Paper Number- TTTITM2
Baja 09 design report Raman Sarin Captain, Member design team
Ajay Goyat Steering and brakes department Copyright © 2006 SAE International
ABSTRACT
1. Maximum speed – 45 km/hr
The design report focuses on explaining engineering and design process behind each system in the Baja vehicle that is developed till now. The report also throws some light on the alternatives considered. The design of the vehicle is in accordance with the specifications laid down by the rule book. This design report is a cumulative effort towards explaining the design process to the readers.
2. Weight – 270 kg 3. Ground clearance – 20 cm or 8 inch 4. Track width – 160 cm or 64 inch approx 5. Wheel base – 190 cm or 75 inch approx 6. Braking distance – 1400 cm
INTRODUCTION The design process of the vehicle is iterative and is based on various engineering and reverse engineering processes depending upon the availability, cost and other such factors. So the design process focuses on: Safety, Serviceability, Cost, Standardization, Strength and ruggedness, Driving feel and ergonomics, Aesthetics The design criterion followed here is design for the worst and optimize the design while avoiding over designing, which would help in reducing the cost. We proceeded by setting up the budget for the project. Throughout the design process we distributed the budget in such a way that if we assign more money to one system, we reduce that amount from some other system. Our last year vehicle design was based on the criterion of prevention of failure, as that year no one knew the track and the obstructions prevalent over there. So the procedure of over designing was followed as the safety of the driver is of utmost importance. The design targets of our vehicle for Baja 09 are as follows:
7. Turning radius – 240 cm or 96 inch Further, as designing is based on prevention of failure so let me define the condition of failure of each system of our vehicle. •
For roll cage, failure is yielding as this would change the distance between various parts and thus their working is affected. It should be rigid and the mountings should be able to bear its load.
•
For brakes, failure is their inability to lock all the four tires simultaneously.
•
For tires, it is failure to transmit the required torque maintaining the traction with the track surface.
•
For suspensions, failure occurs if they are not able to isolate the driver from the shocks or if they are so soft that they compress to their solid length while working.
•
For transmission there is less scope of failure but failure is if any part is not able to transmit the required torque or also if torque provided in the first gear is unable to drive the vehicle from a halt.
•
For steering, failure is defined in terms of effort applied by the driver and ability of the various components to facilitate the function of steering.
•
Failure of various other mechanisms like pedals, levers, electrical components occurs if they are not able to fulfill their desired function.
Hence, our designing process targets on the above lying facts to ensure the proper working of our vehicle.
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) Km/hr Where, G=gear ratio N=revolutions per minute R=outer radius of the tire in meters.
MAIN SECTION ENGINE AND TRANSMISSION
Some of our calculations for reverse and forward orientation are as follows:
A quick look at the engine:
Normal orientation Final Gear Ratios
Speed (km/hr)
First
31.45:1
0.65D
14.5
15.8
Second
18.70:1
1.109D
24.4
26.6
Third
11.40:1
1.82D
40
43.6
Forth
7.35:1
2.82D
62
67.7
Reverse
55.08:1
0.38D
10
9
Power - 8 kW at 4400 rpm Max Torque – 19 Nm at 3000 rpm About gear box, we have 4 forward and 1 reverse gear box with built in differential and universal joint. As engine and gear box were given to us. Thus we had a little choice while working on transmission. Configuration of our vehicle would be rear engine rear wheel drive. We decided to keep the maximum speed of the vehicle at 45 km/hr as the vehicle is not about larger speed but greater torque and stability. For attaining this speed, the only thing we can vary was the outer diameter of the driving tire. For 45 km/hr O.D. of the tire came out to be 16 inch. This diameter is too small as ground clearance decreases.
Speed (km/hr) D=22 inch
D=24 inch
Reverse orientation Hence in order to counter this problem options available were:
Final Gear Ratios
Speed (km/hr )
First
55.08:1
0.38D
8.3
9
10.5
Second
32.75:1
0.63D
13.9
15.2
17.7
Third
19.96:1
1.04D
22.8
24.9
29
Forth
12.87:1
1.61D
35.45
38.7
45
Reverse
31.45
0.65D
14.5
15.8
18.5
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. We decided to work on the latter option and so did reverse engineering process trying to find if the gears would be able to transmit the increased torque. Also following this method, 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.
