The Propulsion System Engineering Essay

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The propulsion system with respect to this application can be defined as the system which provides vehicle motion. Thus, this project involves the design of a system for efficient power generation and transmission of power from power plant to the driving wheels with minimum power train losses.

All design features must comply with the shell Eco-Marathon Asia 2010 rules and regulations [1]

Main objectives

Selection of a suitable energy source to power the vehicle

Overall system design

Stock selection and design of components

Detailed analysis and optimization of each sub system for maximum fuel efficiency

CAD modeling and detailed drawing

Verification of selected and designed components through calculation and suitable simulation software

Final performance estimation

All design features must be approved by shell Eco-Marathon organizers

Parts quotation

1.2 About the competition

The principle of the Shell Eco-Marathon is simply to design and build the most fuel efficient vehicle while producing the fewest emissions. [1], [30]

Teams can enter two main categories:

Futuristic prototypes

These are streamlined vehicles where the primary design consideration is reducing drag and maximizing power train efficiency. This category has fewer restrictions. [1]

Urban Concept vehicles

These are built to more conventional 4-wheel roadworthy criteria. [1]

Design requirements and other rules related to the propulsion system are listed in Appendix F

1.3 Competition category and energy source selection

After few days of research and discussion considering time, cost and expertise of team, it was decided to compete under the prototype category with an internal combustion engine running on ethanol as the power plant.

CHAPTER 2.0

LITERATURE REVIEW

2.1 Introduction

This literature review provides details on past endurance vehicles and the latest developments in the area of fuel efficiency, alternative fuels and future propulsion systems.

Articles and reports were found related to vehicles designed and developed mainly for Shell Eco-Marathon and SAE Supermileage competitions.

2.2 The world record

The most fuel efficient car in the world, PAC Car II designed and developed by ETH Zurich (Swiss Federal Institute of Technology) was powered by a Hydrogen fuel cell and it had a record of 12600 Miles per Gallon (US) during the Shell Eco-Marathon, France in 2005 [3]. This clearly indicated the level of competition, amount of potential for fuel efficiency and alternative fuels.

2.3 Ethanol as vehicle fuel

Articles [4] on ethanol combustion and conversion of gasoline engines were found which provided detailed practical explanation.

An ethanol powered car engineered by French high school students from Lycée La Joliverie had achieved the best fuel efficiency at the European Shell Eco-marathon 2006, winning the race at the Nogaro auto racing circuit in southwest France by travelling 2885 kilometers per liter of gasoline equivalent. It also took the Climate Friendly prize for producing the least greenhouse gas emissions. [6]

Currently ethanol E85 (85% ethanol) powered vehicles are produced by leading automotive manufacturers such as Ford, Chevrolet, Chrysler, Toyota, Nissan etc [8]

2.4 Engine and drive train

Vehicle designed and built by Alerion Supermileage team from Laval University of Quebec, Canada won the grand prize of the Shell Eco-Marathon Americas 2010 recording 1057.5 kilometers per liter. This vehicle consisted of an internal combustion engine.

The Dalhousie University team had used the Honda GX35 engine for their vehicle with a direct drive transmission system for Shell Eco-Marathon Americas in 2008/2009 which travelled 332.8 km/l gasoline [5]. Experimental values obtained by dynamometer testing for the GX35 engine were also published. [2]

2.5 Fuel delivery systems

Most endurance vehicles in previous competitions had mechanical fuel pumps to pressurize fuel. The Dalhousie team used a pressurized fuel system in 2008/2009 which yielded successful results. Fuel injection systems are known to be more efficient than carburetor systems since there is more control over the spray of fuel.

2.6 Intake and exhaust system

The internal combustion engines and fluid mechanics online lecture notes published by the Colorado State University provided basic explanation about intake/exhaust tuning. Current developments in this field are related to variable valve timing.

Fiat was the first automotive manufacturer to patent a variable lift system. Developed by Giovanni Torazza 1960, the system used hydraulic pressure to vary the fulcrum of the cam followers (US Patent 3,641,988). The hydraulic pressure changed according to engine speed and intake pressure. [43]

After continued improvement, a system with variable valve timing, two stage valve lift on the intake valves and variable timing of the exhaust valves was developed by Porsche in 2009. [40]

In 2010, Mitsubishi developed and started mass production of the 4N13 1.8 L Diesel Overhead Cam Inline 4 cylinder engine. This is the first passenger car with a diesel engine that features a variable valve timing system. [39]

2.7 Latest trends in fuel efficiency

There is much research and development in the area of hydrogen fuel cells and hybrid systems. Although the principle of the fuel cell was discovered in 1838, it has not been a popular topic until recent years. Currently, hybrid systems and hydrogen fuel cells are considered the future of vehicle propulsion systems.

CHAPTER 3.0

ENGINE - THE SYSTEM POWER PLANT

3.1 Introduction

A small four stroke gasoline engine was required to be selected and modified for ethanol combustion with maximum fuel efficiency.

