Institute Of Technology And Management

Gurgoan, Haryana

SAE Baja Asia 2010

The Techie Tyros

Design Report 2010

DESIGNED BY

Vikrant Dalal Vice-captain and Head of Design Team

#TTT-3

SAE BAJA ASIA 2010 Design Report

Vikrant Dalal Vice-captain, Head of Design Team

Copyright © 2009 SAE International

1.0 ABSTRACT

The objective of the Baja competition is to design a Baja all-terrain vehicle that embodies innovation, simplicity and functionality, delivering high performance and safety at a reasonable price. This report details the considerations, functions and processes behind the separate vehicle subassembly

2.0 INTRODUCTION

Baja-2010 is an international competition sponsored by the Society of Automotive Engineers (SAE). Engineering students are given a challenge to design, simulate and manufacture a “fun to drive”, versatile, safe, durable, and high performance off road vehicle.

The Techie Tyros 2010 Baja team consists of 22 undergraduate students in Automotive and Mechanical Engineering. The goal of the 2010 car is to improve on some of the key areas that have caused the team problems over the last few competition years. These areas are: suspension, steering, driveline, hub and fabrication tolerances. It was decided that, while making components lightweight is important, strength and durability of key components would not be sacrificed for weight reduction. All subassemblies and components were researched and designed to meet pre-established team expectations.

For designing, simulation, analysis and optimization of the vehicle components various software such as Pro-E (design and analysis), Cosmos (analysis and simulation), Optimum K and suspension analyzer (Suspension design and analysis), ADAMS and IPG car maker (vehicle dynamics) are used.

The design targets of our vehicle for Baja 2010 are as follows:

1. Maximum speed – 60 km/hr

2. Weight of vehicle – 270 kg

3. Ground clearance – 20 cm or 8 inch

4. Track width – 140 cm or 56 inch approx

5. Wheel base – 130 cm or 52 inch approx

6. Braking distance – 1400 cm

7. Turning radius – 240 cm or 96 inch

3.0 MAIN SECTION

The design of Baja 2010 is divided into follow section:

1. CHASIS AND ANALYSIS

2. ENGINE AND TRANSMISSION

3. TYRES

4. BRAKES

5. STEERING

6. SUSPENSION

7. HUB DESIGN

3.1 CHASIS AND ANALYSIS

The kind of body we are required to manufacture is a unitized body. The roll cage is of utmost importance for us as it would be the one which would provide safety to the driver, mounting points for various systems and even ergonomics and looks to the vehicle.

3.1.1 MATERIAL

The material used in vehicle must fulfill the SAE Baja requirements. The proper equilibrium should be acquired between the design requirements, cost and weight to achieve an unbeaten design. The available materials that fulfil the requirements are AISI 1018, 1020, IS1239 part-1 and 4130.A comparison was done to select the material by considering various properties and cost of each material.

Table 1: Material property

MATERIAL

YIELD STRENGTH

(KSI)

MODULUS OF

ELASTICITY (KSI)

COST PER

METER (RS)

ELONGATION AT YIELD POINT

(%)

AISI 1018

53.7

29700

600

15

AISI 1020

42.7

29700

2200

36

IS 1239

59.12

29700

765

18.5

AISI 4130

65.1

29700

2500

25.5

Graph 1: Weight and Cost comparison

AISI 4130 and IS 1239 part-1 having good yield strength will allow the usage of tubing with smaller wall thickness. This will in turn reduce the weight of our chassis. Also 4130 and IS 1239 part-1 are more ductile than other materials so it will deform more before its ultimate failure. But cost of per meter length of 4130 is 2.5 times more expensive IS 1239 part-1. So considering the above said factors we have chosen IS 1239 part-1 pipes to be used for our chassis.

3.1.2 FEA ANALYSIS

The initial design is shown in the Figure 1. Some notable features are the fact that the design consists of 4 main members: the roll hoop, the horizontal hoop, and the two perimeter hoops. As mentioned above the design was made using the Pro-E solid modelling package.

Figure 1: Roll cage model in Pro-E

In order to optimized the strength, durability and weight of Chassis cosmos was used to analyze the chassis for all six loading condition. The six analysis tests conditions are front impact, side impact, rollover impact, heave and the loading on the frame from the front and rear shocks

After running all five analyses it was found that there is a need of additional member. After having added these members, a second analysis using identical loading constraints was completed and results of these tests are shown in table 2.

For front collision test stress diagram and displacement diagram is shown in figure 2 and 3.

