND IMPLEMENTATION :-

Design Review And Project Plan.

Maintaining Quality in Fabrication.

Follow up And Project Plan.

DFMEA AND VALIDATION :-

Maintain DFMEA And DVP.

Validate of The Vehicle For Designed

Aspect.

TESTING :-

Testing The Vehicle For All the

Terrains.

Expecting Failures And Correcting

Them.


3/60


Page 3

CHAPTER N0. 3

ROLL CAGE :

The purpose of the Roll cage is to provide a safe environment for the

occupant while supporting other vehicle systems. Several steps were taken to ensure

this objective was met. For the frame design, we focused on a lightweight and safe

frame that still meets all of the requirements set forth by SAE. Special considerations

were given to safety of the occupants, ease of manufacturing, cost, quality, weight,

and overall attractiveness. Other design factors included durability and maintainability

of the frame.

The frame design incorporated bends instead of miters in many of the

structural members, believing that this allowed for faster construction, and increased

material strength from cold working resulting in an overall increase in product

quality. Although there was added cost associated with out-sourcing tube bending,

this cost was offset by a reduction in fabrication man hours through decreasing the

amount of mitered and welded joints and eliminating man hours and material needed

to fabricate fixtures for fit-up ,The Roll cage consists of two main criterions as

follows:

3.1 MATEARIL SELECTION:

The materials used in the cage must meet certain requirements of geometry

and minimum strength requirements found in SAE. Since the frame is being used in a

racing vehicle rather than a recreation vehicle, weight and cost is a very large factor in

the shape and size of the frame. The proper balance of strength, weight and cost is

crucial for the teams overall success.


4/60

Design Anal

TABLE N

In addition to th

availability of the mater

best suitable for the ro

satisfying the bending sti

was decided to be 2.1

chassis and for the sec

thickness of 2.1 mm

3.2 FINE ELEMEN

In order to opti

Work Bench 14 was us

six analysis tests conditi

Torsional ansys heave a

After running a

member. After having ad

constraints was complete

the safety of the roll ca

which is tabulated as foll

Materia

l

UTS UY

Mpa Mp

1 AISI

4130

560 45

2 AISI

1020

394.7 294.

3 AISI

1018

440 38

sis And Optimization Of Al l -ter r a

.- 3.1 GRA

e above table, selection depended mainly on

al. From the above tables, we concluded that

ll cage with economical cost and easier av

ffness criteria and bending strength the thickne

m for the O.D. of 28 mm for the primary m

ndary members, O.D. was selected as 25.4

ANALYSIS:

ized the strength, durability and weight of Ch

ed to analyze the chassis for all six loading c

ns are Front Impact, Side Impact, Rollover

d the loading on the frame from the front and r

ll five analyses it was found that there is a nee

ded these members, a second analysis using ide

d and results of these tests are shown in table; f

ge, proper analysis was done in the ANSYS

ows:

Elongation Youngs

modulus

% Gpa

21.5 210

7 36.5 200

15 205

n Vehicl e(ATV)

Page 4

H NO- 3.1

the cost and

ISI 1018 was

ilability. For

ss of the pipe

mbers of the

mm with the

ssis Analysis

ndition. The

Impact, and

ear shocks.

of additional

ntical loading

or confirming

Workbench


5/60


Page 5

(Assuming the total weight of the vehicle is 320 kg)

Setting up the analysis:

Component Material

Ultimate

Tensile

Strength

(MPa)

Yield

Strength

(MPa)

Modulus of

Elasticity

(GPa)

Percentage

Elongation

(%)

Hardness

(BHN)

Roll Cage 1018 steel 450 380 265 16 130

Hub 6082 Al alloy 225 186 70 12 75

Adapter EN8 660 530 206 7 120

TABLE NO.- 3.2

TABLE NO.- 3.3

RESULTS:

TABLE NO.- 3.4

DETAILS MAX

FORCE

MAX

FORCE

TIME OF

IMPACT

(kN) (in terms of gs) (s)

1 Front impact 30 10 0.2

2 Side Impact 9 3 0.2

3 Roll Over Impact 6.4 2 0.2

4 Torsional

analysis

1.88 3 FRONT -

2.82 3 REAR -

DETAILS MAX STRESS MAX

DEFORMATION

FOS

(Mpa) (mm)

1 Front impact 385.49 3.67 1

2 Side Impact 303.09 1.02 1.2

3 Roll Over Impact 272.64 4.74 1.3

4 Torsional ansys - 1.84(F) 3.64(R) 1.26


6/60

Design Anal

3.3.1 FRONT IMPA


CT: (8-10G)

n Vehicl e(ATV)

Page 6


7/60

Design Anal

3.3.2 SIDE IMPAC


:(3G)

n Vehicl e(ATV)

Page 7


8/60

Design Anal

3.3.3 ROLL OVER:


(2G)

n Vehicl e(ATV)

Page 8


9/60

Design Anal

3.3.4 TORSTIONA


ANSYS:(3G FOR FRONT AND RE

n Vehicl e(ATV)

Page 9

R)


10/60


Page 10

CHAPTER N0. 4

SUSPENSIONS:

4.1 FRONT SUSPENSION:

The problem that was encountered was to design a competitive front

suspension for the ATV . To do this the operating conditions of the competition had to

be researched, and from that design considerations had to be decided

Consideration Priority Reason

Simplicity Essential Main objective

Lightweight Essential Lower weight means Faster car.

10 of travel High To ensure ground contact always.

Durability High It should be durable and reliable for

any condition.

Shock Absorbing Desired High shocks in the front.

Adjustable Desired To adjust camber, toe in and out for

improving handling.

Compatibility with

Steering

Desired It must be compatible because

suspension geometry is linked with

steering geometry.

From the above considerations to balance weight and cost savings for the

manufacturers, and comfort and handling for the customer, several options for front

suspensions were analyzed. For the best handling characteristics the front wheels must

always be in perpendicular contact with the ground. Bump steer and camber gain must

be minimized in both ride and roll changes. Two possibilities for the front suspension

were a double a-arm and a single arm McPherson Strut suspension. The double a-arm

suspension is the most feasible design according to our design, thus double A- arms

were selected for the front suspension. To design the front suspension several

software packages were utilized to ensure the best possible results. LOTUS SHARK


11/60

Design Anal

simulation software was

wheel travel.

