41802098 sleeve valve report

Cranfield University

Adriaan Moolman

Modelling of a 4-Stroke Sleeve

Valve Engine

School of Engineering

MSc Automotive Product Engineering

Cranfield University

School of Engineering

MSc Automotive Product Engineering

Academic Year 2006-2007

ADRIAAN MOOLMAN

Modelling of a 4-Stroke Sleeve Valve

Engine

Supervisor: Professor Douglas Greenhalgh

August 2007

This thesis is submitted in partial fulfilment of the requirements for the degree of

Masters in Science

Cranfield University 2007. All rights reserved. No parts of this publication may be

reproduced without the written permission of the copyright owner.

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Modelling of a 4-Stroke Sleeve Valve Engine August 2007

ABSTRACT

In the highly competitive automotive industry where ever increasing demand on

higher performance is overshadowed by emission regulations, downsizing engines

becomes an attractive solution. To ensure sufficient breathing capacity of the

downsized engine, the higher possible valve area of the sleeve valve coupled with the

possibility to optimize the combustion chamber and the reduced mechanical losses

present a plausible alternative to poppet valve engines.

The aim of this study is to develop a simulation model in order to predict the

performance of a sleeve valve engine. Little theoretical or empirical models are

available for sleeve valve engines because the use of sleeve valve engines deteriorated

before the widespread use of computer simulations. The major focus for the

simulation is on the modelling of the flow through the sleeve valves. The modelling

consists of the exact valve areas and the accompanying valve discharge coefficients.

The study subsequently developed a method of determining the valve areas as a

function of the engine crank angle from the arbitrary shaped valve profiles. It also

identified experimental discharge coefficients in the open literature that could be used

for flow analyses and it determined a new set of discharge coefficients by way of CFD

simulations. These CFD derived discharge coefficients compared well with the

experimental coefficients and can subsequently also be used for sleeve valve

modelling.

WAVE models were developed for a sleeve valve engine using the sleeve valve models

as determined in the study. These WAVE models produced satisfactory results,

reiterating the need for accurate valve models.

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ACKNOWLEDGEMENTS

Firstly I thank my Lord and Saviour Jesus Christ for the opportunity He gave me to

study this course and for the abilities and intellect to complete this study.

I thank my parents for all their support and love, emotionally and financially and I

thank my brother and sister for their support and love as well.

My thanks go to my supervisor for his help and guidance during this study as well as

my fellow students working with me on the sleeve valve project for their support and

help. I thank Mahle for providing the experimental engine as well as help and

assistance regarding this project.

I thank my flatmates and my classmates who helped me through this year of study and

for helping me make this a very memorable year in my life. I also thank the staff of the

Automotive Product Engineering course for their teachings and guidance throughout

the year.

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TABLE OF CONTENTS

Abstract ....................................................................................................................... iii

Acknowledgements...................................................................................................... iv

Table of Contents .......................................................................................................... v

List of Figures .............................................................................................................. vii

Notation ....................................................................................................................... x

1. Introduction .......................................................................................................... 1

2. Literature Review .................................................................................................. 4

2.1 Engine Downsizing .......................................................................................... 4 2.2 The Use of Sleeve Valve Engines ..................................................................... 7

2.2.1 Brief History of Sleeve Valve Engines ....................................................... 7 2.2.2 Sleeve Valve Operation ............................................................................ 7 2.2.3 Advantages of Sleeve Valves .................................................................. 10 2.2.4 Disadvantages of Sleeve Valves .............................................................. 12

2.3 Engine Modelling .......................................................................................... 13 2.3.1 Sleeve Valve Flow Coefficients ............................................................... 14 2.3.2 Sleeve Valve Area .................................................................................. 20 2.3.3 Heat Transfer in Small Engines ............................................................... 21

2.4 Conclusion .................................................................................................... 22

3. Initial WAVE Model ............................................................................................. 23

3.1 Determining the Port Positions ..................................................................... 23 3.2 Determining the Valve Areas......................................................................... 28

3.2.1 Initial Method of Calculation .................................................................. 28 3.2.2 Automated Method of Calculation ......................................................... 30

3.3 Valve Models ................................................................................................ 35 3.4 Intake Flow Path ........................................................................................... 38

3.4.1 Geometry .............................................................................................. 38 3.4.2 Heat Transfer ......................................................................................... 41 3.4.3 Junction ................................................................................................. 45

3.5 Engine Model ................................................................................................ 46 3.5.1 Engine Geometries ................................................................................ 46 3.5.2 Combustion Model ................................................................................ 50 3.5.3 Engine Heat Transfer ............................................................................. 50

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3.6 Exhaust Flow Path ......................................................................................... 57 3.6.1 Ducts ..................................................................................................... 57 3.6.2 Junction ................................................................................................. 59

3.7 Initial Results and Discussion ........................................................................ 60 3.8 Conclusion .................................................................................................... 65

4. Valve Discharge Coefficients with Computational Fluid Dynamics ....................... 67

4.1 Model Generation......................................................................................... 67 4.1.1 Model Layout ......................................................................................... 68 4.1.2 Valve Geometry ..................................................................................... 69 4.1.3 Gambit Models ...................................................................................... 70 4.1.4 Meshing ................................................................................................. 71

4.2 Simulation Specifications .............................................................................. 72 4.2.1 Solver Models ........................................................................................ 72 4.2.2 Boundary Conditions ............................................................................. 74 4.2.3 Convergence .......................................................................................... 75

4.3 Post Processing ............................................................................................. 76 4.4 Results and Discussion .................................................................................. 79 4.5 Conclusion .................................................................................................... 83

5. Experimental Facility ........................................................................................... 85

5.1 Assembly of Test Setup ................................................................................. 85 5.1.1 Belt Driven ............................................................................................. 86 5.1.2 Direct Coupling ...................................................................................... 87

5.2 Conclusion .................................................................................................... 89

6. Final WAVE Model ............................................................................................... 91

6.1 Changes from Initial Model ........................................................................... 91 6.2 Equivalent Poppet Valve Model .................................................................... 92 6.3 Results and Discussion .................................................................................. 94

6.3.1 Initial Model vs. Updated Model ............................................................ 94 6.3.2 Sleeve Valve Model vs. Poppet Valve Model .......................................... 96

6.4 Conclusion .................................................................................................... 99

7. Final Conclusion and Further Work.................................................................... 101

7.1 Conclusion .................................................................................................. 101 7.2 Recommendations for Further Work........................................................... 102

8. References ........................................................................................................ 104

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LIST OF FIGURES

Figure 1: VSC Core Engine (Hendrickson, 1999) ........................................................... 6

Figure 2: Sleeve Valve Motion ..................................................................................... 8

Figure 3: Various Sleeve Port Arrangements (Ricardo, 1931) ....................................... 9