Speed (km/hr) D=22 inch
D=24 inch
D=28 inch
Hence for maximum speed of 45 km/hr, we selected tires of 28 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 (from engine characteristics graph; fig e1). Thus, for better economy, the range of speed in each gear, for the driving tires of O.D. 28 inches; operating in reverse orientation is: First Second Third Forth Reverse
- 6.7 to 9 km/hr - 11 to 14 km/hr - 18 to 24 km/hr - 29 to 37 km/hr - 12 to 15 km/hr
Apart from this, for mounting the engine we are going to use neoprene rubber mountings.
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.
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 of tire ÷ gear ratio) For the reverse orientation of the transmission system and maximum speed of the vehicle as 45 km/hr radius comes out to be 28 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.
ADVANTAGES: 1. Built with a 6 ply rating and a reinforced casing makes these one of the most puncture resistant tires in the market today.
LAST YEAR: Front and rear same tires Outer diameter of tire – 24 inch Outer diameter of rim – 12 inch Tread width – 6 inch Aspect ratio - 1 Number of plies – 6 Tread design – mud cutting Side with – 210 mm
THIS YEAR: FRONT
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.
Outer diameter of tire – 24 inch Outer diameter of rim – 12 inch Tread width – 8 inch Aspect ratio - 1 Number of plies – 6 Side with – 198 mm
6. Use of the larger outer diameter tire at the rear helps to provide good ground clearance and also 10 inch treads provides good traction to the power wheels.
REAR
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.
Outer diameter of tire – 28 inch Outer diameter of rim – 12 inch Tread width – 10 inch Aspect ratio - 1 Number of plies – 6 Side with – 231 mm Shown in fig t1
BRAKES
LAST YEAR‘S BRAKING SYSTEM: Front Rear
THIS YEAR:
Disc brake of Maruti800 (91 mm) Drum brake of APE (180×30mm)
In the last years vehicle we found that the braking force was not enough to lock all the four wheels simultaneously 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 ÷ g) ×W× (H ÷ L) Rear axle dynamic load = w2 – (α ÷ g) ×W× (H ÷ 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 front and drum brakes in rear. Initially we thought of using disc brakes for all four wheels but disc with parking brakes have higher cost and we found it necessary to use the parking brakes to increase the all terrain capabilities of the vehicle. Some formulas that we used for designing our brakes: T (disc) + T (drum) = m × f ×R T (disc) = µ×R1× (P × A) ×2 T (drum) = (P × A) × Brake factor×R2 Where, T(disc) = Frictional torque on the disc T(drum) = Frictional torque on the drum f = deceleration m = mass of vehicle R = radius of tires P = Pressure applied by the TMC. µ= Coefficient of friction R1=Radius of the disc A= Area of the caliper for disc brake and wheel cylinder for the drum brake.
Using these formulae, we have done our calculation and selected our brakes. Some of calculations are shown in the table: F kg
Pr
D1
D2
D3
R1
R2
R3
R4
16.2 5 22.9 6 17.7 8 19.8 8 20.6 4
50
30
91
90
12”
14”
50
30
91
90
12”
14”
50
30
91
90
12”
14”
50
30
91
90
12”
14”
50
30
91
90
12 ”
14”
1
20
2.5
2
20
5
3
20
3
4
15
5
5
26.9 3
3
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) D3=Diameter of the wheel cylinder for the drum (mm) R1= Radius of the disc (mm) R2=Radius of the drum (mm) R3=Outer radius of the front tires (inch) R4=Outer radius of the rear tires (inch) The above highlighted specifications have been selected for our vehicle. These are the standard specifications of Maruti Zen’s braking system with vacuum booster. We selected these as per our design of the braking system for 5.9 m/s^2 deceleration. The pedal force would decrease further by a factor 3 due to the use of booster. So the force on pedal would be 9 kg approx.
COMPARISON WITH PREVIOUS YEAR BRAKES: 1. The friction material is semi-metallic which has got better frictional properties. So we have a higher coefficient of friction 2. Vacuum booster is used for giving the better comfort in applying the paddle force. 3. Use of X-type brake fluid lining which will give us better response and has a higher reliability. 4. Ventilated disc for higher heat dissipation rather than a single disc.
ADVANTAGES 1. Standardization of parts is there. Thus reliability is there. 2. The cost of standard parts is lower.
3. Ergonomically, the use of booster would get the pedal force to a lower value. Thus facilitating the driver. 4. Friction material is in the disc and drum is of semi-metallic which has very good frictional property.
STEERING SYSTEM LAST YEAR: • • • • • • • • • •
Reticulating ball type, pitman arm 4 tie rods used Steering ratio 19.53:1 Turning radius - 9 feet IBJ to IBJ - 50 mm OBJ to OBJ - 1090 mm Column inclination from horizontal - 40 degree One flexible coupling is used in column One universal joint was used. Two pivots were used as shown in fig above
THIS YEAR: • • • •
Reticulating ball type (ZF steering) Turning radius – 9 feet Gear ratio - 19:1 Steering ratio – 17:1
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.
response. The only problem lying with it is its higher weight than rack and pinion type steering.
