IRJET- Design and Analysis of Gearbox of an All-Terrain Vehicle
0 likes97 views
This document describes the design and analysis of the gearbox for an all-terrain vehicle (ATV). It includes specifications for the engine, CVT, and other powertrain components. Gear ratios were selected to provide maximum acceleration while achieving the target speed of 50 km/hr. The gearbox was optimized to reduce weight and size. Calculations were shown for torque, tractive force, acceleration, gradeability, and stresses on gears and shafts. A two-stage gearbox with gear ratios of 2.411 and 3.06 was designed to transfer power efficiently from the engine to the wheels.
![International Research Journal of Engineering and Technology (IRJET) e-ISSN: 2395-0056
Volume: 06 Issue: 05 | May 2019 www.irjet.net p-ISSN: 2395-0072
© 2019, IRJET | Impact Factor value: 7.211 | ISO 9001:2008 Certified Journal | Page 393
Tyre Size = 23”-7”-10”.
3. Selection of Gear-Ratio.
Rolling Resistance = μmg
= 0.3*9.81*200
= 588.6 N.
Air Drag = 1⁄2*Cd*A*ρ*V^2
= 1⁄2* 0.23*0.838*1.225*229.129
= 25.05N.
Gradient Resistance = mg*sinӨ
= 200*9.81*sin (35)
= 1125.356 N.
Total Resistance = Rolling Resistance + Air drag
+ Gradient Resistance
= 588.6 + 25.05 +1125.356
= 1739.006 N.
Radius = 0.2921 m (23” x 7” x 10” in)
Torque = Total Resistance * Radius
= 1739.006 * 0.2921
= 507 Nm.
Powertrain Efficiency to be considered as 96%
507/0.96 = 526.5 Nm.
The above obtained value is for maximum resistance on
loose sand. Hence, the torque obtained from the abovevalue
will be the maximum torque. But in practical that value is
never achieved. Hence, we choose a value (for torque) that
nearest to the above calculated value.
Selected Gear Ratio = 526.5 / (18.3*3.9)
= 7.36.
Speed = [RPM (engine)* 2π * Radius]/ [60*CVT
ratio* Gear ratio]
= [3600 * 6.28 * 0.292] /[60*0.9*7.36]
= 16.6 m/s
= 60 km/hr.
Total Tractive Force = T*G*r /R
Where T = Engine Torque
G = Gear Ratio
r = CVT low ratio.
R = Radius
Total Tractive Force = 18.3*7.36*3.9 /0.2921
= 1799 N.
Acceleration = (Tractive Force-Resistances) / m
= (1799 –613.65) / 190
= 6.23 m/s.
Gradeability =
M*a = Tractive Force – (Rolling Resistance +
Air Resistance) – Gradient Resistance
190*0=1799–(588.6+25.05) –
190*9.81*sinӨ
Θ =39.419.
Gradeability % = tanӨ*100
= tan (39.41) * 100
= 82.17%.
4. Gear Calculations
For First Pair:
Gear Geometry Data Data
Gear Mesh type External
Helix type Single Helix Type
Normal Pressure angle
PHI(n)
20
Standard helix Angle
beta
15
Required gear ratio u 2.411
Note: Dimensions are in mm, all angles in degrees, and all
stresses in N/mm2.
Materials / Heat
Treatment Data
For Both Pinion
and Gear.