Table 2: FEA Analysis results

Type of case Force applied Result

Front Collision 4500 KN Passed

Side Collision 1200 N Passed

Rollover 1800 N Passed

Front Bump 810 N Passed

Figure 2: Stress analysis for front collision

Figure 3: Stress displacement for front collision

7677787980818283

0

500

1000

1500

2000

2500

3000

AISI 1080AISI 1020 IS 1234 AISI 4130

Wig

ht

(KG

)

Co

st (

Rs)

cost(*10) weight

Comparison between previous years roll cage design is done and the results are shown in table3.

Table 3: Comparison of pervious 3 years design

2007-BAJA ROLLCAGE

DESIGN

2008-BAJA ROLLCAGE

DESIGN

2010-BAJA ROLLCAGE

DESIGN

SAFETY

Poor

High

High

COMFORT

Average

High

High

ERGONOMIC

Poor

Good

Very good

SPECE FOR

ELECTRONIC DEVICES

More

Less

Sufficient

WIEGHT To heavy

Medium

Light

COST OF ROLLCAGE

High

Low to high

Low

STANDARDIZATION

9/10

4/10

8/10

3.2 ENGINE AND TRANSMISSION

A quick look at the engine:

Power - 8 kW at 4400 rpm

Max Torque – 19 Nm at 3000 rpm

Engine was given to us. Thus we had a little choice while working on engine. Configuration of our vehicle would be rear engine rear wheel drive. We decided to keep the maximum speed of the vehicle at 60 km/hr as the vehicle is not about larger speed but greater torque and stability.

As per the rules of the competition, the engine cannot be modified in any way. This restriction causes the design emphasis to be placed on the choice of transmission. For the transmission we have several options:

A manual transmission (4 or 5 speed): this system would allow the driver to select the right gear from the available gears allowing more control over the vehicle. This is seen on most manual cars with a standard “H” pattern. A sequential transmission: this is similar to the manual transmission, but the “H” pattern is eliminated and replaced with a different shifting pattern. For example in a race car, the motion of the shift lever is either “push forward” to up-shift or “pull backward” to downshift. These transmissions are usually found in either motorcycles or all terrain vehicles. This type is most suitable of our vehicle as it has good sensitive i.e. why

we are using M&M champion transmission. To increase

the torque following options were available: 1. Manipulation of power transmission outside the gear

box using gears, sprockets and chain.

2. Engaging the reverse gear lever while driving in all the forward gears and using the first gear in forward as reverse gear.

3. Using the transmission system in normal conditions.

We decided to work on the 3rd option due to following reason:

1. We were able to check the weight

2. Reduce the cost of the vehicle as we avoided the use of additional gears, sprockets and chains.

3. We used standard parts, thus increased the reliability of the transmission system.

To find the speed of the vehicle corresponding to different gear ratios, the formulae used is

Velocity on road = 2π×N×R×60

1000 ×𝐺

Where, G=gear ratio N=revolutions per minute R=outer radius of the tire in meters.

Some of our calculations for normal orientation are as follows:

Table 4: Normal orientation

Final Gear

Ratios

Speed (km/hr)

Speed (km/hr)

D=22 D=24 inch inch

First 31.45:1 0.65D 14 16

Second 18.70:1 1.109D 24 27

Third 11.40:1 1.82D 40 44

Forth 7.35:1 2.82D 60 68

Reverse 55.08:1 0.38D 10 9

Hence for maximum speed of 60 km/hr, we selected tires of 22 inch outer diameter.

Further, for better economy, we assume engine rpm to be ranging from 2750 to 3250 as maximum torque produced by the engine is at 3000 rpm. In between this range the torque produced by the engine is almost

constant. Thus, for better economy, the range of speed in each gear, for the driving tires of O.D. 22 inches; operating in normal forward orientation is:

First - 10 to 12 km/hr Second - 15 to 18 km/hr Third - 25 to 33 km/hr Forth - 40 to 51 km/hr Reverse - 8 to 11 km/hr

Apart from this, for mounting the engine we are going to use neoprene rubber mountings.

3.3 TIRES

Selecting the tires is one of the most important things as the whole vehicle is in contact with the road on these 4 points or rather patches. Also for designing an all terrain vehicle tires form the most important part. They should be such that they are able to provide enough traction on all kind of surfaces so as to transmit the torque available at the wheels without causing slipping.