3.2 Fuel - Ethanol E100

The main idea behind selecting ethanol as the fuel for the engine is that ethanol has a high octane value (Higher auto ignition point). Therefore the engine will not knock at higher compression ratios. [8], [4]

It is also a renewable fuel produced by corn, sugar cane etc and although controversial is regarded as generating less toxic emissions. [8]

3.3 Stock engine selection

Main specification guidelines set in order to select few potential engines that can be used for the system:

Engine type:

Air cooled, four stroke, single cylinder petrol engine

Displacement:

35 to 50 cm3

Power Output:

Maximum

Mass:

Minimum

Table 1.0: Specification guidelines for selection

Adhering to the set guidelines (refer table 1.0) few potential engines were selected.

Scooter engine, 139QMB, [9]

350R/S 35cc Robin/Subaru 4-Stroke engine used for bicycles, [10]

Honda GX35 mini 4 stroke used in lawn mowers, [11]

Gasoline engine (142F) - manufactured and supplied by Shandong Huasheng Zhongtian Machinery group CO.LTD, [12]

[Detailed specifications for these engines are given in appendix C]

3.3.1 Parametric study

Main parameters considered for this study were specific fuel consumption, mass, ease of modification, availability, versatility and reliability

The fuel efficiency of an engine is directly related to the brake specific fuel consumption. [13]

During the autumn semester, the drive train was assumed to be 100% efficient and mass of car without engine was estimated as 330lbs.

These values were only used for comparing engines.

The miles per gallon values were estimated using the Bowling and Grippo program for various BSFC values. [7]

The estimates for coefficient of drag, frontal area, tire inflation pressure, vehicle weight were obtained from other team members in charge of each sub assembly.

Program inputs

Coefficient of drag

0.13

Frontal area (Square feet)

8.04

Vehicle miles per hour (MPH)

18

Vehicle weight in lbs

330

Tire inflation pressure in psi

80

Engine Brake Specific Fuel Consumption (gal/hr-hp)

Inputs from 0.01 to 0.25

Drive train horsepower loss

Assumed 0 for engine comparison

Table 2.0: Bowling and Grippo program inputs

(Refer Table 6.0 in Appendix D for results)

Figure 1.0: Estimated MPG vs. BSFC

The specific fuel consumption of the engine should be a minimum to obtain high miles per gallon of fuel (from figure 1.0).

The engines selected were further short listed considering the information available and ease for modification.

142F Gasoline engine:

Brake specific fuel consumption = 480 g/kW-h

Density of gasoline = 2790.38 g/gal, [15]

480 g/2790.378 g/gal = 0.1720 gal

0.1720 gal/1.341 hp-h

BSFC = 0.1283 gal/hp-h

Honda GX35

Brake specific fuel consumption = 360 g/kW-h

This value in gal/hp-h = 0.0962 gal/hp-h

Miles per gallon was estimated (assuming no power loss in drive train) for each engine while considering the mass of each engine.

Estimated weight of vehicle without engine = 330 lbs

Weight of vehicle with 142F engine = 335.5 lbs, this gives an estimated MPG of 790.84.

Weight of vehicle with Honda GX35 = 333.8 lbs, this gives an estimated MPG of 1058.84.

[7]

Figure 2.0: MPG for 142F and GX35 engines

Although miles per gallon value would be lower when drive train power loss is considered, the engine was compared assuming a 100% efficient drive train.

These calculations are based on gasoline fuel for engine comparison purposes.

Study results clearly indicate that the Honda GX35 is the most suitable engine for this system and also considering the reliability factor of Honda further proves that this engine should be selected for this application.

3.4 Honda GX35

Main advantages of Honda GX35 engine [11]

Lower brake specific fuel consumption

Mass is almost 2 kg less than 142F

Better reliability and it has been improved over the past 10-11 years.

Over Head Cam engine: Carrying out modifications on the cylinder head is easier.

360o inclinable

The only disadvantage is that it consists of a carburetor. A fuel injection system would have been more fuel efficient but electronic fuel injection would also require an Engine management system and alternator which would add more weight to the vehicle. Therefore weight is less with carburetor.

Performance curves

C:\Documents and Settings\Pulsara\Desktop\FYP RESEARCH\curve_GX35.gif

Figure 3.0: Performance and fuel consumption curves [11]

Basic calculations

Clearance Volume (Stock GX35)

Swept Volume = 35.8 cm3, Compression ratio, rc = 8:1

Compression ratio = Total Volume, (Vc+Vd)/Clearance Volume, Vc

Clearance Volume, Vc = 5.114 cm3

Brake mean effective pressure (BMEP), (Stock GX35)

Maximum power output = 1.3 HP at 7000rpm

L is displacement in liters, L = 0.0358 l

BMEP = 67.44 lbf/in2

Calculation of specific fuel consumption with ethanol without any system modifications

Fuel

NCV (KJ/l)

NCV (KJ/gal(US))

Gasoline

31627.84

119724.42

Ethanol

21229.48

80362.34

Table 3.0: Net calorific values by volume [1], [15]

BSFC with ethanol

0.0962 gallons of gasoline = 119724.42 x 0.0962 KJ = 11517.45 KJ

Therefore BSFC of Honda GX35 engine with ethanol fuel = 0.1433 gal/hp-h

(GX35 specifications in appendix C).