Material (Pinion) EN-24
Material (Gear) EN-24
Heat Treatment Toughened
Surface Hardness HRC 55
Load Data Data
Transmitted Power (kw) 6.19
Pinion Speed (rpm) 821](https://image.slidesharecdn.com/irjet-v6i580-190813061243/85/IRJET-Design-and-Analysis-of-Gearbox-of-an-All-Terrain-Vehicle-2-320.jpg)
IRJET- Design and Analysis of Gearbox of an All-Terrain Vehicle
- 1. International Research Journal of Engineering and Technology (IRJET) e-ISSN: 2395-0056 Volume: 06 Issue: 05 | May 2019 www.irjet.net p-ISSN: 2395-0072 © 2019, IRJET | Impact Factor value: 7.211 | ISO 9001:2008 Certified Journal | Page 392 Design and Analysis of Gearbox of an All-Terrain Vehicle Shivshankar Angadi1, Prajyot Palande2, Anurag Kandke3, Pratik Manjari4 B.R. Patil5 1,2,3,4Student, Department of Mechanical Engineering, MIT Academy Of Engineering, Maharashtra, India 5Assistant Professor, Department of Mechanical Engineering, MIT Academy Of Engineering, Maharashtra, India ----------------------------------------------------------------------***--------------------------------------------------------------------- Abstract - This is a detailed report about the powertrain system of the ATV vehicle designed for SAE BAJA India. Powertrain system designed with the objective to utilize the power available at the engine output shaft by reducing transmission losses comprises of main components such as engine, gearbox, CVT, half shaft, joints, etc. CVT is tuned to its ultimate efficient output by confirmingitschanging values, i.e. from 3.9 to 0.9. Gears are designed in a way to provide an aggressive launch with appropriate output torque while the optimized maximum speed could be obtained within time. Gearbox is optimized in its dimensional view reducing its weight and space required to mount it. Joint is selected to accommodate the extreme value of travel for the vehicle with considering its transmission efficiency according to varying articulation angle. Powertrain system is designed and manufactured to attain vehicle’s top acceleration as well as drawbar pull capacity and gradeability. Key Words: Efficiency, Power, Gearbox, CVT, Torque, Acceleration, Speed. 1. INTRODUCTION. The powertrain system includes engine as a power producing device. CVT assists in torque multiplication to produce the required output torque. Gearbox is Customized as per the ATV requirement as a single Speed 2 stage reduction system. The joint used is OEM withsometolerable travel and allowing varying articulation. The transmission system multiplies the Torque of the engine such that power peak is achieved to the maximum potential. The design of Components is done considering the losses of every Component both individually and as a part of the System wherever it was necessary. All the resistances were mathematically formulated in order to set the design requirements. Material for every component was selected giving preference to its design as well as performance requirement. 1.1 Engine Specifications. We Used Briggs and Stratton engine which is 4 stroke air cooled petrol engine. Torque 18.3 Nm. EngineDisplacement 305 cc. No of Cylinders Single. EngineConfiguration Horizontal Shaft. Engine Technology OHV. Mass (Kg) 28 kg. Bore (in) 3.12. Stroke (in) 2.44. Oil Capacity (dry) 24 ounces. Rpm 3800. Compression ratio 8.1 to 1. Power (HP) 10 hp. Fuel Type 87 Octane. 1.2 CVT Specifications. We used Gaged CVT gx9 Model of center to center distance 8.5 inches with high tunability options. Low ratio = 3.9. High Ratio = 0.9. Weight = 5kg. Cost = 100K. 2. Design Targets. To obtain an output speed of 50 Km on road. To reduce the overall weight and cost of theSystem. To avoid losses to the maximumpossible extent and increase the efficiency of the transmission system. To achieve maximum acceleration. To minimize the vibrations. To reduce the maintenance, increasing serviceability. To design components considering all parameters so they don’t undergo any failure after manufacturing. 2.1 Design Consideration. Coefficient of friction between tyre and road = 0.31. Coefficient of Rolling Resistance = 0.31. CVT Ratio – 1. Low ratio – 3.9. 2. High ratio – 0.9. Vehicle Mass = 220 Kg.