Figure 4: Maximum Available Valve Areas (Ricardo, 1931) .......................................... 9

Figure 5: Cylinder Junk Head (Dardalis, 2004) ......................................................... 10

Figure 6: Typical Valve Flow Coefficient for Poppet Valves (Cole, 2006)..................... 15

Figure 7: Shape of Sleeve Valve Openings (Waldron, 1940) ....................................... 15

Figure 8: Cylinder of Waldron Experimental Engine (Waldron, 1940) ........................ 16

Figure 9: Flow Coefficients for Centre Inlet Valve (Waldron, 1940) ............................ 17

Figure 10: Flow Coefficient for Centre Valve at Different Openings (Waldron, 1940) . 18

Figure 11: Flow Coefficient for End Inlet Ports (Waldron, 1940) ................................ 18

Figure 12: Manifold Pressure with All Inlet Ports Open (Waldron, 1940) ................... 19

Figure 13: Flow Coefficient for Exhaust Valves (Waldron, 1940) ................................ 20

Figure 14: Valve Movement with Respect to Crank Angle (Hendrickson, 1999) ......... 21

Figure 15: Traced Sleeve Ports................................................................................... 24

Figure 16: Traced Cylinder Wall Ports ........................................................................ 24

Figure 17: Coordinate Points on Sleeve Port Profiles ................................................. 25

Figure 18: Coordinate Points on Cylinder Wall Port Profiles....................................... 25

Figure 19: Port Layout at 0 Crank Angle ................................................................... 26

Figure 20: Elliptical Motion of Sleeve ......................................................................... 26

Figure 21: X and Y Coordinates of Ellipse at Crank Angle ........................................ 27

Figure 22: Curves Fitted to Points Describing Sleeve Port .......................................... 29

Figure 23: Points Describing the Sleeve Port Profile................................................... 31

Figure 24: Piston Movement with Crank Angle .......................................................... 33

Figure 25: Trapezoid from Adjacent Valve Opening Points ........................................ 34

Figure 26: Valve Areas Plotted Against Crank Angle ................................................... 35

Figure 27: Typical Input Page for Effective Valve Area ............................................... 36

Figure 28: Discharge Coefficients as taken from (Waldron, 1940) .............................. 37

Figure 29: Input Page for Valve Discharge Coefficient ................................................ 37

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Figure 30: WAVE Layout of Intake Flow Path ............................................................. 38

Figure 31: Inlet Manifold Duct and Cylinder Barrel .................................................... 39

Figure 32: Inlet Manifold Geometry .......................................................................... 39

Figure 33: WAVE Layout of Intake Including Heat Transfer Ducts .............................. 44

Figure 34: 3-D Layout of Y-Junction Element for Intake Manifold .............................. 46

Figure 35: Volumes for Compression Ratio Calculation .............................................. 48

Figure 36: Fin Geometry and Equations ..................................................................... 54

Figure 37: Efficiency of a Rectangular Annular Fin (Incropera & De Witt, 1996) ......... 56

Figure 38: Schematic of Exhaust Flow Path ................................................................ 58

Figure 39: 3-D Layout of Y-Junction Element for Exhaust Pipe ................................... 60

Figure 40: Brake Power and Torque Calculated with Initial Model ............................. 61

Figure 41: Volumetric and Thermal Efficiency Calculated with Initial Model .............. 61

Figure 42: Indicated and Brake Mean Effective Pressure Calculated with Initial Model

................................................................................................................................... 62

Figure 43: P-V Diagram Calculated with Initial Model at 4000 rpm ............................ 62

Figure 44: Effective Valve Areas for Initial Model ...................................................... 63

Figure 45: Mass Flows through Valves Calculated with Initial Model ......................... 64

Figure 46: Pressure Difference across the Valves ....................................................... 65

Figure 47: Layout of CFD Model................................................................................. 68

Figure 48: Indication of Valve Opening Profiles Simulated ......................................... 70

Figure 49: Solver and Viscous Model Input Pages of Fluent ....................................... 73

Figure 50: Centre Inlet Valve Discharge Coefficients .................................................. 79

Figure 51: End Inlet Valve 1 Discharge Coefficients.................................................... 80

Figure 52: End Inlet Valve 2 Discharge Coefficients.................................................... 81

Figure 53: (Waldron, 1940) End Inlets (left) vs. Experimental Engine End Inlets (right)

................................................................................................................................... 81

Figure 54: Exhaust Valve 1 Discharge Coefficients ..................................................... 82

Figure 55: Exhaust Valve 2 Discharge Coefficients ..................................................... 83

Figure 56: Experimental 4-Stroke Sleeve Valve Engine............................................... 85

Figure 57: Engine Belt and Pulley Layout ................................................................... 87

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Figure 58: Engine CV Joint Layout .............................................................................. 88

Figure 59: Engine Mountings ..................................................................................... 89

Figure 60: Updated Discharge Coefficient Input Page for Inlet Valve ......................... 91

Figure 61: Updated Discharge Coefficient Input Page for Exhaust Valve .................... 92

Figure 62: Valve Configuration for Poppet Valve Model ............................................ 93

Figure 63: Brake Power and Torque Calculated with Updated Model ........................ 94

Figure 64: Effective Valve Areas Updated Model Left & Initial Model Right ............ 95

Figure 65: Valve Mass Flow Rates Updated Model Left & Initial Model Right .......... 95

Figure 66: Brake Power and Torque Calculated with Poppet Valve Model ................. 96

Figure 67: Valve Effective Areas Sleeve Valve Model Left & Poppet Valve Model

Right ........................................................................................................................... 97

Figure 68: Valve Mass Flow Rates Sleeve Valve Model Left & Poppet Valve Model

Right ........................................................................................................................... 97

Figure 69: Brake Power of Sleeve and Poppet Valve Models ..................................... 99

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NOTATION

Variables

Variable Description

Speed of sound

Constants

Area

Flow coefficient

Discharge coefficient

Specific heat at constant pressure

Compression ratio

Diameter

Hydraulic diameter

Discretization length

Heat transfer coefficient

Height

Thermal conductivity

Connecting rod length

Mass flow rate

Length, lift

Number of fins

Nusselt number

Static pressure

Total pressure

Prandtl number

Wetted perimeter

Crank shaft radius, radius

Universal gas constant

Reynolds number

Surface area, fin and adjacent wall thickness

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Piston vertical position

Time, thickness

Temperature

Volume, velocity

Greek Symbols

Variable Description

Coordinates

Crank angle

Emissivity

Efficiency

Ratio of specific heats

Viscosity

Density

Sleeve angle

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1. INTRODUCTION

The automotive industry is a very contradictive industry in that the research and

development is driven by two conflicting factors. It is dictated by regulations and

legislations set out by governments, which at this point in time focuses on

environmentally friendly and safety driven vehicles. Direct consequences of these

focuses are vehicles with lower performance in order to emit less harmful exhaust

gasses and slower vehicles in order to be safer. However, the automotive industry is

dependent on its customers to survive financially and the customers desire faster

vehicles with ever increasing performance. It is therefore the task of the automotive

engineer to satisfy the customers while adhering to the regulations and legislations.