3.5 Modifications

Figure 4.0: Engine modifications

3.5.1 Mandatory modifications for ethanol combustion

Carburetor modification

Main jet changes

Since the energy density of ethanol is lower than gasoline, the fuel/air ratio should be increased. The main jet orifice can be bored out to increase the size of the orifice by around 30% of the original size. The air/fuel ratio for ethanol combustion should be 10.07:1. [4], [1]

Idle orifice changes

When the throttle plate is at idle position, the air/fuel mixture is only allowed to enter the manifold through the idle orifice. The idle mixture screw could be loosened or orifice could be bored out to increase the size by 30% in order to provide sufficient ethanol to keep the engine running at idle speed. [4]

Overall engine and piping system

Ethanol is a strong cleaning agent and has the ability to degrade certain engine parts such as, natural rubber, plastics, and even metals over time. Therefore, all rubber and plastic components should be replaced by synthetic material. [4]

It is recommended to use neoprene hoses for the fuel delivery system. [4]

Durability of various plastics: Ethanol vs. Gasoline in table 3.1in appendix D.

3.5.2 Modifications for maximum fuel efficiency

Compression ratio alteration

This is discussed with detailed analysis in chapter 5.0

Intake and exhaust optimization

This is discussed with detailed analysis in chapter 6.0

Starting system

An electric starter would be installed which would enable the driver to turn off the engine and coast after reaching a particular speed and restart later with ease.

Choke

A manually controlled choke is better for ethanol engines and especially for this competition. Therefore if the engine is equipped with an automatic choke; it can be adapted for manual control using a manual choke conversion kit.

CHAPTER 4.0

COMPRESSION RATIO ALTERATION

4.1 Introduction

This is the main advantage of using ethanol as fuel. The compression ratio can be increased up to 16-20:1 without engine knock. [20]

Increasing the compression ratio increases the thermal efficiency of the engine but it should only be increased to an extent to which the engine could withstand the pressure and temperature.

Methods to increase the compression ratio [21]

Cylinder head and block can be shaved by milling (planning) the surfaces.

Modify or change the piston head.

Inlet conditions (High pressure, temperature etc)

Reduce gasket thickness

4.2 Analysis

For this analysis, the combustion chamber of the engine was assumed to be cylinder shaped

r = Bore/2 = 1.95cm

X

VC (Stock) = 5.114 cm3

Ï€r2 X = 5.114

X = 0.428 cm

Milling the head/block or reducing thickness of the gasket would reduce X which would result in a smaller clearance volume, VC.

A smaller clearance volume results in a higher compression ratio which also generates more power.

Valve clearance for Honda GX35

This is the maximum distance the valve travels beyond the engine head.

Intake Valve clearance

Exhaust Valve Clearance

0.08 +/- 0.02mm

0.11 +/- 0.02mm

Table 4.0: GX35 Valve clearance [17]

The value of X after modifications must be greater than 0.13 mm to avoid valve/piston collision.

Let Y be the amount of head/block milled or reduced from gasket

X = (4.28 - Y) mm

There is no direct theoretical relationship between horsepower and compression ratio but the Bowling and Grippo program provides a rough estimate which was tabulated in table 6.1 in appendix D. [7]

For specific new fuel consumption

Tabulated results can be found in Table 6.1 in Appendix D

Figure 5.0: Compression ratio vs. reduction in combustion chamber height (Y)

Figure 5.1: Estimated engine HP vs. reduction in combustion chamber height (Y)

Figure 5.2: BSFC vs. reduction in combustion chamber height (Y)

Figure 5.3: Estimated miles per gallon vs. reduction in combustion chamber height (Y)

From research it was found that the compression ratio could be increased to 16-20:1 with ethanol fuel without knock problems but there was no credible information on how much compression the engine could withstand. Therefore, it was specified to increase the compression ratio only up to 12:1.

This increase in compression ratio would result in an increase of 55 miles per gallon (US)

(Refer figures 5.0, 5.1, 5.2 and 5.3)

4.2.1 Engine cyclic analysis

Figure 6.0, P-V diagram for naturally aspirated Spark ignition engine [25]

Inlet conditions:

Pressure (P) = 1 bar, Temperature (T) = 303 K, Ideal Gas constant (R) = 287 J/kg. K,

Ratio of specific heats (γ) = 1.4, CV = 717.6,

For perfect gas,

Where, is the total mass of charge mixture

From fuel consumption calculations using net calorific values:

Fuel /Air ratio (FAR) of Ethanol = 1.49 x FAR of gasoline

FAR (Ethanol) = 1/15 x 1.49 = 0.0993

, where

From 1 2 (Refer figure 6.0)

Isentropic compression

From 2 3 (Refer figure 6.0)

Constant volume heat addition

Energy density of Ethanol = 30 MJ/kg

Most small engines have thermal efficiencies between 40 and 45%. Therefore with a compression ratio of 12:1, conversion efficiency (Formation and combustion) can be assumed to be 45% to obtain an overestimate of the increase in pressure and temperature. [22]

Above calculations were repeated for the original compression ratio (8.0:1) of the stock engine which gave the following results

C:\Users\PMG\Desktop\FYP RESEARCH\T0512e0v.gif

Figure 7.0: Thermal efficiency increase with increase in compression ratio, [23]

Assuming a conversion efficiency of 40% and an air/fuel ratio of 15:1

Therefore, the peak in cylinder pressure has been increased by a factor of 1.64. This factor is also the factor of increase in force on piston, head, valves etc.

