Go Kart Report

DESIGN REPORT OF THE ECOKART VEHICLE– - BY HAMMER HEADS TEAM This design report of the Eco kart vehicle – The Hammer Heads, deals with the engineering and designing processes that we followed in the development of each and every working system of the vehicle. The system designs were made through detailed analysis of the user and vehicle requirements. With this as the base of the designing criteria for the systems are drawn using SOLID WORKS. They are virtually simulated under real time simulations by ANSYS and SOLIDWORKS taking possible number of parameters as constraints. Design validation, DFMEA and Project Planning were done using Microsoft office and Smart Sheets.  Introduction The initial point of design process started with the driver ergonomics and adding to it was safety and cost effectiveness. A tentative design was made tailoring to the initial target values. Then further changes were made to the design considering the driver sitting virtually inside an assumed model and finally modified. The design was done under newer constraints and real time simulations were done to analyze the systems. TQM tools and Design tools were used to validate the final output. 2. Vehicle Technical Specification S.no Technical description Specification 1 Dimensions Overall length Overall width Overall height 1935 mm 1280 mm 547 mm 2 Motor 2 pole, PMDC, 48 V, 800 W 3 Drive Train Continuously variable transmission CVT– Chain sprocket ratio 0.76 SCOOTY PEP 4 Frame Material ASTM A106 -Grade B Yield strength = 303 MPa Carbon = 0.3 % Density=7700Kg/m3 5 Weight 175 kg 6 Steering Ackerman steering Ratio = 1:1 7 Brakes Disc Brakes Calipers – Maruti Alto Master Cylinder – Maruti Omni 8 Wheels 16*4*3.5 Inch Honda Activa 9 Ground Clearance 5.12 Inch 3. Target Performance Value The target performance of various systems are calculated through various methods which are explained in detail in the later section. The following are the target constraints of our ECOKART vehicle we were indented to meet the during the design process. 3.1 Braking System 3.1.1 Braking force Target braking for our vehicle was calculated as 2271.5 N. The calculations are explained in the later section of braking system. 3.1.2 Stopping Distance Target stopping distance optimal for our vehicle was calculated to be 10.94m. 3.2 Steering System The target criterion for the steering system is 100 % Ackerman steering with a 1:1 ratio. 3.2.1 Turning radius The target turning radius of the vehicle was calculated to be 2.3m. 3.3 Drive and Transmission System 3.3.1 Maximum velocity The maximum velocity to be obtained by our vehicle is targeted to be 16.16 m/s taking all other constraints of the vehicle dynamic parameters. 3.3.2 Gradability The Gradability of our vehicle is targeted to be 25.8̊ 4. 3D Model of the vehicle The 3D model of the vehicle shown here is designed in the modeling software SOLID WORKS. The Frame design was made according to the loops and it is assembled with various other system components modeled in SOLID WORKS. System models very virtually evaluated through various simulation soft wares such as ANSYS, SOLID WORKS, Etc… The driver is seated and ergonomics were checked virtually. 5. Frame Design 5.1 The Frame Material Selection The design was made based on the driver ergonomics. A material of less weight without compensating the strength was taken into consideration to make the vehicle lesser in weight and optimizing the overall performance. The material was found to suit our consideration well as it had yield strength 303 MPa and a density of 7700 Kg/m3.The simulation was done with ANSYS with ASTM A 106 Grade B and its physical properties were given as input parameters. The dimensions we settled for the frame material was 1.18 inch diameter (30mm) and 5mm thickness. The thickness was made minimal to have possible weight reduction and the section is tubular with above mentioned dimensions. 