The reduction of carbon dioxide and other harmful exhaust gas emissions are very

important issues and consume vast amounts of research and development resources.

Various techniques are investigated and employed, and one of the techniques

currently being developed is downsizing. This consists of decreasing the engine

displacement in order to reduce the exhaust gas emissions. It is, however important to

maintain satisfactory performance and therefore boosting is usually employed with

downsizing.

Decreasing the engine displacement involves reducing the piston bore and stroke.

When reducing the piston bore, the diameter of the conventional poppet valves

subsequently also reduces, resulting in smaller air flow area and increased pumping

losses due to increasing friction of the flow and the surrounding surfaces. Decreasing

the air flow into the engine will reduce the amount of fuel that can be burnt per cycle,

thus, together with the increased pumping losses, reducing the engine performance.

One possible way to counter this problem is by using sleeve valves to facilitate the air

induction and exhaust gasses of the engine. However, sleeve valve engine design have

not enjoyed as much research and development as the poppet valve engine designs

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and problems with high harmful exhaust emissions, unwanted sleeve friction and

ineffective sealing still needs to be resolved before sleeve valve engines can be used

productively. The majority of the sleeve valve development occurred before the

1950s, therefore before the widespread use of computer simulation software to

determine engine performance and optimize designs. The present study thus is aimed

at developing models for simulating sleeve valve engine using present engine

simulation software and an experimental 4-stroke sleeve valve engine. Attention will

also be paid to develop these models so that it could be utilized in simulation of

downsized engines employing sleeve valve engines. The software that will be used is

called WAVE. It is a 1-dimensional engine simulation package developed by Ricardo.

From the onset of the project the importance of accurately determining the sleeve

valve area was realised. Accompanying the sleeve valve areas are the discharge

coefficients that combine to produce the effective area of the valves. A major focus of

this report was to determine these two valve characteristics for an experimental sleeve

valve engine provided by Mahle for this study. A method of calculating the valve areas

from traced drawings of the port profiles are presented. Valve discharge coefficient

from available literature is presented as well as a set of simulated discharge

coefficients specifically characteristic to the valves of the experimental engine. These

coefficients were simulated using computational fluid dynamic (CFD) software.

It was planned to perform experiments with the engine and to use the experimental

results to calibrate the WAVE engine models. However, due to unforeseen

circumstances and the time constraint on this project, the experimental results did not

materialise, but still a chapter was dedicated to explaining the experimental setup and

lessons learned during the attempts to acquire these results.

Three WAVE models were developed and the results compared in order to gain

understanding into simulating sleeve valve engines. One model was done with

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discharge coefficients found in the open literature, one model with CFD derived

discharge coefficients and one model with poppet valves.

Finally conclusions were drawn and further recommended work discussed. This

project served as one in four projects performed on the particular sleeve valve engine.

The other projects address different parts of the engine and although the projects

were all separate, some information and knowledge were shared. The other projects

are (Chabert, 2007), (Franco Sumariva, 2007) and (Vasudevan, 2007).

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2. LITERATURE REVIEW

In order to gain a better understanding of sleeve valve engines, a literature review was

undertaken. It also serves as a tool in performing the project and identifies previous

work done in the relevant fields so that unnecessary duplication of work will be

avoided. The literature review firstly focuses on the topic of engine downsizing after

which the focus is shifted towards sleeve valve engines and then the modelling of this

type of engines. A few key modelling issues are identified and existing literature

assembled to aid in the understanding and completion of the task at hand.

2.1 ENGINE DOWNSIZING

One of the possible methods of reducing engine exhaust emissions while maintaining

sufficient performance is by downsizing the engine. The problem with current

production small engines is that they are not designed to meet any emission

regulations and fuel consumption is of low importance. These two factors, however,

are major design criteria for modern automotive engines.

Small engines show the tendency to produce low brake thermal efficiencies and (Lowi,

2003) describes a few causes for this. When downsizing an engine the surface to

volume ratio becomes an important design consideration. The smaller cylinder exhibit

higher heat transfer areas which could result in over cooling thereby impairing

effective combustion, but the cylinder head has the tendency to under-cool resulting

in excessive spark plug temperatures. The cooler cylinder walls do however reduce the

tendency for end gasses to auto ignite, allowing the use of higher compression ratios.

Downscaling of the cylinder results in viscous effects influencing the air stream and

causing small scale turbulence. This causes insufficient air/fuel mixture and flame

speeds which can be resolved by introducing large flow areas into the cylinder. This is

however difficult to achieve with conventional poppet valves, thus promoting the use

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of sleeve valves. (Yagi et al. 1970) also states that one of the major design

considerations to achieve high volumetric efficiency is to maximize the valve area in

order to increase engine breathing.

Furthermore, using carburetion with a short inlet manifold will cause incomplete

vaporization of the fuel (especially wide-boiling hydrocarbon fuels) resulting in

unburned fuel being passed through the engine causing high fuel consumption and

hydrocarbon emissions. Port- or direct fuel injection might solve this when high

atomization injection is used. These factors must be taken into account when

simulating and designing a downsized engine.

The design of the combustion chamber is one of the most important components in

designing a small engine. A high compression ratio and combustion speed is required

in order to maximize the thermal efficiency while flame travel and heat transfer must

be minimized so that higher indicated efficiency can be reached. Decreasing the travel

that the flame must undergo to engulf the end gasses will result in a higher usable

compression ratio. In order to ensure sufficient turbulence in the air flow into the

cylinder, the combustion chamber design in a small engine needs to promote swirl

motion of the air. Careful consideration is required not to invoke excessive turbulence

so that the flame kernel is extinguished before the fuel is burnt completely.

(Lowi, 2003) describes the design considerations for a combustion chamber of a small

cylinder engine and concluded that the design used by (Hendrickson, 1999) is

sufficient. This design consists of a small spherical open chamber with a spark plug

locater centrally with a small squish land on the cylinder perimeter. (Ricardo and

Company, 1947) also confirmed that decreasing the combustion chamber diameter

with the use of a squish land on the cylinder perimeter increases the swirl inside the

combustion chamber. This minimized the volume of the chamber as well as the flame

travel and the surface area. This arrangement however, deems it improbable to use

poppet valves and (Hendrickson, 1999) also describes using sleeve valves to overcome

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the lack of area in the combustion chamber to yield to the poppet valves. Figure 1

illustrates the so called VSC core engine by (Hendrickson, 1999). Note the

combustion chamber shape as described above and the lack of space for poppet valves

in the combustion chamber, necessitating the use of a sleeve valve.