Brake mean effective pressure (BMEP) is a valuable measure of the capacity of an engine to do work and is independent of displacement (Size of engine). [24]

The BMEP of the stock engine was 67.44 (from calculations under parametric study).

Therefore, increasing the compression ratio has increased the capacity of the engine to do work significantly.

CHAPTER 5.0

INTAKE AND EXHAUST OPTIMIZATION

5.1 Introduction

A pressure wave is created when an intake or exhaust valve is opened/closed. The wave propagates through the pipe at the speed of sound. When this wave encounters a change in cross sectional area, such as the end of the pipe, a wave of opposite sign will be reflected which would travel back towards the port. Based on the time taken for this wave to return to the valve and also considering the open/close durations of the valves, the optimum length for the pipe can be calculated. This would increase the volumetric efficiency of the engine. [16]

5.1.1 Optimum intake pipe length

Experiments have revealed that there is a significant gain in volumetric efficiency when the reflected compression wave returns when the piston is at a crank angle of 90o. At this point the piston would be moving at maximum speed. Matching the time taken for the wave to return with engine speed, the required length of the pipe can be found. [16]

Velocity of wave = Distance/Time, (where distance = 2L)

Time = 900/ RPM (revolutions/minute)(minute/60s)(3600/revolution) = 15/RPM

[16]C:\Documents and Settings\Pulsara\Desktop\FYP RESEARCH\fluid0{image1}.gif

Where c is the speed of sound which depends on the temperature

Where = Ratio of specific heats

R = Ideal gas constant

T = Temperature

5.1.2 Optimum exhaust pipe length

At blow down (exhaust valve opens), a compression wave is propagates through the pipe and when it meets the end of the pipe an expansion wave returns back to the port. Experimentally it has been revealed that the optimum position of the piston when the wave returns is 120o. At this position the exhaust gas can be scavenged from the combustion chamber efficiently. [16]

Time = 1200/RPM (360/60) = 120/RPM

[16]C:\Documents and Settings\Pulsara\Desktop\FYP RESEARCH\fluid0{image4}.gif

Graphs were plotted using these formulae

A detailed calculation also considering the valve timing of Honda GX35 could be found under detailed calculations (5.2)

Figure 8.0: Intake pipe length vs. engine RPM at different temperatures

(Tabulated results in table 6.2 in appendix D)

Figure 8.1: Exhaust pipe length vs. engine RPM at different temperatures

(Tabulated results in table 6.3 in appendix D)

5.2 Optimum pipe length calculations in detail

5.2.1 Intake pipe length considering valve timing

Intake valve opens at 10o before top dead centre (BTDC) and intake valve closes at 57o after bottom dead centre (ABDC). [18]

Duration of 247o , [18]

Intake valve opens once every two revolutions.

Therefore (360 x 2 - 247) o = 473o

After closing, the intake valve would open again after 473 crank angle degrees

473o =

Speed of sound at an intake temperature of 30oC

Ratio of specific heat, γ = 1.4 at 30oC

Distance travelled is two times the pipe length,

Therefore,

The optimum pipe length for GX35 engine to run at 5100 RPM is 2.705 m

Due to the space constraint of the engine compartment pipe length can be shortened by a factor of four, making it 0.677 m in length. By this method, the wave would travel up and down the pipe four times before the intake valve opens again. Although the effectiveness would be less, it would still arrive at the correct time to force more air into the cylinder.

Using this result a custom intake pipe was designed with a length of approximately 0.5 m leaving the remaining 0.177 m for intake runners, carburetor, etc

5.2.2 Exhaust pipe length using valve timing

There are various methods and theories used for calculating the exhaust pipe length. The intake and exhaust can be treated separately to find the optimum length for each pipe and also both can be treated as one system during valve overlap to gain an added advantage during the overlap period.

Method 1 (Considering exhaust system only)

The reflected pulse could be set to arrive at the engine just as the exhaust valve starts to open, which would help to expel the exhaust gas without using up excess energy.

Exhaust valve opens at 48o before bottom dead centre (BBDC) and exhaust valve closes at 28o after top dead centre (ATDC). [18]

Duration of 256o, [18]

Port opens/closes once every two revolutions. Therefore, exhaust valve opens 464 crank angle degrees after closing

Speed of sound, c at the exhaust will depend on the exhaust temperature

Thermodynamic calculations were continued from point 3 (Refer Figure 6.0) in order to calculate the temperature at blow down.