Fig. No.4.1 :-

Fig. No.4.2 :- FR

The front

as possible, while maintmaintained by welding

within a tolerance of 1 m

The variati

obtained from Lotus shar


used to create simulations for both parallel

SIMULATION OUTPUT AND ROLL CENT

NT SUSPENTION ASSEMBLY EXPLODED

uspension arms were designed to be as easy t

aining the high strength as desired. The Builhe A arms mounting brackets at the designe

m

on of the toe angle and camber with respec

k

GRAPH NO- 4.1

n Vehicl e(ATV)

Page 11

and opposite

R

VIEW

manufacture

quality wasd hard points

to bump as


12/60

Design Anal

4.2 REAR SUSP

There wer

process of designing an

trailing arm design with

inches of travel, and are

about half way up the

travel while staying with

that trailing arms were

trailing arms allow fo

interference by the susp

for maintenance, and

Consideration Priority

Simplicity Essential

Lightweight Essential

8 of travel High

Durability High

The rear suspension geo

below:

Fig. No.4.3 :- RE


NSION:

many objectives and considerations to look

building the rear suspension. The rear suspe

only one arm per side. The Fox Float air s

ounted near the bearing carrier, near the end o

ear main roll hoop. This allows for maximu

in the range of the rear axle CV joint travel. A

sed was that the drive train design was to be

the full drive train assembly to be rem

nsion. This enables the drive train to be pulle

eeps the overall design of the rear of th

Reason

Easier to fix, build, design, analy

Lower weight means Faster car.

To ensure ground contact always

Withstand abusive driving during the endu

etry and modeling was done in Catia and it is a

R SUSPENTION ASSEMBLY EXPLODED

n Vehicl e(ATV)

Page 12

at during the

sion is a full

ocks have 6

f the arm, and

suspension

nother reason

modular. The

ved without

from the car

e car simple

e.

.

rance race.

s shown

IEW


13/60


Page 13

The car has been driving for two weeks, there has been testing done to

see if the suspension reacts the way intended by design. It turns out that the design of

the rear suspension is working as well or better than expected.

TABLE NO.- 4.1

4.3 FOX RACING SHOCKS:

Right from the beginning we focused on reducing the weight of the

vehicle. The customized Spring and damper assembly of the vehicle was way too

bulky to be used in ATV, thus team emphasized on Fox Shocks which reduced the

weight of the vehicle to a large extent and provided easy adjustable stiffness to the

shocks.

From the market survey, the fox shocks were selected on the following criteria:

Travel of the Shocks.

Total extended Length of the shocks.

Cost and availability.

Thus, FOX FLOAT 2 air shocks were selected and procured. It provides 6 inches of

travel and 19.8 inches of extended length, which is perfect from our design point of

view.

Parameter Values

Front

Suspension

Rear

Suspension

Wheel

Travel254 mm 206 mm

Wheel Rate 9.294 N/mm 19.90 N/mm

Jounce 117.4 mm 117.4 mm

Rebound 39.14 mm 39.14 mm

Camber

Gain1.85 0

PARAMETERS VALUES

Caster 6

Kingpin inclination 14

Static Camber Set as Zero

Static Toe In Set as Zero

Roll angle @Speed 30 km/hr

Roll Angle 172

Turning Radius 5m

Weight Transfer 90.77kg


14/60


Page 14

CHAPTER N0. 5

STEERING:

On the rough terrains it is very essential to have the steering must be lightand should give quick response on turns. The design considerations are as follows:

CONSIDERATION PRIORITY REASON

Simple Design Essential Easy to repair during

competition

Light Weight Essential Lower weight means

Faster car.

Low Steering Ratio Essential Quick steering response

Ackerman geometry High To make understeer.

Minimize Bump steer Desired Conserve momentum

while

Steering

Rack and Pinion steering system was selected due to its easy availability, easier

maintenance, feasibility to modifications and the cost. Most of the analysis was

focused on the steering system. The primary focus was on decreasing the steering

effort. The team also focused on decreasing the amount of steering wheel travel, and

increasing the steering responsiveness.

In the normal rack & pinion vehicle the driver had to turn the steering

wheel 540 to bring the wheels from the center to lock. The driver had to remove his

hand from the wheel at least once to complete the turn. The goal was to allow the

driver to use only 290 of steering wheel travel from the center to maximum wheel

travel. The goal was accomplished by using a REDUCTION GEARBOX after the

pinion .A new system provided a motion ratio of 6.5 to 1, or 70 mm of rack travel

per revolution of steering wheel travel. The higher ratio rack has inherently larger

steering effort, however using a longer moment arm tie rod mount offset this effect.

The Ackermann angle was selected by analyzing wheel angles from previous years.


15/60

Design Anal

The steering calculation

TABLE NO.- 5.1

Fig.

PARAMETERS

Rack travel(mm)

Steering Wheel lock

from centre

Turning circle

Radius(m)

Scrub Radius (mm)

Steering Ratio

Steering Effort (N)

Percentage Ackermann

Tie rod Length (mm)


s are tabulated as:

Fig. No. 5.1 :- STEERING

No. 5.2 :- ACKERMANN GEOMETRY

VALUES

57

109

2.48

36.57

6.59

108

98.99

400

Fig. Ackermann geometry

Fig. St

asse

n Vehicl e(ATV)

Page 15

ASSEMBLY

ering

bl


16/60

Design Anal

5.1 STEERING RA

In the normally v

the wheels from the centat least once to complet

only 290 of steering wh

was accomplished by usi

This gearbox con

and the other smaller ge

to the column of the stee

by the Universal joint.