Figure 1: VSC Core Engine (Hendrickson, 1999)

Turbocharging a downsized engine may lead to impractically small turbomachinery.

Too tiny components would have to run at too high rotational speeds resulting in low

Reynolds numbers which is not practical for manufacture and service. In these cases

positive displacement pumps would result in a more practical solution (Hendrickson,

1999).

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2.2 THE USE OF SLEEVE VALVE ENGINES

2.2.1 Brief History of Sleeve Valve Engines

In 1903 Charles Yale Knight designed the first sleeve valve engine. This sleeve valve

mechanism consisted of a double sleeve arrangement with reciprocating movement.

Six years later two separate designers filled patents for single sleeve valve mechanism

combining reciprocating and rotating movements to produce an elliptical path of valve

movement. These two inventers were Peter Burt and James H K McCollum (Wells).

Various sleeve valve engine designs enjoyed moderate success in the automotive

industry with the high production cost of the engines limiting their use to upmarket

vehicles. Sir Harry Ricardo noticed the sleeve valve engine and realized its potential as

a high performance aero engine. He performed much development work on sleeve

valve engines and many different sleeve valve design aero engines were employed

during the Second World War. Among them the Bristol Centaurus and the Napier

Sabre, two of the worlds most powerful spark ignition engines.

The sleeve valve engine was a very competent alternative to the poppet valve engine,

showing very high levels of performance for spark ignition engines and many other

advantages (as described in the subsequent sections). The advent of the jet engine in

the aero industry however, halted the use of the sleeve valve engine in that industry.

At that stage no other markets existed for very high performance spark ignition

engines and subsequently sleeve valve engines was lost to the world.

2.2.2 Sleeve Valve Operation

The sleeve is located between the cylinder wall and the piston. Port openings at

various locations along the cylinder wall serve as inlet and outlet passages. The sleeve

consists of a number of pie-shaped openings situated along its circumference. These

openings are aligned with the applicable ports in the cylinder wall at the appropriate

sectors in the intake and exhaust strokes, thereby creating inlet and outlet valves

respectively. The sleeve motion is produced by a gear driven cam connecting to the

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sleeve and delivering reciprocating as well as rotating motion to result in an elliptical

path being followed by the sleeve (Figure 2).

Figure 2: Sleeve Valve Motion

Various port arrangements are illustrated in Figure 3, with the subsequent maximum

valve areas available for some of these arrangements at different bore diameter

illustrated in Figure 4.

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Figure 3: Various Sleeve Port Arrangements (Ricardo, 1931)

Figure 4: Maximum Available Valve Areas (Ricardo, 1931)

At TDC the sleeve ports are above the junk head rings (Figure 5), effectively

shrouding the ports from the combustion chamber and protecting the ports from the

combustion gasses. This is however a place of concern when sealing is considered and

blow-by of gasses occur around these rings.

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Figure 5: Cylinder Junk Head (Dardalis, 2004)

2.2.3 Advantages of Sleeve Valves

Sir Harry Ricardo realised the potential of the sleeve valve engine as a high

performance aero engine, but described the following advantages of the general use of

sleeve valve engines (Lowi, 2003):

The spark plug could be located in the centre of the combustion chamber,

thereby minimizing the required flame travel to engulf all the charge in the

combustion chamber. This is also applicable to very small cylinder engines and

is exactly the design consideration required as described in Section 2.1. This

use of sleeve valves which permits the designer to optimize the combustion

chamber shape for desired combustion was also realized by (Hendrickson,

1999) and (Lowi, 2003).

The lack of high temperature resistant materials in the early part of the 20th

century caused problems for exhaust poppet valve design. The use of sleeve

valves eliminated problematic exhaust poppet valves while also eliminating the

source of unwanted auto ignition in the form of the hot exhaust valves. The

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absence of the hot exhaust valves subsequently allows for higher tolerable

compression ratios, thereby increasing engine performance.

The geometry and layout of the sleeve valve generates high levels of natural

turbulence (in the form of swirl) when the valves initially opens aiding in

air/fuel mixing and flame propagation. These high levels of swirl was studied

and documented by (Ricardo and Company, 1947).

The sleeve valve results in a breathing capacity (in other words flow area) at

least equal to that of any accommodated poppet valve arrangement and that

this larger valve area could be opened more rapidly than a poppet valve

counterpart.

The use of a sleeve valve mechanism results in a more compact and less

complex engine with a smaller frontal area.

Sleeve valve engines also showed higher mechanical efficiencies due to

reduced friction and lower actuation force of the valve train. The lower friction

also resulted in less wear of the engine components.

The sleeve valve ensures noiseless operation (Ricardo, 1931).

It is more robust than the poppet valves and requires less attention.

Opposed to the sleeve valve, (Yagi et al. 1970) describes abnormal valve motion of

poppet valve trains as a major obstacle in high speed engines. The rigidity and the

inertia of the valve train is a source of loss in the engine, reducing the volumetric

efficiency. Sleeve valves reduce these mechanical losses due to a lower power

consumption of the valve train.

One of the limitations on engine speed of a normal poppet valve engine as pointed out

by (Lumley, 2001) is valve float. This happens when the engine speed becomes too

high, and the valve spring is not strong enough to prevent the valve from breaking free

from the cam profile. When using a sleeve valve, this limitation in engine speed is

eliminated entirely, because the sleeve valve is operated by a fixed cam and not

controlled by a spring.

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At first it was believed that the extra surface contact areas between the piston and the

sleeve, and the sleeve and the cylinder wall would increase mechanical friction,

thereby reducing engine performance. However, according to (Dardalis, 2004)

experiments have shown that the total friction of the sleeve valve was usually lower

than conventional poppet valve designs, believed to be due to the rotary movement of

the sleeve.

The maintanance records of over 60 000 sleeve valve engines used during the war

suggested the absence of localized cylinder wear paterns, observed in engines without

the resiprocating sleeve valve, and 10 times lower overall bore wear (Dardalis, 2004).

The wear was so low that it did not determine the engine life as was the case in more

conventional engines. Unfortuanately the major manufacturers at the time, Bristol

and Napier, was more conserned about engine performance than cylinder wear (or

lack thereof in this case) and very little effort was spent on quantifying this benefit.

According to (Dardalis, 2004), the sleeve valve engines illustrated high values of BMEP

and the engines could be maintained indefinitely at these peak pressures rather than

only 15 minutes as the poppet valves was limited to.

Sleeve valve engines are relatively insensitive to high exhaust pressures because of the

increased exhaust valve area allowing quick discharge of exhaust gasses through the

exhaust ports. This results in ideal conditions for using a turbocharger with the sleeve

valves.