,

From 3 4

Isentropic expansion

This length can be shortened by a factor of four allowing the wave to travel up and down four times before the valve starts to open, which gives 1.163 meters.

Method 2 (Considering valve overlap period)

If the reflected expansion wave reaches the opened exhaust valve just before closing but after the intake valve opens, the expansion wave will travel across the cylinder (since effective cylinder volume is small near TDC) through the intake port up to the intake atmosphere. This would result in an increased aspiration.

Intake/exhaust valve overlap period of 38o

Blow down shock wave leaves at 48o BBDC and the expansion wave must be set to return at around18o ATDC. [18]

This gives duration of 246o, [18]

Exhaust valve opens once every two revolutions.

To obtain maximum volumetric efficiency by gaining advantage of the valve overlap period the exhaust pipe should be 2.46 meters in length

This can be shortened by a factor of two which would make the wave travel up and down twice before making use of the valve overlap period but this method may not be effective since the exhaust port will be open when the valve returns for the first time.

5.3 Custom parts in the intake/exhaust system

Intake pipe

This pipe was designed considering the calculation results (figure 2.0) and compartment space

Length is approximately 0.5 meters, (CAD drawing in appendix A)

Exhaust pipe

This was designed considering the calculations (figure 2.1), compartment space and also the Shell Eco-Marathon rule which states that exhaust should be evacuated outside the vehicle but the pipe should not be longer than the body

Length is approximately 1.5 meters, (CAD drawings in appendix A)

Velocity stack

Figure 9.0 Inlet flow [19]This is a pipe with a curved inlet which should be fixed to the end of the intake pipe. This would give a smoother flow of air into the intake pipe which would result in better atomization of fuel in the carburetor. Also this allows the full cross section of the intake pipe to be used whereas without a curved inlet, the flow area would be reduced due to the sharp entry. Therefore the velocity stack helps to aspirate more air into the system. [19]

(CAD drawing in appendix A)

5CHAPTER 6.0

FUEL DELIVERY SYSTEM

6.1 Introduction

The basic function of this system is to deliver the fuel to the carburetor. In regular vehicles, either mechanical or electric fuel pumps are used to pressurize and drive the fuel into the system. Stock GX35 engine uses gravity to pressurize the fuel when used in lawn mowers.

The Shell Eco-Marathon rules state that electric fuel pumps are not allowed. [1]

Therefore the possible methods would be:

Using a Mechanical fuel pump

Using gravity

Pressurized fuel delivery system using compressed air

6.2 Selection and design

A mechanical pump would have to be powered from the engine output, which would result in an additional load on the engine. This would result in a reduction in specific fuel consumption.

The pressure due to height may not be large enough due to the space constraint in the engine compartment if gravity is used. Also, the shell fuel tank could be pressurized up to 5 bar which makes the pressurized system ideal for this application

Figure 10.0:

PRESSURIZED FUEL DELIVERY SYSTEM LAYOUT

Valve to drain the fuel (Shell requirement)

Solenoid cut off valve

Pressure control valve

A 1.5L pop bottle is used to store air at high pressure and air is regulated using a pressure control valve to control the pressure of air entering the fuel tank. This air at high pressure is used to push the fuel through the system.

Pressure gauge is positioned close to the fuel tank to indicate the pressure of air entering the fuel tank.

An air pump (hand held or foot pump) can be used to pump in air through, the valve stem (No return valve).

1.5 L pop bottles are rated at 72 psi, therefore it is recommended to pump the bottle to approximately 60 psi.

6.3 Advantages of this method

No extra load on the engine to drive a mechanical pump

Less weight

Low cost

Also, this is a proven method which has been used in successful endurance vehicles in past competitions. [2]

(CAD drawings can be found in appendix A)

Figure 10.1: Screenshot from CAD model showing the fuel delivery system

CHAPTER 7.0

DRIVE TRAIN

7.1 Introduction

The vehicle consists of three wheels, two in front and one at the back. The vehicle was designed to be a rear wheel drive and a chain is used to drive the wheel. Design and selection of transmission system and parts, gear ratio calculations, overall system layout and basic stress analysis is discussed in this chapter.

7.2 Transmission system

There were few potential transmission concepts that could be used

Continuously Variable Transmission (CVT). [26]

Derailleur [27]

Manual two/three speed gear box

Direct drive system [2]

CVT is known to be more efficient than a manual gear box but after further review it was found that CVT is less efficient at low speeds. [41]

Derailleur system is highly efficient and simple but previous vehicles with this system had problems with the chain slipping out of the sprocket during gear change. [2]

Although a manual gear box is well suited for this system, it would add extra weight and also more moving parts results in additional power loss.

Therefore a direct drive system with one gear ratio was selected as the most suitable transmission system.