It is as shown be

Fig. No. 5.3 :- REDUCT

GEARBOX EXPLODE


IO REDUCTION GEAR BOX:

ehicle the driver had to turn the steering wheel

r to lock. The driver had to remove his hand fthe turn. The goal for 2014 was to allow the

el travel from the center to maximum wheel tr

g a REDUCTION GEARBOX after the pinion

sists of two gears: One bigger gear with diam

r with the diameter of 35.5 mm. The bigger g

ing wheel and the smaller gear is attached to t

ow:

ON TABLE NO.- 5.2

VIEW

Part

Witho

Reducti

gearbo

Steering Ratio 13:1

Rack travel per revolutionof steering wheel

35 m

Required Rack travel

(Centre to lock)70 m

Rotation of steering wheel

(Centre to lock)540

Steering Effort 68 N

n Vehicl e(ATV)

Page 16

540 to bring

om the wheeldriver to use

vel. The goal

.

ter of 68 mm

ar is attached

e pinion side

t

on

Without

Reduction

gearbox

6.5:1

70 mm

57 mm

290

108 N


17/60


Page 17

CHAPTER N0. 6

BRAKES:

The braking system for the vehicle is responsible for stopping the vehicle at alltimes and is integral for the drivers safety. That why the brake must be capable of

locking all the four wheels when applied so we incorporated disc brakes in the front

and rear.

CONSIDERATION PRIORITY REASON

Simplicity High Overall goal of vehicle.

Light Weight High Lightweight parts to

minimize total weight.

Performance High Capable of decelerating a

320 kg vehicle.

Reliability Essential Reliable to provide hard

braking always.

Ergonomics Essential Optimal pedal assembly

fitment to suit every

driver.

According to the rim size and the braking calculations we chose to use Bajaj

Discover ST discs that will be mounted on the hub in the front. Disc brakes were

chosen because of the ace of compatibility, the availability of the replacement parts

and the overall effectiveness that the system provides.

For the rear design, rear disc brakes of Bajaj Pulsar 220 were used. It provided

the required diameter of the disc and the required braking torque could be achieved.

The design calculations are tabulated as follows:


18/60


Page 18

PARAMETERS FRONT REAR

Outer diameter(Custom) 190 218

Effective Rolling

Radius(mm)

81 95

Thickness(mm) 3.47 3.47

Material Perlite Grey Cast Iron

Radius Of Gyration(mm) 170 280

Moment of Inertia(kg/m^2) 0.289 1.176

Calliper BAJAJ DISCOVER 125 ST

Calliper Piston

Diameter(mm)

28

Coefficient of friction 0.45

Tandem Master Cylinder Maruti 800

TMC diameter(mm) 19.05

TABLE NO.- 6.1

PARAMETERS VALUES

Braking distance(m) 17.66

(Deceleration 0.8kg )

Pedal Force(N) 130

Pedal Ratio 1:4

Inline Fluid Pressure 0.5bar

Dynamic load Transfer(kg) 83.63

Single Stop Temp. Rise(c) 22.5

TABLE NO.- 6.2


19/60


Page 19

CHAPTER N0. 7

POWERTRAIN:

The goal of the drive train is to transfer power from the engine of the vehicleto the wheels. The power transferred must be able to move the vehicle up steep grades

and propel it at high speeds on level terrain. Acceleration is also an important

characteristic controlled by the drive train. Calculations were done according to the

considerations, looking at gear ratios, engine power and wheel size. After the

calculations were re verified no reduction is to be given was decided. Hence direct

line was given. Also during design, the angle of the propeller shafts was taken care.

The drive train for the car has been radically overhauled to improve overall

car performance and correct vulnerabilities. The Drive Train Based of Mahindra GIO

was used based on the traction and speed calculations. The system benefited with

simplicity and low cost.

GIO transmission was used in forward configuration, this year to enhance

torque the transmission is used in Reverse configuration. It can be tabulated below :

GEARBOXMULTIPLATE

CLUTCH

GEAR RATIOS Initial

Tractive

Effort

(N)

Acceler

ation

(m/s)G1 G2 G3 G4 R

PIAGGIO APE

PASSENGERYES 25.52 15.16 9.25 5.96 30.62 1702.8 2.80

MAHINDRA

ALFA CHAMPIONYES 31.48 18.7 11.4 7.35 55.08 2100 3.76

MAHINDRA

ALFA

PASSENGER

YES 25.52 15.16 9.25 5.96 30.62 1702.08 2.80

TATA NANO NO 27.6 15.6 10.08 6.64 31.42 1841.58 3.14

MAHINDRA GIO YES 27.66 14.86 8.48 5.55 33.66 1845.58 3.15

MAHINDRA GIO

IN REVERSEYES 33.66 18.08 10.32 6.76 27.66 2245.93 4.11

FORCE MINIDOR

PICK UPNO 24.42 14.58 8.22 4.8 23.4 1629.40 2.63


20/60

Design Anal

AUTORICKSHAW

MAHINDRA

CHAMPION

The transmission was

is explained below.

Specification

Fig. No. 7.1

Gear ratio 4.979

Overall gearratio

4.925

Max.

Velocity54 km/hr

Max. Torque 586.18

Clutch type Multipla

Gearbox type Trans-ax

Shifter type sequenci

DRIVELINE

2 CVJ C

Sleeves


ES 23.55 18.8 8.66 5.27 27.98 15

NO 25.081 15.12 9.33 7 23.54 16

TABLE NO.- 7.1

coupled directly to the engine with a adapter.

:- ADPATER ASSEMBLY EXPLODED VIE

@Top gear

m @ First gear

e wet type clutch

le Constant mesh gearbox

l single wire shifter

nnection Stock maruti 800

n drive shaft for length correction

n Vehicl e(ATV)

Page 20

71.35 2.49

73.50 2.73

his assembly


21/60

Design Anal

7.1 ANALYSIS OF


HE ADAPTER:

TABLE NO.- 7.2

GRAPH NO.- 7.1

PARAMETERS V

Max. Equivalent

Stress

193.

Max. Shear Stress 104.

Max. Deformation 0.

Factor of safety

n Vehicl e(ATV)

Page 21

LUES

46 mpa

46 mpa

3mm

.74


22/60

Design Anal

CHAPTER N0. 8

WHEEL ASSEMB

In an all-terrain v

steering and getting theand rotational of inertia

has low weight and low

various surfaces and be c

For the front, s

control. Therefore, tires

diameter of the rim will a

For the rear, req

with specifications of 25

The Front wheel

Bearings were selected a

used. The Rear wheel h

disc onto the hub.