5.2 Properties of Frame Materials Material –ASTM A106 Grade B Yield strength – 303 MPa Carbon composition – 0.3% c Frame material weight – 30Kg The strength is good and the weight is comparatively very less to other available materials as per specifications. Moreover the material we chose is easily available in local markets. Hence this becomes the ideal selection of material for frame. 5.3 Frame Design Consideration The main criteria in analysis were the factor of safety, even stress distribution and the maximum stress induced. The energy rate is used as the load on the nodes. 5.3.1 Front Impact The energy for the impact is calculated as 12160N, by means of the following formula, The mass is considered to be 175 Kg with a desirable factor of safety and the velocity is considered to be 16.16 m/s. The energy calculated for one second is the load to be applied for the above mentioned test. The load for one second impact comes to 11500 N. For the front impact test the front nodes are applied with the load calculated. The rear is completely constrained and the allowing displacement to occur only in direction of the load applied. The stress and displacement values are within the permissible values. 5.3.2 Rear Impact The rear impact load is same as the front impact considering a vehicle at 16.16 m/s speed and is applied to the nodes with the front completely constrained this time. The stress and displacement values are well within the permissible levels. 5.3.3 Side Impact For the side impact, the calculated preset load is used for font impact and rear impact test. In this the load is applied along the side face of the frame and the other side wheels are constrained referring to the frame. The stress and the displacement values are within the permissible values. 5.3.4 Bump For bump test the energy rate is calculated as of And the impact time is calculated for one second. Hence load is calculated from the energy rate for the rear bump, driver weight is included in the load applied. The bump test is carried out by applying loads at the respective wheel mounting location and constraining all other three wheel mounting points. The stress and displacement values are well within the permissible levels. 5.3.5 Frame Performance Parameters The minimum factor safety from the various tests comes to around 1.94. This shows that our frame design is safe. The real time simulations were done using ANSYS and SOLIDWORKS. The loops were designed with in the guide lines of ECOKART rule book. The simulation test results prove that the design is completely safe. 6. Transmission The arrangement of transmission system in our eco vehicle starts from the motor that is connected to a 0.76:1 ratio sprocket with freewheeling by means of a chain linkage. Freewheeling is provided to avoid reversing of the direction of the motor’s rotor. The motor sprocket is connected to a Continuously Variable Transmission (CVT) which is further connected to a 1:1 ratio sprocket with freewheeling by means of a second chain. This chain connects a sprocket attached to the power shaft or the rear shaft in which the rear axle is attached and the rear axle rotates as motor rotates by means of this mechanism. 6.1 Chain – Sprocket : No.1 (B/w Motor and CVT) No of teeth in motor sprocket: 17 teeth No of teeth in Pinion: 13 teeth Ratio : 0.76 6.2 Chain – Sprocket : No.2 (B/w CVT and Rear shaft) No of teeth in Pinion and CVT sprocket : 13 The CVT has a centrifugal clutch that engages at 1400 rpm. 6.3 Calculations 1) Motor RPM : 1320 rpm Wheel diameter : 0.4128m (16 inch) Primary Reduction = 0.76 CVT input RPM=1737 RPM CVT Ratio at 1737 RPM is 22.11 Power shaft rpm = 1737/22.11 = 78.56 RPM Speed (Kmph) = Power shaft rpm * * Wheel Dia * 60 /1000 = 78.56*3.14*0.4128*60*10-3 = 6.1 Kmph 2) Motor RPM : 1680 rpm Wheel diameter : 0.4128m (16 inch) Primary Reduction = 0.76 CVT input RPM= 2210.52 RPM CVT Ratio at 2210.52 RPM is 16.71 Power shaft rpm = 2210.52/16.71 = 132.28 RPM Speed (Kmph) = Power shaft rpm * * Wheel Dia * 60 /1000 = 132.28*3.14*0.4128*60*10-3 =10.29 Kmph 3) Motor RPM : 2100 rpm Wheel diameter : 0.4128m (16 inch) Primary Reduction = 0.76 CVT