Main reasons behind selecting direct drive transmission system

Honda GX35 has relatively flat curves for torque, power and fuel consumption. The fuel consumption curve is almost flat from 3000 to 6000 RPM. Therefore the engine can run at a wide range of speeds and still supply adequate power with the same fuel consumption. (Refer figure 3.0)

A manual gearbox would add extra weight and benefits of it would be negligible due to the linear performance curves.

7.3 System layout

Few drive train system layouts were drawn in order to allocate space for each part. The most suitable layout was selected and modified accordingly.

FUEL TANK

CLUTCH

Figure 11.0: Selected layout

(Initial concept layouts in Appendix F)

Figure 11.1: Screen shot from CAD model of the propulsion system showing the layout

(Overall CAD assembly and exploded drawings in appendix A)

7.4 Gear ratio

Since a direct drive system was selected, the drive train would have one fixed gear ratio from engine to rear sprocket.

An overall gear ratio of 16:1 was chosen and calculations were carried out to verify that this ratio is suitable for our application.

7.4.1 The torque required at the rear wheel to move vehicle from rest

Rolling resistance,

Where,

= Coefficient of rolling resistance

m = Total mass of the vehicle

g = Acceleration due to gravity

Coefficient of rolling resistance for pneumatic tires on a dry surface can be approximated by the following equation

Where, P = Tire Pressure (bars)

U = Vehicle velocity (km/h)

Overall estimated mass of the vehicle = 140 kg

Maximum tire pressure = 85 psi (5.8605 bar)

Rear wheel diameter = 0.508 m

Torque required to move the vehicle = Rolling resistance x driving wheel radius

The efficiency of the drive train was approximated as 90% [28] and the overall gear ratio was achieved in two steps.

First step from engine to clutch with a speed ratio of 2.73:1and second step from clutch to rear wheel sprocket with a speed ratio of 5.85:1.

(Torque calculations in table 6.4 in appendix D)

Figure 12.0: Torque at engine, clutch and rear wheel for a range of engine speeds (RPM)

Torque values from the engine performance curve were used to calculate torque at the clutch and rear sprocket using the gear ratio and efficiency. (Figure 2.0) [11]

7.4.2 Vehicle speed calculation

At 3000 RPM,

The rear wheel speed = 3000 RPM/16 = 187.5 RPM

Therefore, since the rear wheel radius = 0.254 m

This is equal to 17.954 km/h

The engine idles at 3000 RPM [29]. Also, at 3000 RPM the torque at the rear wheel is approximately 14 Nm (Refer figure 3.0) which is greater than 2.339 Nm required for initial movement.

Therefore, the vehicle would travel at a speed of 17.954 km/h at idle RPM.

The calculation shown above was repeated for engine speeds from 3000 to 8600 RPM and a graph was plotted to observe how vehicle speed varies with RPM. (Figure 12.1)

(Results in table 6.5 in appendix D)

Figure 12.1: Vehicle speed versus RPM

The graph in figure 12.1 can be used to determine the speed of the vehicle at any given engine speed between 3000 and 8600 RPM.

The minimum average speed requirement to complete the required distance for the prototype category is 30 km/h [30]. When the average speed is lower, fuel consumption would also be lower. Therefore driving strategy would be to travel at a constant speed of 31 km/h except during the start and end of the race.

To travel at 31 km/h, the engine speed should be constant at approximately 5180 RPM. (Figure 12.1) The fuel consumption of the engine increases significantly at engine speeds above 6000 RPM (Figure 2.0). The driver is advised to keep the speed of the vehicle below 35 km/h at all times.

7.5 Clutch

A clutch was required in order to immobilize the vehicle in the starting line. [1]

There were few potential clutch concepts.

Centrifugal clutch [31]

Friction plate clutch [2]

Manual clutch [32]

The stock Honda GX35 consists of a centrifugal clutch. This engages at speeds above 3000 RPM which shows that this is not suitable for our purpose.

Friction plate clutch is a simple system that has been used in previous vehicles. [2]

Considering, the availability of information and simplicity of engagement mechanism a manual scooter clutch was selected for the system. This clutch can be assembled easily and the mechanism would be the same as in the scooter.

7.6 Starter system assembly

The stock GX35 is a pull/kick start engine. This does not match the driving strategy and competition rules because it would be difficult for the driver to start/stop the engine while inside the vehicle.

An electrical starter system would enable the driver to switch off the engine and coast after attaining a particular speed and later restart the engine when the speed decreases.

An electric starter motor was selected and a sub assembly was designed. [33]

The starter motor was connected to the drive shaft using a chain and sprocket. A freewheel clutch was included in between sprocket and shaft in order to restrict free rotation of the sprocket only in one direction. [34] (Appendix E)

The driver is advised to disengage the clutch during starter motor operation because, the competition rules state that the starter motor is not allowed to provide any forward propulsion force. [1]

7.7 Other drive train components and basic stress analysis

Stock parts found in the market were specified for sprockets, chains, bearings etc. The engine-shaft coupling was custom designed to suit the requirements.

7.7.1 Engine shaft coupling

The coupling used to connect the flywheel to the drive shaft was designed. This rotates at high speed (Engine RPM) but at low torque. Therefore this undergoes relatively low stress.