Fig. No. 8.1 :- FRON

8.1 BODY PANEL

To reduce the w

Mild steel sheets. These

sheet of 0.7 mm thickne

weight. For the side pane


IES AND BODY PANELS:

ehicle, traction is one of the most important a

ower to the ground. Tire configuration, treadre critical factors when choosing proper tires.

internal forces. In addition, it must have stro

apable of displacing water to provide power wh

aller diameter tires were used to allow bet

with specifications of 21x7x10 were selected.

llow the brake components to fit inside the whe

irements are better traction and larger diame

10x12 were selected.

ub was made from Aluminum this year to redu

ccording to required design, thus, Maruti Alto

b is the OEM part and modifications were mad

AND REAR WHEEL ASSEMBLY EXPLOD

:

ight of the vehicle, aluminum sheets were u

sheets were bolted to the chassis. For the fire

ss was used. This provided the required streng

ls, aluminum sheet of 0.5 mm were used.

n Vehicl e(ATV)

Page 22

pects of both

epth, weight,The ideal tire

g traction on

ile in mud.

ter maneuver

The 10-inch

el.

er, thus, tires

e the weight.

earings were

e to assemble

ED VIEW

ed instead of

all aluminum

th with lower


23/60


Page 23

CHAPTER N0. 9

ERGONOMICS AND SAFETY:

Ergonomics is the science of equipment design intended to maximize

productivity by reducing driver fatigue and discomfort. The ergonomics aspect of theSAE Baja vehicle is crucial in ensuring that the car will both meet all of the rules

stated in the SAE rule book as well ensuring that all of the components of the car will

function properly when assembled together. It is essential that each member of the

team is able to safely and comfortably operate the vehicle.

ParameterStd.

Value

Design

ValueParameter Std. Value

Design

Value

Angle at

elbows110-130 110

Steering Wheel

Dia. (mm)- 320

Angle at knees 120-150 130Angle of Steering

Wheel20-45 20

Clearance from

vehicle

(inches)

- 4 Head Clearance - 6.5

TABLE NO.- 9.1

Drivers should be able to experience fast pace, exciting racing without

risking major injury. Car 80 meets or exceeds all of the minimum safety requirements

composed by the Society of Automotive Engineers and the event coordinators. In

addition, a number of safety features have been added to further reduce the possibility

of personal injury.

An LED brake light warns other drivers of deceleration.

A safety helmet and neck support protects the driver.

A Six-point safety harness keeps the driver adequately restrained.

Roll cage padding protects drivers head from impact.

The remaining standard safety equipment, including arm restraints, fire

extinguisher, and two kill switches were all placed for easy access and use, as well as

maximum optimization of their functions during an emergency.


24/60

Design Anal

CHAPTER N0. 10

ELECTRICALS:

The electrical s

electrical circuitry is to bswitch.

The electrical cir

Fi


stem was proposed to work on many road

e done mainly for the brake light, horn, reverse

uit for the vehicle is as shown below:

. No. 10.1 :- ELECTRICAL CIRCUIT

n Vehicl e(ATV)

Page 24

vehicles. The

light and kill


25/60

Design Anal

CHAPTER N0. 11

CONCLUSIONS:

The Team used

and prototype constructiSeveral team members

gather ideas and informa

could be incorporated int

After initial testin

in this years competitio

and durability of all the

the leaves for the compet

Fig. No. 11.1

ERGONOMI

CONSIDERA


xtensive physical testing, hours of simulation

n to create a vehicle that is fast, maneuverablettended the time to time workshops arranged

ion about what design choices were successful

o our design.

it can be seen that our design should be a stro

n. There will be extensive testing done to pro

ystems on the car and make any necessary ch

ition.

:-

CS

ION

Fig. No. 11.2 :- PVC M

n Vehicl e(ATV)

Page 25

and analysis,

, and reliable.by BAJA to

and how they

g competitor

ve the design

nges up until

OCKUP


26/60

Design Anal

CHAPTER N0. 12

FRONT VIEW

TOP VIEW


CAD MODELS:

SIDE VIEW

REAR VIE

n Vehicl e(ATV)

Page 26


27/60

Design Anal

CHAPTER N0. 13

DESIGN FAILURE

SYSTEM COMPONE

NT

P

IA

FE

Suspension

TrailingArm

T

B

Knuckle K

Di

on

Power

Train

Drive Shaft

D

endi

ial

M

wfai

Steering Pinion PiF

Brake

Tandem

MasterCylinder

F

ofro

Lof

fr

fl

Engine Engine Air

Intake

Cl

gfil

Pedal Fof

li


MODE EFFECTIVE (DFME) ANAL

TENT

L

ILUR ODE

S O D R

P

N

ACTION TAKEN

rsion 7 7 3 147

Optimum designconsideration to

reduce tensional&bending force; co-

linear line of actionof wheel& spring

centerline.

nding

uckle

slocati

7 6 5 210 Soldering of

Knuckle

tachm

t fromferent

8 6 4 192 Perfect Shaft

length

uff

ldlure

10 8 4 320 Align using V

block& Drill holein Muff coupling

for excess weldmaterial

Penetration .

ionilure

9 5 2 90 Replacement ofPinion

ilure

Pushd

8 3 6 144

Push rod should

have adequateD.O.F; co-linear of

line of action thepedal & pushrod.

akageoil

m

id line

9 4 3 108 Refill of brake oil

oggin

f Airter

9

7 4

252 Rerouting of Air

intake above thedriver seat through

the firewall

7 196

8 224

ilure

kage

10 4 3 120 Replacement oflinkage

n Vehicl e(ATV)

Page 27

SIS:

S O D R

P

N

7 4 2 56

7 2 2 28

8 2 3 48

8 3 3 72

9 2 2 36

8 2 4 64

9 2 2 36

10

2 3

60

7 42

8 48

10 2 2 40


28/60


Page 28

CHAPTER N0. 14

DESIGN VALIDATION PLAN

System Parameter Method Of Checking Design

Value

Steering

Minimum

Turning

Circle

Diameter

Vehicle is to be taken to a surface

with loose soil and a circle with

steering wheel locked at full travel

is to be negotiated. The distance

between the two diametrically

opposite points is to be measured.