Front

Outer diameter of tire – 21 inch Outer diameter of rim – 10 inch Tread width – 6 inch Aspect ratio – 0.68 Number of plies – 6

Rear

Outer diameter of tire – 22 inch Outer diameter of rim – 10 inch Tread width – 8 inch Aspect ratio – 0.75 Number of plies – 6 One of the most important parameter for the selection of the outer diameter of the tires in rear was the maximum speed of the vehicle. The relation between outer diameter of the tires and the vehicle speed is as given below:

Velocity on road = 𝑨𝒏𝒈𝒖𝒍𝒂𝒓 𝒗𝒆𝒍𝒐𝒄𝒊𝒕𝒚 ×𝑶𝒖𝒕𝒆𝒓 𝑹𝒂𝒅𝒊𝒖𝒔

𝒈𝒆𝒂𝒓 𝒓𝒂𝒕𝒊𝒐

For the normal orientation of the transmission system and maximum speed of the vehicle as 60 km/hr radius comes out to be 11 inches. Apart from outer radius of the tire, other factors for the selection of tires include tread width, tread design, side wall width, load handling capacity, number of plies and treads on side wall etc which define the traction ability, tire resistance to wear and puncture and performance of the tire on various terrains.

Reason:

1. Built with a 6 ply rating and a reinforced casing makes these one of the most puncture resistant tires in the market today.

2. Large shoulder knobs wrap down the sidewall to provide excellent side to pull out of the ruts without causing sidewall failure.

3. The deep tread and open wing design provides excellent clean-out with each lug and an improved

traction.

4. Special natural compound delivers added traction.

5. Smaller tires in front results in a smaller magnitude of moment on the wishbones due to cornering forces during steering.

6. Use of the larger outer diameter tire at the rear helps to provide good ground clearance and also 8 inch treads provides good traction to the power wheels.

3.4 BRAKES

The criterion for designing the brakes stated as per the rule book is that all the four wheels should lock simultaneously as the brake pedal is pressed.

For designing the braking system this year, we calculated the weight of our vehicle in static condition as well as in dynamic condition as per the deceleration (0.6 g) and stopping distance. In static condition it is around 60kg on each front tire and 110kg each on the rear tire. But in dynamic conditions, we consider weight to be 85kg on each tire, the front and the rear. We have calculated the dynamic weight using the formulae as given below:

Front axle dynamic load = 𝑤1 +𝛼×𝑊×𝐻

𝑔×𝐿

Rear axle dynamic load = 𝑤1 −𝛼×𝑊×𝐻

𝑔×𝐿

Where, W1=Weight on the front axle in the static condition. W2=Weight on the rear axle in the static condition.

g = Acceleration due to gravity. W= Total weight of the vehicle. H=Height of center of the gravity. L= Length of the wheel base. Deceleration of the vehicle is α. We planned to use disc brake in all four wheels. Initially we thought of using disc brake in front and drum brakes in rear but problem with drum brake is of locking .For achieving the condition for locking at once on the application of brake paddle, it is preferred to use disc in all four wheels. Some formulas that we used for designing our brakes:

T (disc) = 𝑊1 ×𝑓

𝑔× 𝑅1 + 𝑊2 ×

𝑓

𝑔× 𝑅2

T (disc) = 𝜇 × 𝑅 × 𝑃 × 𝐴 × 2 × 𝑛𝑜. 𝑜𝑓 𝑑𝑖𝑠𝑐 𝑝𝑎𝑑

Where, T (disc) = Frictional torque on the disc f = deceleration W = weight of the body R = Effective radius of disc R1= Radius of front tire R2= Radius of rear tire P = Pressure applied by the TMC. µ= Coefficient of friction R=Radius of the disc A= Area of the caliper for disc brake P= Pressure applied by the master cylinder. Using these formulae, we have done our calculation and selected our brakes. Some of calculations are shown in the table 5: Table 5: Brake pedal force calculation

F

kg

Pr

D1

mm

D2

mm

R

inch

R1

inch

R2

inch

3.0 3.21 16.25 16 98 10.5 11

2.5 3.86 16.25 16 98 10.5 11

3.0 3.84 17.78 16 98 10.5 11

3.8 3 17.78 16 98 10.5 11

3.2 3.58 17.78 16 98 10.5 11

3.0 4.44 19.05 16 98 10.5 11

3.2 3 16.25 16 98 10.5 11

Where the parameters shown above are as under: F=Pedal force required for braking (kg) Pr = Pedal ratio D1=Diameter of the TMC (mm) D2=Diameter of caliper cylinder for the disc (mm) R = Effective radius of the disc R1=Outer radius of the front tires (inch) R2=Outer radius of the rear tires (inch)

The above highlighted specifications have been selected for our vehicle. We selected these as per our design of the braking system for 5.9 m/s^2 deceleration.