- 2. International Research Journal of Engineering and Technology (IRJET) e-ISSN: 2395-0056 Volume: 06 Issue: 05 | May 2019 www.irjet.net p-ISSN: 2395-0072 © 2019, IRJET | Impact Factor value: 7.211 | ISO 9001:2008 Certified Journal | Page 393 Tyre Size = 23”-7”-10”. 3. Selection of Gear-Ratio. Rolling Resistance = μmg = 0.3*9.81*200 = 588.6 N. Air Drag = 1⁄2*Cd*A*ρ*V^2 = 1⁄2* 0.23*0.838*1.225*229.129 = 25.05N. Gradient Resistance = mg*sinӨ = 200*9.81*sin (35) = 1125.356 N. Total Resistance = Rolling Resistance + Air drag + Gradient Resistance = 588.6 + 25.05 +1125.356 = 1739.006 N. Radius = 0.2921 m (23” x 7” x 10” in) Torque = Total Resistance * Radius = 1739.006 * 0.2921 = 507 Nm. Powertrain Efficiency to be considered as 96% 507/0.96 = 526.5 Nm. The above obtained value is for maximum resistance on loose sand. Hence, the torque obtained from the abovevalue will be the maximum torque. But in practical that value is never achieved. Hence, we choose a value (for torque) that nearest to the above calculated value. Selected Gear Ratio = 526.5 / (18.3*3.9) = 7.36. Speed = [RPM (engine)* 2π * Radius]/ [60*CVT ratio* Gear ratio] = [3600 * 6.28 * 0.292] /[60*0.9*7.36] = 16.6 m/s = 60 km/hr. Total Tractive Force = T*G*r /R Where T = Engine Torque G = Gear Ratio r = CVT low ratio. R = Radius Total Tractive Force = 18.3*7.36*3.9 /0.2921 = 1799 N. Acceleration = (Tractive Force-Resistances) / m = (1799 –613.65) / 190 = 6.23 m/s. Gradeability = M*a = Tractive Force – (Rolling Resistance + Air Resistance) – Gradient Resistance 190*0=1799–(588.6+25.05) – 190*9.81*sinӨ Θ =39.419. Gradeability % = tanӨ*100 = tan (39.41) * 100 = 82.17%. 4. Gear Calculations For First Pair: Gear Geometry Data Data Gear Mesh type External Helix type Single Helix Type Normal Pressure angle PHI(n) 20 Standard helix Angle beta 15 Required gear ratio u 2.411 Note: Dimensions are in mm, all angles in degrees, and all stresses in N/mm2. Materials / Heat Treatment Data For Both Pinion and Gear. Material (Pinion) EN-24 Material (Gear) EN-24 Heat Treatment Toughened Surface Hardness HRC 55 Load Data Data Transmitted Power (kw) 6.19 Pinion Speed (rpm) 821
- 3. International Research Journal of Engineering and Technology (IRJET) e-ISSN: 2395-0056 Volume: 06 Issue: 05 | May 2019 www.irjet.net p-ISSN: 2395-0072 © 2019, IRJET | Impact Factor value: 7.211 | ISO 9001:2008 Certified Journal | Page 394 Required Life (hrs) 1000 Overload factor (Ko) 1.75 Dynamic Factor ( Kv) 1.0858 Load Distribution factor (Km) 1.2 Pitting safety factor (SH) 1.5 Bending Fatigue Safety Factor (SF) 2.5 Reliability 99% Driving PINION Number of contacts per revolution 1 Stress cycle factors, curve chosen, figs. 17 & 18 = Lower (For Critical application). ***DESIGN OPTIONS*** Operating Centre Distance: By Input. Operating Centre Distance (mm) a = 60. Last Precise Solution Input Output Pinion Operating Pitch Dia (mm) 34.622 34.622 Face Width (mm) 15 15 Normal Module (mm) 2 2 Selection of Variants:- Z1 Z2 Ratio Ratio Wt. Wn. w 16 36 2.25 3.211 25.897 25.064 15.575 17 36 2.118 -2.86 23.525 22.779 15.295 17 41 2.411 -0.162 20.904 20.248 15.024 Tooth number Combination chosen (Z1, Z2):- 17, 41. Geometry Summary Pinion Gear Tooth Number (Z1,Z2) 17 41 Net Face Width (b1,b2) 15 15 Outside Diameter (do, Do) 39.199 88.800 Normal Module (mm) 2 Normal Pressure Angle ( Φ) 20 Standard Helix Angle (β) 15 Operating Centre Distance (a) 60 FOR SECOND PAIR: Gear Geometry Data Data Gear mesh type External Helix type Single helical gear Normal pressure angle, (φ) 20° Standard helix angle, (β) 15° Required gear ratio, (µ) 3.06 Materials / Heat Treatment Data For Both Pinion and Gear. Material (Pinion) EN-24 Material (Gear) EN-24 Heat Treatment Toughened Surface Hardness HRC 55 *** DESIGN OPTIONS *** Operating centre distance: By Input Operating centre distance (mm) a= 100