The stress analysis was carried out in Pro Engineer Mechanica using the highest engine torque output with a safety factor of two.

Figure 3.0Shows few stress concentrated areas close to the shaft key.

The material specified for this component was Aluminium alloy 5086. [35]

Figure 13.17.7.2 Rear 76 tooth sprocket

This is a low speed and high torque part. Since this sprocket is connected to the driving wheel (i.e. last stage of the overall gear ratio), this undergoes the highest torque

The highest torque was used to analyze the stress on this component. As this is a component selected from the market it can was verified that this component is suitable for this application

7.7.3 The system holder/carrier

Figure 13.2Although this is not a part in the drive train, this component carries the total load of the propulsion system including the engine.

The extrusions shown on the component were made in order to help an intake of air into the engine compartment from below for engine air intake, cooling and also for aerodynamic air channeling purposes.

Aluminium alloy 5086 was specified also for this component. [35]

(CAD drawings for these parts can be found in appendix D)

CHAPTER 8.0

ELECTRICAL SYSTEM

8.1 Introduction

The basic electrical circuit to power the electric starter motor, horn, light and emergency shutdown mechanism was designed and it was approved by competition organizers.

8.2 Electrical system layout

Kill Switch 1

Kill Switch 2

Kill Switch 3

Solenoid fuel cut off valve

Kill Switch Relay

Starter Switch Relay

Starter Motor

Light

Starter motor switch

Horn switch

Horn

Figure 14.0

The Shell Eco-Marathon rules state that an indicator light is required to indicate any operation of the starter motor [1]. Therefore a light was connected parallel to starter motor to be triggered using the starter relay switch.

Three kill switches are present to isolate the battery and stop the engine for safety purposes which is a mandatory requirement in the competition. [1]

One kill switch should be accessible to the driver and the other switches should be easily accessible from the exterior. [1]

A solenoid fuel cut off valve is used to stop the engine. The solenoid is set to be in energized state when the engine is running. [36]

8.3 Verification calculations

Battery: 12V, 12Ah

Starter motor: 12 V DC, 25 A

Power = 12 V x 25 A = 300 W,

Energy required to operate starter for 0.25 hours = 300 x 0.25 = 75 Wh

Horn: 12 V, 4 A

Power = 48 W, Energy required for 0.25 hours = 48 x 0.25 = 12 Wh

Energy required for the solenoid fuel cutoff valve and indicator light is negligible.

Fully charged battery contains = 12 V x 12Ah = 144 Wh

Therefore this battery could store enough and more energy to power the vehicle electrical components for the duration of the race.

CHAPTER 9.0

PERFORMANCE ESTIMATION

9.1 Introduction

The inputs from other team members in charge of each sub assembly were obtained to carry out the final miles per gallon estimation considering aerodynamic drag, rolling resistance and drive train power loss.

9.2 Calculations

Coefficient of drag (CD) = 0.16

Body frontal area (A) = 0.855 m2 (9.2 sq feet)

Tire Pressure (Ptire) = 85 psi (5.8605 bar)

Engine brake specific fuel consumption, BSFC = 0.128 gal/hp-h (6.498x10-4 l/Wh)

Efficiency of drive train = 90%

Drive train power loss, PDL at 5200 RPM = 0.08 kW (Refer figure 3.0)

Vehicle travels at 31 km/h (19.26 MPH) for the main duration of the race

Power lost due to aerodynamic drag

Density of air = 1.164 kg/m3 at 30oC [38]

Power lost due to rolling resistance

Fuel consumption

Kilometers per liter

Therefore according to calculations, the car could travel 233.26 kilometers per liter. This is equal to 548.88 Miles per gallon (US).

The inputs were also entered to the Bowling and Grippo online automotive calculations program which computed an estimate of 472.81 miles per gallon (US). [7]

The increase in volumetric efficiency was not considered for this calculation, thus the vehicle may travel more distance than the value calculated.

Gasoline equivalent

[1]

[15]

[1]

[15]

The vehicle could travel 233.26 km using 21224.1 kJ of energy

Therefore with 31617.3 kJ:

This would be the value recorded for the competition ranking.

CHAPTER 10.0

PARTS LIST AND QUOTATION

10.1 Introduction

This section of the report contains the list of parts selected and designed along with the estimated price/cost of the components.

10.2 Quotation

Component Name

Material

Manufacturer/ Supplier

Item No.