4.9 m

Suspension

Spring

Stiffness

Load is to be added to the spring

while holding it in a vertical block.

Load required to cause unit

deflection is to be noted.

Front-

26.73N/mm

Rear-

40.10N/mm

Travel

The vehicle is to be loaded on a jack

and front wheels and spring are

removed. Damper is mounted in the

designated position and point of

maximum designed travel is

marked. Hub is moved upwards

manually till the point of maximum

travel. Difference between the

initial and final position of hub is to

be noted.

Jounce-

117.4mm

Rebound-

39.14mm

Transmission

Maximum

Gradient

Climbing

Move the vehicle over surface

having inclination of 100

.

Transmission to be set on first gear,

then allow the vehicle to climb the

33


29/60


Page 29

Capacity slope. Then subsequently increase

the slope by 50 .

until vehicle would

not climb the slope. Previous angle

is measured.

Top speed

Vehicle is to be loaded on a jack.

Transmission is shifted to final gear

and full throttle is given for 20

seconds. Tachometer is to be held at

the wheels and maximum reading is

noted. Speed is calculated by using

noted rpm.

54 km/hr

BrakesStopping

Distance

A reference line is to made, from

which the driver is to start braking.

The vehicle is to be at a

predetermined speed while crossing

that line. Maximum force is to be

applied on brake pedal when front

wheels cross the line. Distance is to

be measured from the line to the

front once the vehicle is brought to

a complete halt.

17.5 m

Roll Cage Weld Test

The welded joint is to be taken and

tested on a Universal Testing

Machine. The failure is to be

observSZed. Another method is to

take the welded joint and impactwith multiple hammer blows until

failure. Failure of weld or material

is noted.

Weld is seen

to fail.


30/60

Design Anal

CHAPTER N0. 15

TECHNICAL SPE

ENGINE

Type B

S

O

Displacement, cc 3

Max. Torque,

Nm @ rpm

1

Max Power,

kW @ rpm

7.

Transmission

Mahindra GioGearbox

4 forward 4 reverse

speed

Steering

Front D

Rear T

Brakes

Hydraulic Disc Brakes

DimensionsLength (mm) 2

Width (mm) 1

Height (mm) 1

Weight

Kerb Weight (Kg) 2

Gross Weight (Kg) 3

Wheel size

Front (inches) 2

Rear (inches) 2

Centre of GravityPosition w.r.t.

center of base of

firewall (mm)

X

Y

Z


IFICATIONS

Overall Performance Target

riggs &

tratton 10HP

HV

05

9 @ 2500

.5 @ 3600

ouble

ishbone

railing Arm

286

600

00

0

20

1x7x10

5x10x12

:109.33

:-45.82

:165.26

Light Weight Buggy

Best Driver Safety a

20% Front

Left Weight Distribut

30% Rear

n Vehicl e(ATV)

Page 30

s

d Ergonomics

20%

ion Right

30%


31/60


Page 31

CHAPTER N0. 16

TESTING:

FINAL VEHICLE

BRAKE TEST JUMP TEST


32/60


Page 32

CHAPTER N0. 17

References:

1 Vehicle Dynamics By Thomas D. Gillespie

2 Windsor

3 Mille ken & Millikenh

4 Automobile Engineering volume 1-volume 2 By Kirpal Singh

5 Google Search

6ARAI India


33/60


Page 33

APPENDIX A

SUSPENSION DESIGN

KINEMATIC ANALYSIS

Sample calculation front suspension-

A arm suspension:

The weight of the vehicle is 200 kg but because of 40:60 ratio of weight distribution

between front and rear suspension, the front weight is 80 kg.

F = 8049.81

=3139.2 N

K =?

?(?=allowable travel=117.4 mm)

=3139.2

117.4

= 26.73 N/mm

Wheel rate/ wheel travel = (kw)

Wheel rate is the actual rate of a spring acting at the tire contact patch

kw = ks (M.R)2 sin(s).for trailing arm(from internet reference)

kw = ks (M.R)2.for A-arm ( from Windsor as reference)

for an offroad vehicle ideal value is 8 to 12 inch

unit-N/m or lbs/inch


34/60


Page 34

CALCULATIONS:-

A) Front suspension-

A-arm double wishbone-

Motion ratio= 0.6

Shock travel = 152.4 mm

Motion ratio =????? ??????

? ???? ??????

wheel travel =152.4

0.6

wheel travel = 254 mm = 10 inch

Wheel rate (kw) = ks M.R2

= 26.73 0.6

2

= 9622.6 N/m

B) Rear suspension:-

Trailing arm suspension-

Motion ratio

d1= 369 mm

d2= 497 mm

M.R = d1/d2 =369

497= 0.74


35/60


Page 35

Motion ratio =????? ??????

? ???? ??????

Wheel travel =152.4

0.74

Wheel travel = 205.94 mm

Wheel rate (kw) = ks(M.R)2sin(s)

= 40.10 (0.74)2

sin650

= 19.90 N/mm

= 19901.39 N/m

Roll stiffness (k)

Amount of roll moment needed to roll the suspension by one unit of

rotation guidelines.

Roll stiffness (k) =??

??

? ??.?(from internet)

Unit Nm/deg or lbs/inch

Front suspension

A-arm double wishbone

k =?? ?

?? ??.?

=0.482 9622.8

2 57.3

= 19.34 Nm/deg


36/60


Page 36

Rear suspension

Trailing arm suspension

k =?? ??

? ??.?

t = 885 mm = 0.88 m

kw = 19901.39 N/m

k=0.882 19901.39

2 57.3

= 134.48 Nm/deg

Jounce

It is the upward movement or compression of suspension component.

? =?

?.from internet

Rebound

it is the downward movement or extension of suspension component.

Rebound : jounce = 3:1

Calculation:-


A-arm double wishbone-

Jounce(at 4g load) ? =?

?


37/60


Page 37

=49.8180

26.73

= 117.44 mm

Rebound (at 4g load) ? =?

?=

??.????

??.???