2.2.4 Disadvantages of Sleeve Valves

Sleeve valve engines were developed at a stage where emission control was absent,

and therefore the current design of these engines will not meet modern emission

regulations. The extra set of ring in the junk head attribute to higher hydrocarbon

emissions by trapping fuel and preventing it from combusting during the combustion

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process. These stationary rings in the junk head also cause sealing problems and

subsequent blow-by is observed. This will cause inaccuracies in engine simulations and

therefore the blow-by must be accounted for in the engine model.

The piston movement can restrict the port area when a short stroke is employed.

Furthermore, the sleeve hampers heat dissipation of the piston to the cooling capacity

of the cylinder wall. This justifies research into sleeve materials that would allow

increased heat transfer from the piston.

Companies like Rolls Royce started developing high performance sleeve valve engines

and experimental ultra-high performance 2-stroke sleeve valve engines for aero

applications. However, the advent of the jet engine in the aero industry halted the

production of these engines as well as further development of sleeve valve engines.

2.3 ENGINE MODELLING

Design refers to a situation where the characteristics of a system must be specified so

that it will enable execution of specific functions at an acceptable level of

performance. Simulation on the other hand generally refers to a situation where the

characteristics of the system are known and models must be set up to predict its

functionality and performance level (Rousseau, 2002).

The goal of this study is to simulate the sleeve valve engine in order to be able to use

the simulations to optimize the design. To do this, known models must be employed

to accurately predict the performance so that effective optimization can be done. The

level of complexity of the simulations will be dictated by the available models for

different simulated sections of the entire engine. The thermal fluid flow through the

engine ducting will for instance be modelled with theoretical models based on

fundamental principles. The flow through the valves can also be modelled with

theoretical principles (approximated with orifice flow), but empirical correlations

determined experimentally should produce more accurate results, as observed by

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(Waldron, 1940). The theory is not always completely understood and in such

scenarios empirical model must be used to acquire accurate results.

The sleeve valve engine will be modelled using the Ricardo WAVE software package.

This software is used by many automotive companies and research institutions

(Farrugia, 2004). It is a 1-dimensional simulation package which combines accurate

general model simulations with improved simulation time compared to 3-dimensional

CFD simulations.

2.3.1 Sleeve Valve Flow Coefficients

The major fundamental difference between the poppet valve and the sleeve valve

engines is the airflow into and out of the cylinder. Therefore, the major difference in

the modelling of these two types of engines will be the modelling of the valves. The

fact that the pressure drop across the valves has a significant influence on the engine

performance deems it necessary to accurately model the flow coefficients across the

valves. Figure 6 illustrates a typical flow coefficient curve for normal poppet valves as

a function of the valve lift used in the valve model of Ricardo WAVE. It is therefore

necessary to acquire a similar flow coefficient curve for sleeve valves.

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Figure 6: Typical Valve Flow Coefficient for Poppet Valves (Cole, 2006)

A first method of obtaining such a flow coefficient curve for a sleeve valve is to search

the open literature. This was done and a 1940 NACA report (Waldron, 1940) was

obtained describing the construction of flow coefficients for sleeve valves. The author

used an experimental setup which employed a very similar sleeve valve arrangement

as the engine being used for the present study. In both cases a single sleeve is used

with an elliptical path consisting of 3 inlet valves and 2 exhaust valves.

Figure 7: Shape of Sleeve Valve Openings (Waldron, 1940)

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Figure 7 illustrates the shape of the sleeve and cylinder ports that was used in the

experimental engine of Waldron. Figure 8 illustrates the cylinder and port

arrangement. The experimental engine of the present study consist of a very similar

setup, with three inlet ports spread across 180 of the cylinder and the two exhaust

ports located in the remaining half of the cylinder wall. The inlet duct is also aligned

with the one centre port after which it branches to the two end ports resulting in the

inlet flow entering the end ports tangentially.

Figure 8: Cylinder of Waldron Experimental Engine (Waldron, 1940)

Waldron describes the experimental setup and methods used in measuring the

pressure drop as well as the assumptions made during the entire process and the

claimed accuracy of the results. He calculates the flow coefficient as a function of the

pressure across the valves and it is presented in the following equations.

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(1)

(2)

Waldrons results are illustrated for the different valves in the following figures.

Figure 9: Flow Coefficients for Centre Inlet Valve (Waldron, 1940)

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Figure 9 illustrates the flow coefficients for the centre inlet valve for different

approaching flow field conditions. It can be seen that the flow coefficients are quite

high (>0.8) and that they are independent of approaching flow field conditions. It

should be noted that Waldron ensured that inlet manifold acoustics did not influence

the results.

Figure 10: Flow Coefficient for Centre Valve at Different Openings (Waldron, 1940)

Figure 10 illustrates the flow coefficients for the centre inlet valve for different valve

openings, showing that the flow coefficients are independent of the valve opening.

Figure 11: Flow Coefficient for End Inlet Ports (Waldron, 1940)

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Figure 11 illustrates the flow coefficients of the end inlet valves. Although they are

lower than that of the centre inlet valve (0.62 0.78), the flow coefficient still seems to

be high.

Figure 12: Manifold Pressure with All Inlet Ports Open (Waldron, 1940)

Figure 12 illustrates the pressure in the inlet manifold just upstream of the respective

valves in the case where all the inlet valves are opened simultaneously. It shows that

when the valves are fully open, the pressure just upstream of the end valves are lower

than that just upstream of the centre valve, indicating a pressure drop as a result of

the flow curvature.

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Figure 13: Flow Coefficient for Exhaust Valves (Waldron, 1940)

Figure 13 illustrates the flow coefficients for the exhaust valves at different valve

openings. It is clear that the flow coefficients are independent of valve opening.

These results seem to be very useful for developing a model for simulating the sleeve

valves. However, careful consideration must be done to ensure that the definition of

the flow coefficients as calculated by Waldron is exactly the same as the definition of

the flow coefficients used to describe the eventual valve model. This process will be

described in a later Section where the valve model will be described in detail.

2.3.2 Sleeve Valve Area

The area of the sleeve valves as a function of the crank angle together with the flow

coefficients described in the previous section is used to calculate the flow through

these valves. There is no exact equation for calculating the area for the sleeve valve

areas and therefore the drawings and physical measurements of the experimental

engine will be used to determine the areas graphically.

Figure 14 illustrates the valve movement presented as flow area with respect to crank

angle for the VSC core engine of (Hendrickson, 1999). It shows the upwards

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movement of the piston covering the ports, resulting in a reduced flow area. This must

be considered when calculating the valve areas of the experimental engine being used

in this study.

Figure 14: Valve Movement with Respect to Crank Angle (Hendrickson, 1999)

2.3.3 Heat Transfer in Small Engines

The increased heat transfer area in small engines causes cooler cylinder walls. This

heat transfer phenomena of the small bore engines can adversely affect the efficiency

and torque and must subsequently be taken into consideration when simulating and

designing engine performance of a downsized engine (Lowi, 2003).