ADVANTAGES 1. It has an advantage that being worm type, only driver effort would be transmitted to the wheels. But unlike rack and pinion, the wheels reaction generated from the track would not be transferred to the driver. 2. Reticulating ball type steering has lower wear and tear as compared to rack and pinion steering. 3. It can withstand 25000 cycles at the constant load of 250N/m, at the pitman arm speed of 20 to 25 RPM. Further, apart from keeping the steering ratio to be 17:1. Our main concern in the design of the steering mechanism, using reticulating type steering, is to reduce the weight of the mechanism and to incorporate as minimum joints as possible which would help to reduce the human force required to steer the vehicle.
SUSPENSIONS Suspensions act to provide cushioning action to the driver by absorbing the shocks from the road and also help the tires to maintain good traction. Last year Front - Mac Pherson strut Rear – Mac Pherson strut and rubber bumpers Last year, the problem with out suspensions was that they were too stiff. So the movement of springs was too small ride was not comfortable.
The formulae used for steering calculations are: This Year – C^2 = X^2 + Y^2 Unequal wishbone suspension in both front and rear X = c sin (p) + (a+ b sin (q) – a cos (q)) Reason: Y = b cos (q) + a sin (q) - R
•
Where, C – length of tie rod X, Y – lengths as shown in fig s1 p, q – angles as shown in fig s1 a – length of steering knuckle from center of tire b – perpendicular distance of steering knuckle from pivot point as shown in fig s1
Wishbone suspension give more movement of the tires and hence the vehicle for the same movement of the spring.
•
Independent suspension.
•
In double wishbone suspension, force is distributed at 5 points on the roll cage unlike at only one point in Mac Pherson strut.
We stick to reticulating ball type steering as we had a good experience using it in last year vehicle in terms of
•
It can be slightly adjusted for different parameters of suspension tuning like camber angle, ground
clearance at the time of testing and then finalized (proper adjustments are made at the time of fabrication). Design of suspension system should be such that it is able to sustain the worst of the conditions.
year’s vehicle Estimated weight this year’s vehicle
of
270 kg approx.
Driver with accessories
90 kg approx.
For example, in the case when the vehicle is falling on ground after jumping from a speed breaker, just the two wheels support the vehicle as it lands on the ground. But if we design our springs according to this situation, our spring will be a lot stiffer and hence the ride will not be comfortable. Also if we choose stiffer springs, they would not be able to facilitate tire traction. On the other hand softer spring mean that a larger spring travel should be more otherwise they would reach to their solid length. Hence the suspension system would fail.
Total weight with driver
360 kg approx.
Unsprung mass
75 kg approx.
Sprung mass
300 kg (at max. with driver)
This criterion can be fulfilled by the 2 alternatives:
Since the major components of the sprung mass(in terms of weight) like engine, transmission, driver etc…. are at the back only, the weight distribution is taken as 50 kg on each suspension in the front and 100 kg on each suspension at the back. Also, this was the approximate ratio of distribution of the vehicle weight of last year vehicles.
1. By putting a spring of gradually changing pitch and hence stiffness. This is the best method to encounter this problem but we could not find a vendor who could manufacture for us a continuously varying spring.
The spring design is to be for the total weight of around 300 kg now.
2. By putting a very long soft spring which has enough uncompressed length left so that it would remain in its working range without reaching its failure limit. This method was used by one of the team last year. But the main problem is that the spring might buckle. Even with a damper, the springdamper system might buckle. We might be able to solve the problem using guides but this is making the system unnecessarily complex.
FRONT SUSPENSIONS
3. By putting a system of compound spring (in parallel) in which only one spring is acting in normal conditions and a stiffer spring starts to work only after reaching a certain amount of load. This is the method that we will follow because:
Length of spring = 300 mm
• •
It can be said to be equivalent to the first system. The range of travel is small as compared to the previous two methods and hence our damper buckling problem is also solved to an extent.
