Paper Number 2
Material chosen for the frame is ASTM A106 schedule 40 steel with a radius of 1” and a wall thickness of 3mm. 3m m. Som ome e braci racing ng me memb mber ers s have ave 2m 2mm m wa wall ll thickness. In accordance with section 31.5 sub-section A of the rule book, the material chosen has a carbon content of 0.265% which is >0.18%.
FINAL DESIGN REPORT TEAM STRATOS, BIT Team #8, Car #8
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Copyright © 2009 SAE International
The earlier frame design is shown below. The forces used in its analysis were too low in magnitude. New values of impact and torsional forces were calculated.
ABSTRACT The aim of this report is to highlight the final design of Team Stratos’ mini-Baja vehicle which will compete in Baja SAEASIA 2010.
The team’s primary objective was to design a safe and functional vehicle based on a rigid and torsion-free roll cage cage and and chass chassis, is, we well ll mou mount nted ed powe powertr rtrai ain, n, and and dynamically tested steering and suspension systems. The secondary objective was to enhance performance and maneuverability of the vehicle. The team was divided into core groups responsible for the design design and opt optimi imizat zation ion of major major sub-sy sub-syste stems ms which which were were later later integra integrated ted into into the final final bluepri blueprint. nt. Current Current CAD modeling modeling and FEA approac approaches hes were were used.
INTRODUCTION
fig i
We approached our design by considering all possible alterna alternativ tives es for a system system & mod modeli eling ng the them m in CA CAD D software like CATIA, AutoCAD etc. to obtain a model with maximum geometric details.
When the entire powertrain was modeled, the engine bay area was found to be insufficient. A mockup of the powertrain was done and the engine bay was resized.
The models models were were then subjected subjected to analysi analysis s using using ANSYS FEA software. Based on analysis results, the model was modified and retested and a final design was frozen.
When When spac space e betw between een A-pi A-pill llar ars s wa was s incr increa ease sed d to improve field of vision, the rigidity of the frame was significantl significantly y reduced. reduced. Also, front structural structural members were too complicated to manufacture. Therefore, the front of the car was redesigned.
Dynami Dynamics cs analysi analysis s was done in Lot Lotus us suspen suspensio sion n analysis software and MSC ADAMS. The aim was to optimize suspension variables to improve mane ma neuv uver erab abiility lity.. Theo Theore reti tica call calc calcul ulat atio ions ns of performance characteristics were also done.
The The new new frame frame desi design gn is show shown n belo below. w. This This was subsequently analyzed in ANSYS for frontal impact, torsion and rollover tests.
Extensive weight reduction techniques were followed at every stage of the design to improve performance without sacrificing structural integrity.
DESIGN OF MAJOR SYSTEMS FRAME DESIGN
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fig ii (a & b) Frontal Impact Test: fig iii (a & b)
For a perfectly inelastic collision, energy transferred is DE = ½ (m 1m2 /m1+m2)(u2-u1)2 where m1 and m2 are masses of two vehicles and u1 and u2 are corresponding velocities. Assuming m1=m2=350kg and u2=0 (vehicle at rest),
Torsion Test: For torsion test, a force equivalent to the gross weight of the vehicle (3500N) was applied at one of the 4 corners of the frame while constraining the other 3.
DE = 1/4 m1u12 & F=DE/t where t=100ms
Deformation and stresses were as follows. For a stress of 163MPa, the FOS obtained was 2.12.
Then, F= [.25 x 350 x (16.67)^2] / [10x.1] = 24315N Hence, a frontal impact force of 6000N was applied at 4 points on the frame. The back of the frame was completely constrained. The deformation and stresses are shown below. For a stress of 67MPa, the FOS obtained was 5.15.
fig iv (a & b) Rollover Test: 2
In the rollover test, a force equivalent to the gross weight of the vehicle (3500N) was applied to one of the top corners of the frame while constraining the base. For a stress of 36MPa, the FOS obtained was 9.58.
fig vi For the rear upper arm, a force of 1KN was applied to the hinges and the shock absorber mounting. Max stresses were within limits.
fig vii Hubs: Front hubs are OEM and are made of cast iron with a hardened steel stub axle. Rear hubs are made of mild steel (hardened). Rear hubs were designed to incorporate the double wishbone suspension and also to enable mounting of disc brakes.
fig v (a & b) SUSPENSION DESIGN & WHEELS A double wishbone suspension setup was chosen for the front as well as rear as it is lightweight, independent and prevents deflection during hard cornering which ensures that the steering and wheel alignment stay constant. Other types like McPherson strut and trailing arm were rejected because of weight considerations. Wishbones : Material used for wishbones is same as the frame material. As seen below, for a 1KN force on the ball joint and shock absorber mounting, the max stress obtained is 63Mpa, which gives a FOS of 5.46.
fig viii Front and rear hubs were both analyzed for 3500N force applied at the bearings and were found to be within limits. Front hub shows a stress of 157MPa 3
while the rear hub shows a stress of 65Mpa. The design is well within yield limits for the materials used.
fig x (a & b) Above are the graphs for bump (mm) (x-axis) versus toe, camber and castor angles. For a bump and rebound of 100 mm each the camber was restricted within 0.5 deg and toe within 2 deg. This minimizes the forces on the knuckle ball joints during bumps.