= 39.14 mm

B) Rear suspension:-

Trailing arm suspension-

Jounce(at 4g load) ? =?

?=

49.81120

40.10

= 117.42 mm

Rebound (at 4g load) ? =?

? = 39.14 mm

Natural frequency:-

The natural frequency is rate at which an object vibrates when it is

not disturbed by an outside force.

N.F =???

? ?????? ??????????..(from internet)

Ideal value = 1 to 1.5 Hz

N.F =?

??

? ???? ????

?????? ? ???


A-arm double wishbone suspension

N.F =???

? ?????? ??????????

= 1.45 Hz


38/60


Page 38

B)Rear suspension-


N.F =?

??

? ???? ????

?????? ? ???

= 1.6 Hz

Ride rate

The change of wheel load at thecentre of tire contact, per unit vertical

displacement of the sprung mass relative to the ground at a specific load.

Calculation:-

A). Front suspension-

A-arm double wishbone suspension-

Ride rate =??.??

?? + ??

=26.73 0.421

26.73+0.421

= 0.414

B). rear suspension :-


Ride rate =??.??

?? + ??

=40.10 0.51

40.10+0.51


39/60

Design Anal

= 0.503

Camber gain:-

The amount of angle cha

from the centre of car

A. Front suspension

Camber gain = 1 inch =

Camber gain = joun

= 117.4

= 156.5

= 6.16

= 2.460

Roll centre analysis


ge in front spindles as suspension travels inwa

:-

0.30

for front suspension

ce + rebound

4 + 39.14

8 mm

n Vehicl e(ATV)

Page 39

d or outward


40/60


Page 40

APPENDIX B

STEERING DESIGN

CALCULATIONS:-

A) STEERING TYPE : RACK & PINION

1. Rack Travel: 57mm

2. Steering Wheel Centre to lock Angle 290

3. Rack Used :- MARUTI 800

B.TURNING RADIUS

(R)2

= (R1)2

+ (C)

R = 2.48m

C. STEERING RATIO

S.R. = Steering Wheel Lock Angle / Road Wheel Angle

S.R = 6.59

D. STEERING EFFORT

S.E. = Weight On Front Wheel/ Moment Ratio


41/60


Page 41

S.E. = 108 N

E .TIE ROD LENGTH 400mm

F. PERCENTAGE ACKERMAN

% Ackerman = (Angle Of Inner Wheel Angle Of Outer Wheel) / Angle

Inside

Wheel For 100% Ackerman

% Ackerman = 98.99

Fig. Ackerman Geometry


42/60

Design Anal

G. Bump Steer Con

To minimize bgeometry.

Fi

H. steering reductio


ideration

mp steer, keeping tie rod parallel to Upper A-

g. Bump Steer Correction

gear box concept implemented in ve

Fig. Reduction box Exploded view

n Vehicl e(ATV)

Page 42

rm shown in

icle.


43/60


Page 43

Part Implementation in vehicle

Steering Ratio 6.5:1

Rack travel per revolution

of steering wheel

70 mm

Required Rack travel

(Centre to lock)

57 mm

Rotation of steering wheel

(Centre to lock)

290

Steering Effort 108 N


44/60


Page 44

APPENDIX C

BRAKE

CALCULATIONS:-

A. TMC

1. TMC- Maruti 800

2. Piston dia.- 19.05mm

3. Pedal ratio- 4:1

4. Pedal Force- 130N

B. Stopping Distance-

= V/2

= 17.63.

C. Pedal braking force-

=Total input to each TMC

=Pedal force*No.of TMC

=130*4

=520

D. Coefficient of friction

= = 0.45

E. Dynamic Load Transfer-

W =(/g)*w*(H/L)

=(0.8)*3500*(0.445/1.52) =819.73N

=83.56kg


45/60


Page 45

F. Inline Pressure-

P = F/A

= 520/((/4)(19.05))

= 18.25bar

G. Calliper/Brake force-

F =P*A

=(18.25*10^)*(1.23*10^-)

=2244.75N

H. Rolling radius-

T = F*R

But,

Actual torque = Ideal torque * Brake Force

=39 * 1.25

=48.75kg.m

Torque is divided by 2 wheels,

48.75/2

=24.375kg.m

R= T/F

= 24.375/2.44

Pitch dia. =10.86

Actual dia =217.2+28 . (28=calliper dia)

=245mm


46/60


Page 46

I. Temperature Rise-

=22.5c

J. Front wheel speci.-

1. Type-Disc brake(custom)

2. Size -21*7*10

3. Disc dia.=190mm

4. Front Pistriction-220mm

5. Disc thikness-3.4=4mm

K. Front wheel speci.-

1. Type-Disc brake(custom)

2. Size -25*10*123. Disc dia.=21.83

4. Front Ristriction-260mm

5. Disc thickness-4mm


47/60


Page 47

APPENDIX D

POWERTRAIN DESIGN

CALCULATIONS:-

Calculations for Gear Box Selection

Taking following Assumptions:

IE = 1= 1.3

tot = 1

rdyn = 0.29 m

All calculations for 1st

gear , taking available gear ratios in

consideration.

1. Piaggio Ape Passenger

FzA = Total Available Traction

Fzex = Excessive Traction =(FzA-FzB)

FzB = Total Driving Resistance=534

FzA = (Engine Torque Gear Ratio)(1000 rdyn) = (19.3525.52)(10000.29)

= 1702.8 N

FzEX = (FzA-FzB)

= 1702.8-534


48/60


Page 48

= 1168.8 N

Acceleration =FzEX =mF a

=a = ( FzEX) (mF )

= 1168.8(3201.3)

=a = 2.80 m/s2

Gradient Angle on 1st

gear

FzEX = mF gsin(ast)

=1168.8=3209.81 sin(ast)

= (ast )= 22.9

2.Gio In Reverse Calculations

FzA=(19.3533.66)(10000.29)

=2245.93 N

FzEX =FzAFzB =1711.93 N

Acceleration =FzEX = mF a

=a = 4.11 m/s2

Gradient Angle

1711.93= mF gsin(ast)

= (ast) =33.04

According to above calculations same procedure for following

vehicles.