input RPM=2763RPM CVT Ratio at 2763 RPM is 10.80 Power shaft rpm = 2763/10.80 = 255.833 RPM Speed (Kmph) = Power shaft rpm * * Wheel Dia * 60 /1000 = 255.833*3.14*0.4128*60*10-3 = 19.9 Kmph 4) Motor RPM : 2280 rpm Wheel diameter : 0.4128m (16 inch) Primary Reduction = 0.76 CVT input RPM=3000RPM CVT Ratio at 3000 RPM is 8.997 Power shaft rpm = 3000/8.997 = 333.44RPM Speed (Kmph) = Power shaft rpm * * Wheel Dia * 60 /1000 = 333.44*3.14*0.4128*60*10-3 = 25.93 Kmph 5) Motor RPM : 2520 rpm Wheel diameter : 0.4128m (16 inch) Primary Reduction = 0.76 CVT input RPM= 3315.78PM CVT Ratio at 3315.78 RPM is 6.584 Power shaft rpm = 3315.78/6.584 = 503.61 RPM Speed (Kmph) = Power shaft rpm * * Wheel Dia * 60 /1000 = 503.61*3.14*0.4128*60*10-3 = 39.17 Kmph 6) Motor RPM : 3600 rpm Wheel diameter : 0.4128m (16 inch) Primary Reduction = 0.76 CVT input RPM= 4736.84 RPM CVT Ratio at4736.84 RPM is 6.33 Power shaft rpm = 4736.84/6.33 = 748.31 RPM Speed (Kmph) = Power shaft rpm * * Wheel Dia * 60 /1000 = 748.31*3.14*0.4128*60*10-3 = 58.2 Kmph 6.4 Torque Exerted on the wheels The maximum motor torque = 12.8 Nm Transmission efficiency = 70% = 0.7 Formula Drive torque (Nm) = Motor Torque * combined gear ratio * transmission efficiency Combined gear ratio = Primary reduction * CVT Ratio Calculation 1) Drive Torque = 12.8*0.76*22.11*0.7 = 150.56 N-m 2) Drive Torque = 12.8*0.76*16.71*0.7 = 113.78 N-m 3) Drive Torque =12.8*0.76*10.80*0.7 = 73.54 N-m 4) Drive Torque =12.8*0.76*8.997*0.7 = 61.26 N-m 5) Drive Torque =12.8*0.76*6.584*0.7 = 44.83 N-m 6) Drive Torque =12.8*0.76*6.33*0.7 = 43.104 N-m 6.5 Vehicle Tractive Effort (or) Drive Torque Formula Drive Force = Drive torque / Wheel Radius Calculation 1) 150.56/0.2064 = 729.45 N 2) 113.78/0.2064 = 551.26 N 3) 73.54/0.2064 = 356.29 N 4) 61.26/0.2064 = 296.80 N 5) 44.83/0.2064 = 217.20 N 6) 43.104/0.2064 = 208.84 N 6.6 Gradability Drive force = mg sin Ɵ m = 175 kg g = 9.81 m/s-2 Ɵ = sin-1 ( Drive force / mg ) Calculation Ɵ = sin-1( 729.45 / (175 * 9.81)) = 25.81 Ɵ = sin-1 ( 278.92 / (175 * 9.81)) = 18.7 Ɵ = sin-1 ( 232.46 / (175 * 9.81)) = 11.97 Ɵ = sin-1 ( 189.29 / (175 * 9.81)) = 9.95 Ɵ = sin-1 ( 167.34 / (175 * 9.81)) = 7.26 Ɵ = sin-1 ( 159.39 / (175 * 9.81)) = 6.98 7. Braking System 7.1 Design Methodology List design criteria or requirements Calculate the tyre forces in static condition Assign maximum deceleration of the vehicle Calculate target stopping distance and target braking force Select optimum brake parts which in combination help achieve target Optimize brake design 7.1.1 Tyre force in static condition Total weight of the vehicle under static condition = 175 kg Wheel Base (l) = 51.18” Distance of CG from front axle= 32.18” Distance of CG from rear axle= 19” Force on front wheels (2FZ1) = m*g*a1/l = 1079.42 N Force on rear wheels (2FZ2) = m*g*a2/l = 637.32 N Using the values of 2FZ1 and 2FZ2 we have calculated the weight distribution to be 63/37 in favor of front. 7.1.2 Calculation of target performance of the braking system Maximum deceleration = 1.32g Total weight of the vehicle = 175 kg Maximum velocity of the vehicle= 16.67m/s Target braking force = deceleration * total mass = 12.98 * 175 = 2271.5 N Target stopping distance = Max velocity 2 / (2 * max. deceleration) = 10.94 m 7.1.3 Design of brake system elements Brake pedal Passenger cars generally use a pedal ratio of 4 to 6. We select a pedal ratio of 4, i.e., when the driver applies 1N force it gets multiplied by 4 times to produce a 4 N force. Master cylinder Analyzing the available ones, we chose the master cylinder MARUTI OMNI. The piston diameter of the MC is 19.05 mm. Caliper and Rotor Selection We compared the rotor caliper assemblies of two OE namely Hyundai Santro and Maruti Alto. Finally we zeroed in on Maruti ALTO which had the following dimensions Dimensions of the rotor = 230 mm Caliper piston Diameter = 42 mm 7.1.4 Calculation of stopping distance with the selected components Maximum force a driver can apply is 22 kgf. Force on the Master cylinder piston (FMC) = 22*9.81*4 = 