fig ix Shock Absorbers & Wheels:
Roll:
Shock absorbers used are completely adjustable gas filled dampers (OEM from Maruti Omni) coupled with compression springs. Wheels used are tubeless bias type having R10 175 in front and R10 250 in the rear. Rims used are magnesium alloy. Dynamic Analysis : During wishbone design it was found that size of the engine bay and track width limitations were resulting in extremely short rear wishbone lengths. This would in turn limit travel of the shock absorbers and result in an extremely harsh ride and possible damage to the engine mounts. The back of the frame was then extended as a narrow portion to make longer wishbone lengths possible. Dynamic analysis was done on the front suspension setup to check the response of the vehicle for bump, in roll and while steering. Keypoints were obtained from the CAD model. Variables were tuned to reduce bump steer, camber angles and wayward movement of roll center. Bump:
fig xi (a & b) Values of toe angle, camber angle and roll center height versus roll angle (deg) (x-axis) indicate that driver will experience good control over the vehicle while cornering. Steer:
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fig xiii Brakes are disc type in front and rear, with 180mm discs in front and 130mm in rear. Brake force is distributed via 2 master cylinders so that system is independent. SAFETY & ERGONOMICS
fig xii (a & b) Steering angle (deg) (x-axis) vs. camber angle, toe angle and roll center indicates minimum deviations of all three. The Ackermann error is only 6%, which indicates an accurate and responsive steering.
fig xiv Shown above is the Impact Energy Diffuser (IED) used in the front of the vehicle to absorb energy from impacts and prevent damage to the wishbones and tie rods. It will incorporate springs and dampers to absorb forces and keep vehicle functioning after a crash.
ENGINE & DRIVETRAIN A Mahindra Alfa transmission (4 forward 1 reverse) will be used and will be directly coupled to the wheels. Gear ratios will not be modified. Engine will be mounted on rubber bushings to reduce NVH characteristics. Using a directly coupled final drive also enables the engine to be mounted as low as possible, thus lowering the C.o.G of the vehicle. STEERING & BRAKES Steering is a rack and pinion system having a lock-tolock of 2.5 turns. Steering ratio is 15:1 with Ackermann angles of 24deg and 36deg. The turning radius of the vehicle is 3.46m. The rack is placed ahead of the front wheels’ center axis to improve handling.
fig xv The driver cabin is ergonomically designed keeping anthropometry in mind. The seating is adjustable. Shown above is the tilt steering assembly used to 5
provide different steering settings depending on the user’s preferences. It utilizes a spring loaded locking mechanism to hold the steering column in preset positions. It can also be moved completely out of the way to enhance ease of ingress/egress.
Thus, net deceleration Acc=[2*F(Rw)+2*F(Rr)]/Weight of the vehicle(W) = 2(77.36+218.72)/3500 = 16.9m/s^2. And, Stopping distance (14*14)/2*16.9 = 2.89m.
EXPECTED PERFORMANCE CHARACTERISTICS
D(s)
=
V^2/2*a
=
C.o.G & WEIGHT DISTRIBUTION: C.o.G calculations were done by considering the origin at the front end for X, at the chassis for Z and at the wheels for Y. The final value for Z was arrived at after adding the ground clearance.
POWER & TORQUE: Power to Weight ratio is (10.72/275)*1000 = 39 bhp/ton. Torque is calculated as follows.
BRAKING DISTANCE: Using OEM master cylinders & assuming force applied by driver on pedal to be 85lbs = 386N, force on master cylinder = 386 x 0.26 (dist in m from pedal to cylinder) = 100.36N Now, this is equal to F x ram length, i.e. 100.36=Fx.08 so F=1254.5N Then, pressure delivered by the cylinder P=F/A = 1254.5/314.15e-4 = 39,933N/m^2 Assuming front:rear brake bias as 68:32 gives P(f)=27154.4N and P(r)=12788.6N. Hence, force applied by the rear cylinder F(r) = P(r)*A = 490.9e-4*12788.6 = 627.70N and similarly, F(f) = 1333.1N. Also Force applied on the discs by the cylinder F(R) = 2*F(r)*µ = 2*627.70*0.3 = 376.62N and F(F)=798.7N. Which implies torque on each disc in the rear= T(R)= F(R)*Radius = 376.62*0.06 = 22.6N and that on the front (with radius of the disc=0.08 m) T(F)=63.9N Finally force per wheel in the rear becomes F(Rw) = T(R)/Radius of the wheel (R(w)) = 22.6/0.292 = 77.36N and also F(Rr) = 218.72N.
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Using the graph, this gives our vehicle a four star rating.
FULL VEHICLE 3D VIEWS
fig xvii
Stability Analysis: According to the National Highway & Traffic Safety Administration, most vehicle rollovers occur by tripping over low obstacles. For a Baja vehicle, this would also be the case. Then stability is obtained from the following graph.
fig xviii
fig xix
VEHICLE TECHNICAL SPECIFICATIONS fig xvi Static Stability Factor (SSF) = T/2H where T= track width and H= height of centre of gravity. SSF=1324/(2x553)= 1.19
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Manish O. – Team Captain – 91-9844421914 Mokshith S.N – Design Head – 91-9611666646 Karthik N – Marketing Head – 91-9036227798
CONCLUSION This being Team Stratos’ first attempt at Baja SAE , our team’s objective was to design and build a vehicle that can complete all competition events without failure. All designs and calculations were done to realize this aim. Reliability and safety were considered paramount, keeping the nature of the end-user in mind. Finally, a high level of manufacturability was incorporated to ensure feasibility for mass-production.
REFERENCES 1. Chassis Engineering by Herb Adams 2. Automotive Mechanics by Crouse Anglin 3. Race Car Vehicle Dynamics by Millikens & Millikens
CONTACT 8