1. Mahindra Alfa Champion

FzA= (19.3531.48)(10000.29)

= 2100 N


49/60


Page 49

FzEX = 1566.47 N

Acceleration (a)= 3.76 m/s2

(ast)= 29.93

2. Mahindra Alfa Passenger

FzA=1702.8 N

FzEX =1168.8 N

Acceleration(a)=2.80 m/s2

(ast)=22.90

3. TATA NANO

FzA=184.58 N

FzEX =1307.58 N


(ast)=24.60

4. Mahindra Gio

FzA=1845.58N

FzEX =1311.58 N


(ast)=24.69

5. Force Minidor Pick Up

FzA=1629.40 N

FzEX =1095.40 N


(ast)=20.42

6. Auto Rikshaw

FzA=1571.35 N


50/60


Page 50

FzEX =1037.35 N

Acceleration(a)= m/s2

(ast)=19.29

7. Mahindra Champion Passenger

FzA=1673.50 N

FzEX =1139.50 N


(ast)=29.28

FINAL VALUE FOR TRANSMISSION SURVEY CHART

Sr.

No.

Vehicle Name Initial

Tractive

Effort

Acceleration Gradient

Angle

1) Piaggio Ape Passenger 1702.8 2.80 22.9

2) Mahindra Alfa Champion 2100 3.76 29.93

3) Mahindra Alfa Passenger 1702.8 2.80 22.9

4) TATA NANO 1841.58 3.14 24.60

5) Mahindra Gio 1845.58 3.14 24.60

6) Mahindra Gio in Reverse 2245.93 4.11 33.04

7) Force Minidor Pick Up 1629.40 2.63 20.42

8) Auto Rickshaw 1571.35 2.49 19.29

9) Mahindra Champion

Passenger

1673.50 2.73 21.80

Transmission used in the vehicle on the basis of Accelerationand Traction

Mahindra Gio in Reverse Configuration


51/60


Page 51

Traction available on each gear of Mahindra Gio in

Reverse Configuration .

1

st

Gear = FzA1 = (19.3532.66)(10000.29) =2245.93 N

2nd

Gear = FzA2 = (19.3518.08)(10000.29)

=1206.37 N

3rd

Gear = 688.59 N

4rt

Gear = 451.055 N

Gradient on each gear when used in reverse configuration

1st

Gear = (ast) = 33

2nd

Gear= (ast) = 12.38

2nd

Gear= (ast) = 3

3rd

Gear= (ast) = 2.82 -1.51

Reverse Gear = 24.69


52/60


Page 52

Speed Calculations On Each Gear-

V4 (kmph)= [(3.6/30max)rdyn]iA iE

= [3.6 /3037000.283]3.761.1

V4 = 53.08 kmph

(n4 = 885 rpm)

V3 (kmph) = [3.6 /3037000.283] 10.321.1

V3 = 34.77 kmph ( n3 = 580 rpm)

V2 (kmph)= 19.84 kmph

(n2 = 331 rpm)

V1 (kmph)= 10.66 kmph

( n1 =176.66 rpm= Roadwheel rpm)

Torque Available on 1st

Gear =G1 Max Torque of Engine

T1 =33 19.35

T1 =638.55 N.m

Using 2 CVJ joints at Gear Box side Maruti 800

Sleeve arrangement for drive shaft length correction is made.


53/60


Page 53

Adapter ansys:

PARAMETER VALUE

Max Equivalent

Stress

193.46 Mpa

Max Shear Stress 104.46 Mpa

Max Deformation 0.13 mm

Factor of Safety 2.74


54/60


Page 54

APPENDIX E

STRUCTURAL OPTIMIZATION OF AUTOMOTIVE

CHASSIS:THEORY, SET UP, DESIGN

Marco Cavazzuti and Luca Splendi

(joint with Luca D'Agostino, Enrico Torricelli, Dario Costi and Andrea

Baldini)

MilleChili Lab, Dipartimento di Ingegneria Meccanica e Civile,

Modena, Italy

Universit_a degli Studi di Modena e Reggio Emilia

[email protected]

ABSTRACT

Improvements in structural components design are often achieved on a trial-

and-error basis guided by the designer know-how. Despite the designer experience

must remain a fundamental aspect in design, such an approach is likely to allow only

marginal product enhancements. A different turn of mind that could boost structural

design is needed and could be given by structural optimization methods linked with

niter elements analyses. These methods are here brief introduced ,and some

applications are presented and discussed with the aim of showing their potential. A

particular focuses given to weight reduction in automotive chassis design applications

following the experience matured at Mille Chili Lab.


55/60


Page 55

1. INTRODUCTION

Optimization techniques are very promising means for systematic design

improvement in mechanics, yet they are not always well known and applied in

industry. Despite this, the literature over the topic is quite rich and is addressing boththeory and applications. To cite a few applications in the automotive _eld the works

of Chiandussi et al. [1], Pedersen [2], and Duddeck [3] are of interest. They address

the optimization of automotive suspensions, crushed structures, and car bodies

respectively .Structural optimization methods are rather peculiar ways of applying

more traditional optimization algorithms to structural problems solved by means of

_nite elements analyses. These techniques are an effective approach through which

large structural optimization problems can be solved rather easily. In particular, with

the term structural optimization methods we refer to: (i) topology optimization,

(ii)topometry optimization, (iii) topography optimization, (iv) size optimization, (v)

shape optimization. In the following some of these techniques will be introduced and

their application to chosen automotive structura ldesign problems discussed.


56/60


Page 56

2. STRUCTURAL OPTIMIZATION

In the de_nition of any optimization problem a few elements are necessary,

these are: (i) design space or space of the possible solutions (e.g. in structural

optimization this is often given by the mesh) (ii) variables, (iii)objective(s) (e.g. mass

minimization), (iv) optimization constraints (e.g. stiffness and/or displacementstargets), (v) the mean through which, for a given set of variables, targets and

objectives are evaluated (e.g.,in our case, _nite elements analyses), (vi) the

optimization algorithm (e.g. in structural optimization this is commonly a gradient-

based algorithm, such as MMA).Trying to simplify in a few words a rather complex

and large topic, it could be said that the various structural optimization methods

essentially differ from each other in the choice of the variables of the optimization

problem as follows.