3.5 STEERING SYSTEM

After a comparative study on different steering which are available in the market it was found that the best suitable steering for our vehicle is central roller and rack. Table 6 shows results of our study on steering. Table 6: Steering comparison

Types of steering

cost weight Sensitivity and

response

efficiency

Rack and pinion

low light poor fine

Central roller and

rack

low light good good

Recirculating ball type

high medium poor Very good

Worm and roller

medium heavy poor medium

Worm and sector

medium heavy Very poor good

• Central roller and rack. • Turning radius – 8 feet. • No. of teeth on the Rack bar =36 • Length of rack = 144mm • The ratio of the rack and pinion = 12:1 • The axial pitch of the Rack bar = 6 mm • Steering ratio –7.8:1 • No. of universal joints in column = 2 • Column inclination from horizontal- 45 degree • Removable steering wheel assembly for the ease of driver exit in time specified as per the rulebook. • No. of the tie rods = 2. Figure 4: Central roller and rack

While designing the steering system the constraints that we possessed were centre alignment of steering system, track width, human effort at the steering wheel and the desired response of the steering system.

Apart from deciding the steering ratio we have not been able to design the linkages, tie rods etc as presently we do not have the gear box of steering.

The formulae used for steering calculations are:

𝑪𝟐 = 𝑿𝟐 + 𝒀𝟐

𝑿 = 𝑪𝒔𝒊𝒏 𝒑 + 𝒂 + 𝒃 𝒔𝒊𝒏𝒒 − 𝒂𝒄𝒐𝒔 𝒒

𝒀 = 𝒃 𝒄𝒐𝒔 𝒒 + 𝒂 𝒔𝒊𝒏 𝒒 − 𝑹

Where, C – Length of tie rod X, Y – lengths as shown in fig 5 p, q – angles as shown in fig 5 a – length of steering knuckle from center of tire b – Perpendicular distance of steering knuckle from pivot point as shown in fig 5.

FIGURE 5: Steering knuckle

3.6 SUSPENSIONS

Suspension is the term given to the system of springs, shock absorbers and linkages that connects a vehicle to its wheels. The suspension systems not only help in the proper functioning of the car's handling and braking, but also keep vehicle occupants comfortable and make your drive smooth and pleasant. It also protects the vehicle from wear and tear.

This year we are going to use equal wishbone suspension in both front and rear because of the following reasons:-

* Double wishbone designs allow the us to carefully control the motion of the wheel throughout suspension travel, controlling such parameters as camber angle, caster angle, toe pattern, scrub radius more.

* In a double wishbone suspension it is fairly easy to work out the effect of moving each joint, so you can tune the kinematics of the suspension easily and optimize wheel motion.

*Double wishbones are usually considered to have superior dynamic characteristics, load handling capability and are still found on higher performance vehicles.

Spring Design started with some arbitrary parameters within the constraints

Constraints: Weight, ground clearance required and space limitations

Estimated weight of vehicle

250 kg approx.

Driver with accessories

90 kg approx.

Total weight with driver

340 kg approx.

Unsprung mass 75 kg approx.

Sprung mass 265 kg (at max. with driver)

Now according to design for rear wheel drive 40% of the total weight will be distributed at the front portion and the remaining 60% of the weight will be at the back or rear end.

From the above estimated weight we find that weight distribution at one side of front end will be approximately 70 kg and at one side of rear end will be approximately 105 kg. So, all the calculations will be done taking this weight distribution only.

3.6.1 FRONT SUSPENSIONS

The spring damper would be placed at the centre of the upper wishbone as shown in the figure 5.

Taking ground clearance to be around 8 inches and load of 70 kg on each tire. Thus static load on each spring would be 140 kg as spring is mounted at the centre of the wishbone

Figure 6: Front suspension on optimum k

Front spring design specification of our vehicle is shown in the table 7.

Table 7: Front suspension spring details

Length of spring 171 mm

Total length(spring + damper)

291mm

Wire diameter 7mm

Mean coil diameter 51mm

Allowed travel of spring 100mm

Stiffness 20N/mm

Pitch 19mm

No. of active turns 10

Total no. of turns 12

3.6.2 REAR SUSPENSION

Here also the constraints were ground clearance 8 inches, vehicle weight 110 kg on each tire and movement of transmission shaft as shown in figure 7; full angle being 15 degree, full jounce 3 degree and full rebound 12 degree

In here, we keep the mounting point of the spring on the upper wishbone and at its end. The rear suspension system is as shown in figure 7.