- 4. International Research Journal of Engineering and Technology (IRJET) e-ISSN: 2395-0056 Volume: 06 Issue: 05 | May 2019 www.irjet.net p-ISSN: 2395-0072 © 2019, IRJET | Impact Factor value: 7.211 | ISO 9001:2008 Certified Journal | Page 395 Last Precise Solution Input Output Pinion operating pitch dia. (mm) 49.2611 49.2611 Face width (mm) 15 15 Normal module (mm) 2.5 2.5 Selection of Variants:- Z1 Z2 Ratio Ratio Wt. Wn w 18 56 3.111 1.67 26.347 25.498 15.632 18 57 3.167 3.486 24.738 23.948 15.433 19 56 2.947 -3.681 24.738 23.948 15.433 19 57 3 -1.961 23.024 22.295 15.24 19 58 3.053 -0.241 21.18 20.515 15.051 Tooth number Combination chosen (Z1, Z2):- 19, 58. GEOMETRY SUMMARY Pinion Gear Tooth number 19 58 Net face width 15 15 Outside diameter 54.166 155.8242 Normal module (mm) 2.5 Normal pressure angle 20° Standard helix angle 15° Operating centre distance 100 5. Shaft Calculations:- Input Shaft Weight of pinion: 1.4 N Vertical force analysis ∑𝑀𝐴 =0 (3663.81×14) - (RB×45) =0 RBV=1139.85N ∑𝐹𝑦 =0 RA+RB=3663.81 N RAV=2523.96 N Bending moment acting in vertical plane MAV= 0 Nmm MCV = 35335.44 Nmm MBV = 0 Nmm Horizontal force analysis ∑𝑀𝐵 =0 (1216.61×197) - (RAH×45) 19303.19 + (1380.0314×31) =0 RAH=5847.78 N ∑𝐹𝑦 =0 1216.61-5847.78+1380.0314-RBH = 0 RBH = -3251.14 N Bending moment acting in horizontal plane MDH = 0 Nmm MAH = 184924.22 Nmm MCL= 120088.34 Nmm MCR = 100785.15 Nmm MB = 0 Nmm Load Data Data Transmitted power (kW) 6.19 Pinion speed (rpm) 279.14 Required life (HRS) 1000 Overload (or application) factor 1.75 Dynamic factor 1.0309 Load distribution factor 1.2 Pitting safety factor 1.5 Bending fatigue safety factor 2.5 Reliability 99% Driving: PINION Number of contacts per revolution 1
- 5. International Research Journal of Engineering and Technology (IRJET) e-ISSN: 2395-0056 Volume: 06 Issue: 05 | May 2019 www.irjet.net p-ISSN: 2395-0072 © 2019, IRJET | Impact Factor value: 7.211 | ISO 9001:2008 Certified Journal | Page 396 Maximum bending moment Resultant B.M at D = 0 Nmm Resultant B.M. at A = √184294.722 = 184924.72 Nmm Resultant B.M. at C = √120088.342 + 35335.432 = 125179.0826 Nmm Resultant B.M. at B = 0 Nmm Maximum Bending moment from above Mmax = 184924.72 Nmm Then the permissible shear stress is, 𝜏𝑝𝑒𝑟 = 𝜏 1.64 = 142.39 N/mm2 From Combined bending moment and torsion equation D = 22 mm Intermediate Shaft Pinion and Gear mounted on intermediate shaft have a weight of 2.19 N and 11.33N respectively. Vertical force analysis ∑𝑀𝐴 = 0 (-3663.03×16.5) - (8328.57×33.5) + (RBV×50) =0 RBV=6788.94N ∑𝐹𝑦 =0 RAV+RBV=11991.6 N RAV=5202.66 N Bending moment acting in vertical plane MAV= 0 Nmm MCV =-85706.57 Nmm MDV =-111931.5 Nmm MBV = 0 Nmm Horizontal force analysis ∑𝑀𝐵 =0 (1380.26×16.5) + (54866.179) -(3138.64×33.5) (RBH×50) =0 RBH=-550.079N ∑𝐹𝑦 =0 RAH+RBH= -1758.38N RAH = 2308.459 N Bending moment acting in horizontal plane MAH = 0 Nmm MCL= -38041.32 Nmm MCR = 16829.07 Nmm MDL = -45829.11Nmm MDR = 9034.58 Nmm MBH = 0Nmm Maximum bending moment Resultant B.M at A = 0 Nmm Resultant B.M. at C = 93769.71 Nmm Resultant B.M. at D = 120950.34 Nmm Resultant B.M. at B = 0 Nmm Maximum Bending moment from above Mmax = 120950.34 Nmm Then the permissible shear stress is, 𝜏𝑝𝑒𝑟 = 𝜏 2.26 = 103.3743 N/mm2