In the design process of a small cylinder sleeve valve engine, (Lowi, 2003) used the

following models that influence the combustion process:

Fuel properties and mixture as well as unburned mixture and residual gas

fractions.

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Geometry of the combustion chamber including spark plug location and surface

to volume ratio.

Heat transfer characteristics which are based on the mean velocity, a measure

of the turbulence and the swirl ratio.

These suggested models will be taken into account when preparing the final models

for the engine simulations and will therefore be described in more detail in subsequent

sections.

2.4 CONCLUSION

In this review, a brief description of the sleeve valve engine was given as well as some

comments on the downsizing of spark ignition engines. It was found that the sleeve

valve engine consists of many advantages and therefore justifies a closer inspection.

The fact that the current designs of sleeve valve engines will not meet the modern

emission regulations, together with the advantages of the sleeve valve engines justifies

research into minimizing the emissions of these engines. It was also shown that sleeve

valve engines present a plausible solution for maintaining sufficient breathing for

downsized engines.

With this in mind and the lack of sleeve valve simulation models due to the halted use

of these engines in the non-computer age necessitates the need for accurate

performance prediction models to aid in sleeve valve engine optimization. The major

simulation difference between poppet and sleeve valve engines will be the valve flow

models. Sleeve valve flow coefficients for a very similar sleeve valve was found and

described, but the detail description of the valve model will be described in further

sections of this report. For these models the valve areas of the experimental engine

must be determined and the flow coefficient described in this literature review must

be adapted to serve in the WAVE software.

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3. INITIAL WAVE MODEL

The main aim of this project is to model the 4-stroke sleeve valve engine. This will be

done with specialised engine simulation software called WAVE, being developed by a

company called Ricardo. The software is a 1-dimensional fluid simulation package,

which uses model elements to represent certain parts of typical engine components.

The detail theory behind the models will not be addressed as it comprises mostly of

widely published thermal fluid mechanics.

An initial engine simulation was needed in order to use the experimental data to

calibrate the model. In order to develop an initial WAVE model, various geometries

were needed from the engine. As the engine was available for testing, the engine was

taken apart before any testing was done, to acquire the required geometrical

dimensions. The most important geometries needed for the WAVE model is any

geometrical dimensions determining the flow path of air and exhaust gas through the

engine. The sleeve valve port openings are very important geometries and special care

was taken to acquire these values because of their rather arbitrary and complex

shapes.

This chapter explains the determination of acquiring the sleeve valve flow areas as well

as initial sleeve valve flow coefficients and the subsequent development of an initial

WAVE model.

3.1 DETERMINING THE PORT POSITIONS

With the engine taken apart, the ports in the sleeve as well as the ports on the inside

of the cylinder wall were exposed. There are five ports in the cylinder wall, being one

centre inlet port, two end inlet ports and two exhaust ports. The sleeve has four ports,

as two ports overlap the inlet cylinder wall ports; one overlaps an exhaust cylinder wall

port and the final sleeve port overlapping an inlet and exhaust cylinder wall port.

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The shape of the both sets of ports was captured by fixing a sheet of paper around the

sleeve and around the inside of the cylinder respectively and tracing the particular port

shapes with a pencil. Great attention was paid to obtaining accurate copies of the

shapes and two copies of both sets of ports were made and compared in order to

ensure repeatability of the copying process. Both copies produced the same port

profiles and it was therefore assumed to be sufficiently accurate and repeatable.

Scaled down pictures of the traced sleeve ports and of the cylinder ports are presented

in Figure 15 and Figure 16 respectively.

Figure 15: Traced Sleeve Ports

Figure 16: Traced Cylinder Wall Ports

The next step was to copy these images onto graphical paper in order to determine

coordinates for various points on the profiles of the ports. The profiles were copied

onto the graphical paper and many points along the ports shapes were identified so

that the coordinates of these points would describe the respective port shapes.

Figure 17 and Figure 18 illustrates these coordinate points for the sleeve and cylinder

wall port profiles respectively.

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Figure 17: Coordinate Points on Sleeve Port Profiles

Figure 18: Coordinate Points on Cylinder Wall Port Profiles

The X and Y coordinates of all the various points as they occur on the graphical paper

were read into and Excel spreadsheet and X and Y offset values were added in order to

replicate the positions of the port openings at top dead centre (TDC) for the start of

the combustion stroke (assumed as 0 crank/cycle angle). The origin of the Y-axis was

selected to be the outer rim of the piston at bottom dead centre (BDC) and the origin

of the X-axis was selected to be between the centre inlet wall port and one of the end

inlet wall ports. This position was marked on the traced drawings of the sleeve and

cylinder wall ports in order to obtain the correct X offset values. The circumference of

the sleeve outside diameter and the cylinder inside diameter were rolled out on the

X-axis, and therefore the X-axis stretched from 0 mm to approximately 278 mm (sleeve

outside diameter 89 mm). Figure 19 illustrates the positioning of the various ports at

0 crank angle as reproduced in the Excel workbook. Note the horizontal line

representing the piston at TDC.

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Figure 19: Port Layout at 0 Crank Angle

The procedure described above produced the port positioning of all the ports at TDC.

To determine the port positions at any given crank angle, the port coordinates at TDC

was used as the base coordinates whereby dynamic X and Y offset values would be

added for a certain crank angle. These offset values are determined by the sleeve

motion produced by the rotation of the crank shaft.

Figure 20: Elliptical Motion of Sleeve

(3)

60

80

100

120

140

160

180

0 50 100 150 200 250

Wall Port 1 (End Inlet) Wall Port 2 (Exhaust) Wall Port 3 (Exhaust) Wall Port 4 (End Inlet) Wall Port 5 (Centre Inlet)

Combined Sleeve Port Exhaust Sleeve Port End Inlet Sleeve Port Centre Inlet Sleeve Port Piston

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The mechanism driving the sleeve produces an elliptical motion of the sleeve as

illustrated in Figure 20. The X and Y offset values can therefore be calculated with the

equation describing an ellipse as presented in Equation (3).

Figure 21: X and Y Coordinates of Ellipse at Crank Angle

Figure 21 illustrates the sleeve at crank angle (coordinates (x,y)), which represents

sleeve angle , with the sleeve angle being half that of the crank angle. The sleeve at

TDC is located at the upper most point on the ellipse (coordinates (0,a)). This leads to

an X value as function of the sleeve angle as calculated by Equation (4) and a Y value as

a function of the X value as calculated by Equation (5).

(4)

(5)

X

Y

a

b

y

x

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The values of x and y is subsequently adapted to produce the values for the X and Y

offsets and added to the base X and Y values of the sleeve at TDC in order to locate the

sleeve ports at any given crank angle, .