2.1. Topology Optimization

In topology optimization it is supposed that the elements density can vary

between 0 (void) and 1 (presence of the material). The variables are then given by theelement-wise densities. Topology optimization was _rstly introduced by Bends_e and

Sigmund and is extensively treated in [4]; it has developed in several directions giving

birth to rather different approaches, the

most simple and known of which is the SIMP (Single Isotropic Material with

Penalization).(a) reference model, top view (b) reference model, bottom view (c)

optimum layout Figure 1: Ferrari F458 Italia front hood: reference model and new

layout from the optimization results. The optimization was performed in three stages:

topology, optometry, and size.

(a) reference model, top view (b) reference model, bottom view c) optimum layout

Figure 1: Ferrari F458 Italia front hood: reference model and new layout from the

optimization results. The optimization was performed in three stages: topology,

topometry, and size.

2.2. Topometry Optimization

The idea behind topometry optimization is very similar to that of topology

optimization, the variables being the element-wise thicknesses. Of course, this method

does not apply to 3D elements where the concept of thickness could not be de_ned.


57/60


Page 57

2.3. Topography Optimization

Again topography optimization can be applied only to 2D or shell elements

and aims at _nding the optimum beads pattern in a component. The concept is yet

similar to the previous cases and, simply speaking, the variables are given by the set

of the elements o_sets from the component mid-plane.

2.4. Size OptimizationSize optimization is the same as topometry optimization, but in this case the

number of variables is greatly reduced in that the shell thicknesses of components are

considered in place of the single elements of the domain.

3. APPLICATION EXAMPLES

3.1. Automotive Hood

The internal frame of the Ferrari F458 front hood has been studied aiming at

reducing the weight while keeping the same performance target and manufacturabilityof the reference model. The targets relate to bending and torsion static load cases,

compliance when closing the hood, deformations under aerodynamic loads. A suitable

preliminary architecture has been de-_ned by means of topology optimization. The

results have been re-interpreted into more performing thin-walled cross-sections. A

series of topometry optimizations followed to _nd the optimal thickness distribution

and identify the most critical areas. The solution was re_ned through size

optimization. In the end, the weight was reduced by 12 %, yet in the respect of all the

performance requirements (Fig. 1).

3.2. Rear Bench

The rear bench of a car is fundamental to isolate acoustically the passengers

compartment from the engine. The bench of Ferrari F430 has been analyzed with the

objective of reducing the weight while maintaining the

same vibrational performance of the reference panel. Generally, the damping material

distribution is not known during the numerical veri_cation stage, but is decided later

during the experimental analysis, where the material is added iteratively to counteract

the _rst normal modes. In this study vibration-damping material distribution and panel

design, in terms of beads and thickness, have been optimized through size and

topography optimizations at the same time. Size optimization is applied to control the

thickness of the aluminum plate and of the vibrational-damping material. The

presence of damping material should be limited to essential parts due to its relatively

high weight. Thus, just one thickness variable was created for the aluminum layerbecause its value should be uniform along the plate, whereas several thickness

variables were created locally for the damping layer. Topography optimization was

used to improve the beads disposition in the panel. The objective of the optimizations

was mass minimization, while the _rst normal mode frequency was constrained to be

outside the range of interest (Fig. 2).


58/60


Page 58

3.3. Automotive Chassis

Topology optimization has been applied to the design of an automotive

chassis. The objective of the optimization is still the weight reduction while the

performance requirements regard handling and safety standards, in detail: (i) global

bending and torsional stiffnesss, (ii) crashworthiness in the case of front crash

(a) Size optimization variables (b) Optimum con_guration (c) Damping material optimum thickness

subdivision deformed shape distribution

Figure 2:

Rear bench coupled optimization. In the results, blue stands for low

deformation/thickness, red for high.(a) domain, or design space (b) optimum chassis

con_guration (c) optimum roof con_guration

(a) domain, or design space (b) optimum chassis con_guration (c) optimum roof con_guration

Figure 3:

Automotive chassis topology optimization. In the results, the density range

from 0.1 (blue) to 1.0 (red). (iii) modal analysis, (iv) local sti_ness of the suspension,

engine, and gearbox joints. The initial design space is given by the provisional vehicle

overall dimensions of Ferrari F430 including the roof

(Fig. 3(a)).

The results for the chassis and the roof are shown in

Figs. 3(b) and 3(c).A more detailed discussion on a combined methodology for chassis design

including topology, topography and size optimizations was presented in [5] by the

authors.


59/60


Page 59

4. CONCLUSIONS

A quick overview on structural optimization methods has been given including

various application examples. Their potential has been shown to be large and it is

believed that their spreading in mechanical design could boost innovation in industry

considerably. Examples in the automotive _eld have been provided. To be noted thatthe different methods have different characteristics and in a design process it is

recommended to rely on more than just one technique. For instance, topology and

topometry optimizations are more suitable for an early development stage, whose

outcome could be further re_ned through size and shape optimizations. On a general

basis these techniques do not deliver the shape of the _nal product, but they give

useful hints to the designer in view of the product development and engineering.


60/60


5. REFERENCES

1] G. Chiandussi, I. Gaviglio, and A. Ibba, Topology optimization of an automotive

component without _nal volume constraint speci_cation, Advances in Engineering

Software, 35:609-617, 2004.

[2] C. B. W. Pedersen, Crashworthiness design of transient frame structures using

topology optimization,Computer Methods in Applied Mechanics and En-

gineering, 193:653-678.

3] F. Duddeck, Multidisciplinary optimization of carbodies, Structural and Multi

disciplinary Optimization, 35:375-389, 2008.

[4] M. P. Bends_e and O. Sigmund, Topology optimization: theory, methods and

applications, Springer,2004.

5] M. Cavazzuti, A. Baldini, E. Bertocchi, D. Costi, E.Torricelli, and P. Moruzzi,

High performance automotive chassis design: a topology optimization basedapproach, Structural and Multidisciplinary Optimization, 44:45-56, 2011.

Project Report on ATV prototype BAJA SAE INDIA

Documents

Transcript of Project Report on ATV prototype BAJA SAE INDIA