863.28 N Area of the MC Piston (AMC) = π*0.019052/4 = 2.85*10-4 m2 Pressure developed in the system (P) = force / area = 863.28/(2.85*10-4) = 3031179.775 N/m2 Since the pressure in the system is entirely the same, Therefore, the force on the brake caliper (FCAL) = P * ACAL = 4199.524 N Therefore the force on the rotor = 4199.524 * 2 = 8399.04N Total friction force = Clamp force * Coefficient of friction = 8399.04 * 0.3 = 2519.71 N Therefore the torque on the rotor = Friction force * effective rotor radius = 2519.71 * 0.09165 = 230.814 Nm Therefore the force acting on one tyre = Torque on rotor / radius of tyre = 230.814 / 0.2032 = 1135.8957 N Total brake force using selected brake force = 1135.8957 * 2 = 2271.7914 N Deceleration = Force / mass = 2271.7914/175 = 12.98 m/s2 Stopping Distance = (Max velocity)2 / (2 * deceleration) = 16.672/(2*12.98) = 10.74 m Since this system meets the target performance, we go ahead with the design Maximum Deceleration = 2271.7914 / (175*9.81) = 1.32g 7.1.5 Calculating dynamic weight Transfer and Brake Distribution Weight transfer = Weight of the vehicle * max deceleration* height of COG/Wheel base= 98.77 N Dynamic load on the front wheels during braking = Static front load + weight transfer = 1178.19 N Dynamic load on the rear wheels during braking = static rear load – weight transfer = 538.55 N Torque required to stop front wheel = dynamic load / 2 * Radius/2 * 1.32 = 79 Nm Torque required to stop rear wheel = dynamic load/2*Radius/2*1.32 = 36.11 Nm 8 Electricals The electrical circuit of the design comprises of 4 Batteries (12V, 20 Ah) connected in series as shown in the figure. The series connection of batteries is in turn connected to a Pulse Width Modulation circuit that is used to control the speed of the motor by means of varying voltage supply to it. The Pulse Width Modulation is connected to the motor’s positive terminal and the negative is grounded. 8.1 Pulse Width Modulation (PWM) Pulse-width modulation (PWM), or pulse-duration modulation (PDM), is a modulation technique that conforms the width of the pulse, formally the pulse duration, based on modulator signal information. Although this modulation technique can be used to encode information for transmission, its main use is to allow the control of the power supplied to electrical devices, especially to inertial loads such as motor. The average value of voltage (and current) fed to the load is controlled by turning the switch between supply and load on and off at a fast pace. The longer the switch is on compared to the off periods, the higher the power supplied to the load is. Pulse Width Modulation, as it applies to motor control is a way of delivering energy through a succession of pulses rather than a continuously varying analog signal. By increasing or decreasing pulse width, the controller regulates energy flow to the motor shaft. The motor’s own inductance acts like a filter, storing energy during the ‘ON’ Cycle while releasing it at a rate corresponding to the input or reference signal. In other words, energy flows into the load not so much the switching frequency, but at the reference frequency. PWM is somewhat like pushing a playground style merry go round. The energy of each push is stored in the inertia of the heavy platform, which accelerates gradually with harder, more frequent, or longer lasting pushes. The riders receive the kinetic energy in a very different manner than how it is applied. 9. Steering The steering system used in our vehicle has its tie rod directly connected to the steering rod in the ratio of 1:1. The Ackermann angle of this type of steering is calculated as follows: Track width (TW) = 1040 mm Wheel base (WB) = 1300 mm Where β = Ackermann angle The value of turning radius is assumed to be 2.3m. Also the values of TW and WB are 1040mm and 1300mm. Substituting these values, the value of Ackermann arm is found to be 519.8143mm. Now, the values of inner and outer angles are: 10. Innovation 10.1 Continuously Variable Transmission 10.2 Frame Optimization