3.2 DETERMINING THE VALVE AREAS

In order to correctly simulate the flow through the engine the valve areas must be

known throughout the 720 crank angle cycle. Therefore the valve areas must be

calculated for every crank angle. This can be done by tracing the overlapping sleeve

and cylinder wall port profiles onto a piece of graphical paper and counting the square

millimetre blocks confined within the traced port outline. However, as this must be

done for all five valves at 720 different crank positions, it will result in a very time

consuming and inaccurate process due to the difficulty in correctly tracing the

overlapping port shapes in the confined space of the cylinder. It was subsequently

decided to use the coordinates of the sleeve and cylinder wall ports as determined in

the previous section to calculate the valve areas for all 720 crank angles.

3.2.1 Initial Method of Calculation

At first it was thought to perform curve fitting to various sections of the coordinated

points identified in the port profiles and then to determine the integral of these curves

over their various ranges of applicability and finally to add these areas in order to

obtain the total area of a certain port. Figure 22 illustrates the curves fitted and their

accompanied equations to eight different zones identified around the sleeve port

profile.

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Figure 22: Curves Fitted to Points Describing Sleeve Port

A number of problems arose with this method however. Firstly, the equations for the

curves changes with every change in crank angle (subsequent change in position) and

therefore the integrals must be repeated for every crank angle, resulting in a very

laborious and time consuming process. Secondly, when the ports overlap, the exact X

coordinates where the port profiles overlap are unknown and hence the ranges of the

applicable integrals are unknown, resulting in incorrect calculations of the areas.

Finally, because the curve fittings are just a mathematical approximation, the curves

does not exactly represent the various profiles, resulting in inaccurate calculation of

y = 0.045x3 - 7.623x2 + 424.2x - 7747.y = 120.4

y = -0.016x3 + 1.527x2 - 47.75x + 617.9

y = 3.331x + 45.36

y = 0.5x + 143.2y = -0.003x2 + 0.303x + 154.5

y = -0.165x2 + 15.69x - 210.0

y = -0.124x2 + 10.40x - 47.03

110

120

130

140

150

160

170

0 20 40 60 80

1

2

3

4

5

6

7

8

Poly. (1)

Linear (2)

Poly. (3)

Linear (4)

Linear (5)

Poly. (6)

Poly. (7)

Poly. (8)

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the area. This was realised when the area of the port shown in Figure 22 was

calculated using the method described above. This value was compared to the area

determined by counting the square millimetre blocks on the graphical paper for the

sleeve port in question. Although counting the blocks is also a time consuming

process, it is very accurate and the area calculated using this method resulted in

approximately 1200 mm compared to an area of roughly 1400 mm calculated with

the curve integrals. This confirms the need for a more accurate, generic and quicker

method of calculating the port areas.

3.2.2 Automated Method of Calculation

The points identified on the port profiles are located so that when the points are

connected with a straight line it would still yield a very similar profile as the actual

shape. On curved parts of the profiles the points are highly populated and on

straighter parts the points are more sparsely populated. The region between two

adjacent points could therefore be approximated with a straight line and the area can

easily be calculated as the area of a trapezoidal, being the area from the X-axis to the

straight line for the range on the X-axis.

The port area is subsequently obtained by subtracting the area of the bottom part of

the port profile from the area of the top part of the profile. However, this procedure

works well only when calculating the area of an entire port. Problems arise however,

when the sleeve port and the cylinder wall port overlap and only a certain part of each

profile must be taken into account and the exact points of overlap is unknown. To

overcome this problem, the entire range of each port on the X-axis was divided into

0.25 mm sections. New points were created by linearly interpolating between

adjacent points in order to have points at every 0.25 mm intervals. The linear

interpolation was done by Equation (6) where (x1,y1) and (x2,y2) are two original

adjacent points and (x,y) is the newly created points with at intervals of

0.25 mm.

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(6)

Figure 23 illustrates the original points describing one of the sleeve port profiles

together with the linearly interpolated added points at 0.25 mm intervals. This

procedure was done for all the cylinder wall ports as well as for the sleeve ports.

Subsequently, the biggest interval in X values is 0.25 mm resulting in a very small

potential error in determining the exact point of intersection in the case of port

overlap.

Figure 23: Points Describing the Sleeve Port Profile

The coordinates of the cylinder wall ports remain unchanged when the crank angle

changes, but as described in the previous section, the sleeve port coordinates change.

The procedure of adding points at every 0.25 mm interval on the X-axis was done for

the sleeve ports as well and it is subsequently easy to determine the points of

intersection between the wall and sleeve ports to within 0.25 mm.

120

125

130

135

140

145

150

155

160

165

24 29 34 39 44 49 54 59

Y-A

xis

X-Axis

Original Points Added Points (0.25mm Intervals)

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This results in the range of X values where the wall and sleeve ports overlap being

known, as well as the Y values accompanying these X values, so that the area of the

open valve can be calculated. However, the issue of whether the piston will mask the

valve area at certain crank angles is still unattended. As illustrated by (Hendrickson,

1999) the piston movement covered the valve openings when moving up to TDC in the

exhaust stroke and moving down from TDC in the intake stroke, effectively reducing

the valve areas. An equation presented by (Bosch, 2004) was used to describe the

piston movement and an appropriate Y offset value was added in order to ensure the

piston is at Y = 0 at BDC. The equation for the piston movement is given by

Equation (7).

(7)

The resulting piston movement is presented in Figure 24. A horizontal line was added

to the port coordinates and taken into account when determining the Y values for the

valve opening.

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Figure 24: Piston Movement with Crank Angle

Finally all the necessary data are available to calculate the valve opening area. This

includes the range of X values at which a cylinder wall port and its associating sleeve

port overlap, as well as the accompanying Y values that describes the open part of the

overlap. These Y values also include the presence of the piston where applicable. It

was decided to use the equation for calculating the area of a trapezoid because two

adjacent X values and their respective two associated Y values are situated in the form

of a trapezoid as illustrated in Figure 25. The equation is presented in Equation (8) and

Figure 25 also illustrates the definitions of the terms used in the equation.

(8)

0

10

20

30

40

50

60

70

80

90

0 100 200 300 400 500 600 700

Y-A

xis

[mm

]

Crank Angle [deg]

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Figure 25: Trapezoid from Adjacent Valve Opening Points

All the small trapezoid areas describing each valve opening were added together to

produce the total valve opening for each valve. This was done at all the crank angles

for 1 full cycle (0 to 720) and plotted to produce Figure 26. Note the sudden drop-

offs in the range between 300 to 400 due to the piston masking the valve openings.

91

92

93

94

95

96

97

98

99

100

48.2 48.25 48.3 48.35 48.4 48.45 48.5 48.55

Y -

Axi

s

X - Axis

(x1t,y1t)

(x1b,y1b)

(x2t,y2t)

(x2b,y2b)

y1 y2

x

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Figure 26: Valve Areas Plotted Against Crank Angle

3.3 VALVE MODELS

The WaveBuild software has a number of options available to specify the valve models

with. Unfortunately, there are no models which are directly applicable to sleeve valves

and it was subsequently decided that the best alternative would be to use the effective

area valve model. This model requires the valve area as function of the crank angle, a

diameter and the valve flow coefficients as function of the pressure ratio across the

valve and the valve lift.