- 6. International Research Journal of Engineering and Technology (IRJET) e-ISSN: 2395-0056 Volume: 06 Issue: 05 | May 2019 www.irjet.net p-ISSN: 2395-0072 © 2019, IRJET | Impact Factor value: 7.211 | ISO 9001:2008 Certified Journal | Page 397 From Combined bending moment and torsion equation D = 26 mm. Last Shaft Gear is mounted on last shaft which have a weight of 19.01212 N. Vertical force analysis ∑𝑀𝐴 =0 (8348.69289 ×31) - (RB×45) =0 RBV=5751.3217N ∑𝐹𝑦 =0 RA+RB= 8348.69289 N RAV=2597.3711 N Bending moment acting in vertical plane MAV= 0 Nmm MCV = 80518.5047 Nmm MBV = 0 Nmm Horizontal force analysis ∑𝑀𝐵 =0 (2231.93×69.88) + (3138.7046×31) - (RBH×45) =0 RBH=5628.158 N ∑𝐹𝑦 =0 RAH + RBH = 0 RBH = -2489.4534 N Bending moment acting in horizontal plane MAH = 0 Nmm MCR = 78794.212 Nmm MCL= -77173.056 Nmm MBH = 0 Nmm Maximum bending moment Resultant B.M. at C =√ (78794.212)2 + (80518.5047)2 = 112657.7003 Nmm Maximum Bending moment from above Mmax = 112657.7003 Nmm Then the permissible shear stress is, 𝜏𝑝𝑒𝑟 = 𝜏 1.7 = 137.64 N/mm2 From Combined bending moment and torsion equation D = 32mm Bearing Calculations:- Standard factors: L10= 1000 hrs. a = 3 61904-2RS1 Loads acting due to the gear Pa = 981.3294 N Pr = 1380.017 N Pt = 3662.3113 N N = 820.51 rpm Rated life: L10 = 49.2306 million revolutions Effective Dynamic Load P = 1745 KN. Dynamic Load Capacity C = 8.39 KN. 62/22-2RS1 Loads acting due to the gear
- 7. International Research Journal of Engineering and Technology (IRJET) e-ISSN: 2395-0056 Volume: 06 Issue: 05 | May 2019 www.irjet.net p-ISSN: 2395-0072 © 2019, IRJET | Impact Factor value: 7.211 | ISO 9001:2008 Certified Journal | Page 398 Pa = 2437.8684 N Pr = 3428.309 N Pt = 9098.226 N N = 288.812 rpm Rated life: L10 = 17.3288 million revolutions Effective Dynamic Load P = 4358 KN. Dynamic Load Capacity C = 6.5658 KN. 61906-2RS1 Loads acting due to the gear Pa = 2231.635 N Pr = 3138.2886 N Pt = 8328.5768 N = 101.6944 rpm Rated life: L10 = 6.1017 million revolutions Effective Dynamic Load P = 3989 KN Dynamic Load Capacity C = 4.591 KN. Analysis of Components. Gear 1 Gear 2. G e Gear 3 with output shaft. Gear 1 with input shaft. Gear 3 and Intermediate Shaft. Results. 1. Stress generated in gear 1 is 263.19 MPA and FOS is 3.22. 2. Stress generated in Gear 2 is 199.36 and FOS is 3.51.
- 8. International Research Journal of Engineering and Technology (IRJET) e-ISSN: 2395-0056 Volume: 06 Issue: 05 | May 2019 www.irjet.net p-ISSN: 2395-0072 © 2019, IRJET | Impact Factor value: 7.211 | ISO 9001:2008 Certified Journal | Page 399 3. Stress Generated in 3rd Gear is 454.1 MPA and FOS is 2.5. 4. Stress generated in input shaft is 281.2MPA and FOS is 3. 5. Stress generated in intermediate shaft is 457.3MPA and FOS is 1.853. For above results material’s YTS is 848 MPA. Conclusion. The agenda for designing gearbox was to increase the efficiency and torque of an ATV. The Gears were designed by using Gearcalc software and cad withthehelpofCatiaV5and analysis was carried out by thesoftware‘Ansys’. Considering the efficiency of gears helical gears were selected. Thespeed and torque was selected optimumaccordingtothe event and thus gearbox of 2 stage single speed was designed and analyzed successfully. References. Standard handbook of Machine design by Joseph Shigley and Charles Mischke. Design of Machine elements by V.B.Bhandari. Automotive Transmissions by Wolfgang Novak. Textbook of Machine Design by R.S.Khurmi and J.K.Gupta. Mechanical Engineer’s Handbook by Dan.B.Marghitu. Machine elements in Mechanical Design by Robert.L.Mott. Tune to Win by Carroll Smith. Tyre Vehicle Dynamics by Hans Pacejka. Strength of Material by S. Ramamrutham.