Quantity

Price/ USD

Honda GX35 Engine

------------

Honda power equipment

GX35

1

250

15 tooth sprocket, 1/2" pitch

Steel

Renold

08B1/15T

2

24

47 tooth sprocket, 1/2" pitch

Steel

Reid Supply Company

IDC-40B47F34

1

35

76 tooth sprocket, 1/2" pitch

Steel

Renold

08B1/76T

1

45

110cc Manual Clutch

------

Parts for Scooters

117-5

1

40

Drive shafts

Steel

Linear bearing Pty Ltd

HRC 60-64

2

50

Transmission chains

Steel

Renold

GY08B1

4 m

60

Ball Bearings

------

SKF

6203-RSH

3

21

Sprag clutch (Freewheel)

------

Renold

SA04

1

20

Engine/shaft coupling

Al alloy 5086

Custom made

---------

1

5

Starter Motor

Zhongshan Ruiying Motor & Electric Co. Ltd

RY7917

1

70

Battery

--------

Yuasa

YTX14-BS

1

95

System Carrier/holder

Al alloy 5086

Custom design

---------

1

10

1.5 L Soda bottle

Plastic

Coca-Cola

---------

1

0.5

Shell Fuel tank

--------

Shell

250 ml

1

----

Pipes/hoses for air and fuel 3/16" (5mm) I.D

--------

Sudco fuel system parts

009-040

2 m

10

Valve stem

--------

Ningbo Haofeng Auto parts Co. Ltd

Bicycle valve stem

1

3

Pressure Control Valve

--------

Pneumadyne

Control valve

1

6

Fuel system holder

Al alloy 5086

Custom design

----------

1

3

Pressure gauge

-------

Cooling systems sdn.bhd

ESOP

1

8

Intake, exhaust pipes and velocity stack

-------

Custom design

-------

---

10

Fuel cutoff valve

-------

Vehicle wiring products Ltd

12912

1

70

Bolts and nuts M8x2, M16x2, M12x2

-------

----------

-------

12

4

Table 5.0: Parts list and quotation

(Available specification sheets can be found in appendix E)

10.3 Recommendations

The parts selected for the system are from various manufacturers and suppliers around the world. This was mainly due to the availability of information. If there is a problem in obtaining a particular component during manufacture, it is recommended to use a different part only if it does not cause a change to the system configuration.

CHAPTER 11.0

CONCLUSION

11.1 Engine, ethanol combustion and efficiency

An internal combustion engine was chosen as the power plant and ethanol E100 was selected as the fuel.

A Honda GX35 (4 stroke, Single cylinder, air cooled) engine was selected after carrying out a parametric study on few similar engines.

11.1.1 Compression ratio

The compression ratio of the engine could be increased due to the anti knock feature in ethanol due to the high octane number. From calculation results and literature it is recommended to increase compression ratio to 12:1. This would decrease the brake specific fuel consumption value significantly, which would increase the estimated miles per gallon value by 55 miles. However, this increase in compression ratio results in an increase in cylinder pressure by a factor of 1.64 (according to calculations). The manufacturer was contacted in order to verify that this engine could withstand this increase in pressure and temperature but a reply has not been received to date.

11.1.2 Intake and Exhaust optimization

Optimum Intake and exhaust pipe lengths were calculated using valve timing and pipes were designed also considering engine compartment space. A velocity stack was also designed to be assembled to the end of the intake pipe to further increase volumetric efficiency.

Intake pipe length = 0.5 m

Exhaust pipe length = 1.5 m

11.2 Fuel delivery system

A pressurized fuel delivery system is used where a pressurized air storage bottle connected to the fuel tank provides the required pressure to push the fuel. This system has less weight and no parasitic load on the engine output. Thus this method does not affect the brake specific fuel consumption and miles per gallon.

11.3 Drive train

The drive train design consists of a direct drive transmission system with a manual scooter clutch. The gear ratio was fixed at 16:1. The torque required to move the vehicle from rest is 2.339 Nm, which is much less than the torque at the rear wheel even at idling speed (Figure 12.0). Minimum average speed to complete the race is 30 km/h [30]. At 5200 RPM, the vehicle could travel at 31 km/h (Figure 12.1). Therefore, this confirms that the fixed gear ratio of 16:1 is well suited for this application.

11.4 Electrical system

Basic electrical system layout was drawn (Figure 14.0) showing the starting system, emergency shutdown mechanism and other accessories required for the vehicle.

Three kill switches were specified, one accessible for the driver and two for exterior access. These switches will shut off the engine and isolate the battery for safety purposes.

11.5 Modeling and simulation

The overall system was modeled using Pro Engineer and detailed drawings were obtained. Parts under high loads were analyzed using Pro Engineer Mechanica.

11.6 Entry pack and approval

Technical drawings of the fuel delivery system, electrical system and other design features were approved by Shell Eco-Marathon Asia 2010 organizers.

11.7 Parts quotation

All required parts for the propulsion system were selected and a quotation was prepared

11.8 Sound level

Since, the system is not yet manufactured the sound level of the system could not be found. The Honda GX35 is a quiet engine according to reviews but enough space was allocated in the system for a suitable muffler if the sound level exceeds the limit. [1]

11.9 Final estimation of vehicle performance

The BMEP of the engine was increased from 67.44 to by modification.

The final performance calculations showed that the vehicle could travel 233.26 kilometers per liter of ethanol E100. The vehicle would be ranked according to the gasoline equivalent in the competition. [1]

Therefore, gasoline equivalent value for 233.26 km/l with ethanol is 347.48 km/l which is 817.32 miles per gallon (US)

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