Figure 7: Rear suspension on optimum k

For the smaller half drive shaft, the distance between spring mounting point and shaft hinge point is 12 inch approximately. Thus, for 15 degree spring movement is 80 mm as calculated by the formulae:

LENGTH OF ARC = RADIUS * ANGLE SUBTENDED

So for 1 degree movement of shaft deflection of spring is 5.3 mm

Rear spring specification after designing the rear suspension is shown in table 8.

Table 8: Rear suspension spring details

Length of spring 230 mm

Total length(spring + damper)

490mm

Wire diameter 11mm

Mean coil diameter 80mm

Allowed travel of spring 72mm

Stiffness 30N/mm

Pitch 19mm

No. of active turns 10

Total no. of turns 12

Initial compression (after driver is seated) = 33.3mm

From initial compression we conclude that the movement of shaft required is 6.3 degrees

3.6.3 DESIGN AND ANALYSIS OF WISH BONES

FRONT SUSPENSION

REAR SUSPENSION

3.7 HUB DESIGN

The hub assembly has a very important contribution towards vehicle’s weight. So to achieve our main objective of reducing the overall weight of our vehicle we have to reduce the weight of wheel assembly. We have a detailed study of previous year’s wheel assembly. It was made up of mild steel that why weight of the assembly is extremely heavy. This year we have decided to use aluminium alloy for the manufacture of our hub .we are also using some standard part such as disc, spline cut alto shaft etc to reduce the cost of our hub assembly. Wight of this year hub assembly is about 3kgs and 400gms which is 4 times less than previous year

TABLE 9: Hub weight comparison

Hub assembly weight of 2010

3kg and 440gms

Hub assembly weight of 2009

14kgs and 780gms

Hub assembly weight of 2007

22kgs and 340gms

FIGURE 8: Hub design on pro-e

4.0 CONCLUSION

As discussed earlier, our approach is to design for the worst and still optimize so that we avoid over designing. This would help us to reduce the cost.

The approach that we followed is iterative in nature and processes like reverse engineering are adopted in order to select various systems from the ones, existing in the market. This step would ensure standardization and reliability would follow as a by part.

Our top priority would always be the safety of the driver and working in this direction, we will strive to add aesthetic value and a sense of ergonomics to the vehicle.

5.0 ACKNOWLEDGMENTS

The design process is not a single handed effort and so it is my team, whom I wanted to thank for standing with me under all circumstances. I would also like to express my gratitude towards our Mechanical department and on the whole towards the college for supporting us and believing in us. SAE has provided us with an excellent platform for learning and showcasing real life projects. While working on the project, it was really heartening to see that the people from industry were willing to help us and they provided us with their precious time.

6.0 REFERENCES

1. S.S.Rattan ,2005,”Theory of Machines”

2. V.B. Bhandari ,2007,”Design of Machine Elements”

3. SAE , 2008 ,Advanced Vehicle Technology ”

4. Thomas D. Gillespie ,2008 ,”Fundamentals Of Vehicle Dynamics”

7.0 CONTACT

Vikrant Dalal Mechanical Engineering student Institute of Technology and Management, Gurgaon Web site – www.thetechietyros.com Email I.D. – vikrantdalal@yahoo.co.in Address: V.P.O Goela Khurd, Najafgarh New Delhi 110071

engine

Type 4-stroke, gasoline Lombardini engine

Displacement 305 cc

Compression Ratio 8:1

Power 8 KW

Torque 19 NM at 3000 rpm

Drive Train

Transmission 4 speed manual constant mesh gear box with 1 reverse

Company Mahindra alpha champion

Chassis/Suspension

Chassis Type IS 1239 Steel Pipes

Overall Length 1400 mm.

Wheel Base 1150 mm.

Overall Width 1600 mm.

Front Suspension Double Wishbone

Rear Suspension Double Wishbone

Ground Clearance 250 mm

Shocks coil-over

Front Travel 200 mm. (75 mm rebound and 125 mm jounce )

Rear Travel 100 mm (75 mm rebound and 25 mm jounce )

Vehicle Weight 270 kg

Wheels/Tires

Front Tires 21 in. x 6 in. ITP Holeshots

Front Wheels 10 in.

Rear Tires 22 in. x 8 in. ITP Holeshots

Rear Wheels 10 in.

Performance

Approach Angle 80 degrees

Departure Angle 60 degrees

Top Speed 60 km/hr

Rear Wheel Torque 1584 NM