Spring Design started with some arbitrary parameters within the constraints Constraints: Weight, ground clearance required and space limitations Weight
of
the
last
472
The spring damper would be placed at the centre of the lower wishbone. Taking ground clearance to be around 8 inches and load of 50 kg on each tire. Thus static load on each spring would be 100 kg as spring is mounted at the centre of the wishbone
Total length (spring + damper) = 430 mm Wire diameter (d) = 9mm Mean coil diameter (D) = 70 mm Allowed travel of the spring = 160 mm Maximum travel of the spring = 192 mm Spring stiffness (K) = 20 N/mm Pitch = 25 mm No of active turns = 10 Total no of turns = 12 Springs are squared and grounded
Initial compression (after driver is seated) = 50mm Thus springs can take up a load of 220 kg apart from the weight of vehicle and driver The dimensions and assembly of front suspensions is as shown in fig sus1
Further, if one of the rear tire falls in a ditch, their will be load on the spring. Assuming tire and brake assembly weight to be 20 kg, deflection of spring required is 6.7 mm or in terms of shaft movement we can say that 1.5 degree of shaft movement would be sufficient for allowing the movement of tire if it encounters a ditch. Hence under static but loaded conditions, position of shaft below the horizontal level is 4.5 degree (12-7.5 degree).
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 fig sus2; 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 fig sus3. 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
Now, the allowed movement of shaft under dynamic conditions is 7.5 degree or we can allow spring movement of 39 mm. Thus, the rear suspension can accommodate an additional load of 117 kg.
ROLL CAGE AND MATERIALS 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. It should be strong enough to bear the laden load and should be designed against impact load that it might encounter. The failure criterion for the roll cage is yielding. Our design of the roll cage started with ergonomic and driver comfort study. We also studied the rules and safety instructions as per Baja SAE INDIA 2009 rulebook. This was followed by study of compatibility of various other systems with the roll cage, as these systems were developed in the process. Based on these, we designed a layout which was modified again and again to take the present shape as shown in fig r1. Adjacent to fig r1 we also have the roll cage of last year vehicle as fig r2. The software used by us is Pro-E for 3D modeling and design and Auto-CAD for 2-D drafting.
Now, Length of spring = 230 mm Total length (spring + damper) = 490 mm Wire diameter (d) =11.1 mm Mean coil diameter (D) = 80 mm Allowed travel of the spring = 72 mm
Pitch = 19 mm
Initially I assumed the ratio of total height of driver to length of driver below waist as 1.65 (considering myself as standard) and designed a roll cage model for a person of height 6 feet 3 inches. Then slowly as the other systems of the vehicle were developed, the roll cage design got modified.
No of active turns = 10
Dimensions of the roll cage are see (fig r3):
Total no of turns = 12
Length – 2300 mm
Springs are squared and grounded
Width – Max At front end At rear end
Maximum travel of the spring = 96.8 mm Spring stiffness (K) = 30 N/mm
Initial compression (after driver is seated) = 33.3mm From initial compression we conclude movement of shaft required is 6.3 degrees
that
the
Height – 1440 mm
- 870 mm - 540 mm - 720 mm
The FEM analysis of the roll cage is still pending and would be included in the final design report.
Odometer, speedometer, fuel indicator, oil pressure lamp, brake switch, brake lights, reverse alarm
The material that we are going to use is mild steel, IS: 1239 (part 1):2004. The material has chemical composition as:
We are going to use a separate reversing lever
CARBON
MANGANES E
SULPHUR
PHOSPHORUS
0.2
1.30
0.040
0.040
The pipe we are using is of electric resistance welded type, heavy duty pipe with the following specifications: Bore – 20 mm
We also worked on a gear shifting mechanism which would be available near the steering wheel
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.
Wall thickness – 3.2 mm Outer diameter - 26.5 to 27.3 mm Weight per meter – 1.87 kg
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.
Yield strength – 480 N/mm^2 (as per UTM test) As the yield strength is as per UTM test so we assume working yield strength of 400 N/mm^2 The pipe of above specification has a higher bending strength and rigidity than the material specified by the rule book. For safety of the driver, Ethan foam padding would be used over the pipes in the adjacent of the driver. For fabrication of the roll cage, we are going to use metal inert gas welding and cold bending techniques.
OTHER MECHANISMS This section includes all the levers, electrical equipments etc that form an important part of our vehicle. Apart from the accessories provided by Lombardini, we are going to use
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.
CONTACT
Battery: 12 V, 44 Ah
Raman Sarin Mechanical Engineering student Institute of Technology and Management, Gurgaon
Kill switches: 2
Email I.D. –
[email protected] Address - #1178, Sector 18-C, Chandigarh
Fig e1
REAR and FRONT Fig t1
Steering mechanism (Tie rods to steering knuckle) Fig s1
Fig sus1
Fig sus2
Fig sus3
THIS YEAR ROLL CAGE
LAST YEAR ROLL CAGE
Fig r1
Fig r2
Fig r3