As described in the previous section, the valve area was determined as function of the

crank angle. The area data was entered into a file in the format as specified by the

WAVE user manual for valve effective area files. These files were then specified as the

areas for the various valves leading to input pages similar to the one presented in

Figure 27. Notice that WAVE automatically converts the effective area to valve lift

values.

-100

0

100

200

300

400

500

600

700

800

900

0 100 200 300 400 500 600 700 800

Are

a [m

m^

2]

Cycle Angle [deg]

End Inlet 1 Valve Area [mm^2] Exhaust 1 Valve Area [mm^2] Exhaust 2 Valve Area [mm^2]

End Inlet 2 Valve Area [mm^2] Centre Inlet Valve Area [mm^2]

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Figure 27: Typical Input Page for Effective Valve Area

The valve diameter is used to convert the effective area plot to a valve lift plot

, plotted against crank angle. In the simulations this will be converted

back to an effective area and therefore, any reasonable diameter can be used, as long

as it is used consistently. For the initial model, all the valve diameters were specified

as 20mm.

This leaves only the discharge coefficients to be determined. Due to the fact that this

is an initial WAVE model, it was decided that the coefficients as described in Section

2.3.1 will be sufficient. Discharge coefficient determined from the figures presented

by (Waldron, 1940) was copied into a file with the format of the file as specified by the

WAVE user manual for valve discharge coefficient files. These files were then specified

in the WAVE model as the discharge coefficients for the various valves. Figure 28

illustrates the coefficients used. Notice only one profile per valve, as (Waldron, 1940)

concluded that very similar coefficients were acquired for different valve openings.

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Figure 28: Discharge Coefficients as taken from (Waldron, 1940)

These discharge coefficients were entered as a function of the pressure ratio and

repeated for two different valve lifts, one small lift value (0.1 mm) and one large lift

value (15 mm), resulting in a typical input page presented in Figure 29 (centre inlet

valve in this case).

Figure 29: Input Page for Valve Discharge Coefficient

0.6

0.65

0.7

0.75

0.8

0.85

0.9

0.95

1

1.0000 1.5000 2.0000 2.5000 3.0000 3.5000 4.0000 4.5000

Dis

char

ge C

oe

ffic

ien

t

Pressure Ratio

Centre Inlet Valve End Inlet Valves Exhaust Valves

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3.4 INTAKE FLOW PATH

The flow path of the intake system comprises of an inlet pipe, throttle valve of the

carburettor and then the inlet manifold leading into the three inlet valves. The inlet

pipe and the inlet manifold are modelled using duct elements and these two parts are

joined by a Y-junction element. The throttle is specified as an orifice, splitting the inlet

pipe into two sections before entering the junction element. A fuel injector is also

added to the second part of the inlet pipe to facilitate fuel delivery to the system. The

injector was set to deliver an air fuel ratio (AFR) of 14.7, thereby assuming

stoichiometric combustion. This layout is presented in Figure 30.

Figure 30: WAVE Layout of Intake Flow Path

3.4.1 Geometry

The carburettor is connected to an inlet manifold. The manifold comprises of a C-

shaped steel ducting that bolts over the exposed ports in the cylinder wall. This

ducting directs the flow towards the three inlet ports which are situated at roughly 90

intervals around the barrel. The area around the ports is cleared of cooling fins in

order for the ducting to attach onto the outside of the cylinder barrel. The side of the

manifold connecting to the barrel is open, thus using the barrel as one of the sides

enclosing the inlet flow path. Figure 31 illustrates these components.

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Figure 31: Inlet Manifold Duct and Cylinder Barrel

The geometry of the inlet manifold duct therefore defines the flow path of the air and

it is graphically presented in Figure 32, showing the main dimensions. Inside the

ducting there are no obstructions and the air is free to move undisturbed. The

curvature of the flow around the barrel to the two end inlet valves are supported by

slopping cut-out sections into the barrel to maximize the flow area.

Figure 32: Inlet Manifold Geometry

191

139

26 26

30

30

15

9

0

38

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As illustrated in Figure 30 the inlet flow path will be modelled by using duct, orifice and

Y-junction elements. The input geometrical values for the various duct elements are

taken from Figure 32, with the carburettor having the same diameter as the pipe it

connects to. The two pipes leading to the two end inlet valves are noncircular and

therefore the hydraulic diameter equation (Equation (9)) was used to determine the

input diameter values for the ducts.

(9)

The resulting geometrical input values for the ducts of the intake system are presented

in Table 1. It should be noted that the friction multiplier for the three ducts leading to

the inlet valves are set at 0, implying no pressure loss due to friction. This is done

because the pressure loss due to friction is already taken into account in the discharge

coefficients of the valves.

Table 1: Geometrical Input Values for Intake Flow Path Ducts

Left

Dia

me

ter

[mm

]

Rig

ht

Dia

me

ter

[mm

] D

iscr

etiz

atio

n

[mm

]

Ove

rall

Len

gth

[m

m]

Ben

d A

ngl

e

[deg

]

Fric

tio

n

Mu

ltip

lier

Hea

t Tr

ansf

er

Mu

ltip

lier

Carb1 38 38 15 100 0 1 1

Carb2 38 38 15 15 0 1 1

DuctEI1 33.53 30.875 15 90 90 0 1

DuctCV 38 38 15 10 0 0 1

DuctEI2 33.53 30.875 15 90 90 0 1

The discretization lengths were calculated with an equation given in the WAVE user

manual, Equation (10). The manual suggests using this equation to calculate the

discretization size in order to acquire the best compromise between accuracy and

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computing time, as minimizing the discretization length will increase the accuracy but

also increase the computing time.

(10)

with

where is the engine speed in revolutions per minute, and is the speed of sound.

As the engine will probably not be ran above 6000 rpm, it was decided to calculate the

discretization for this speed and subsequently it will be sufficient for lower speeds.

This resulted in a discretization of approximately 15 mm.

3.4.2 Heat Transfer

Heat transfer inherently implies the transfer of heat from a medium which consist of

heat to a medium which consists of less heat. This phenomenon is therefore driven by

a difference in heat between two mediums which imply a temperature difference

between the two mediums. The three methods of heat transfer are convective,

conductive and radiation heat transfer. All these methods rely on a temperature

difference between two mediums and a higher temperature difference implies higher

heat transfer.

Consider the intake system, remembering that this is a normally aspirated engine.

Therefore, the temperatures throughout the intake system will be at a similar

temperature as the ambient surrounding temperature. Subsequently very little heat

transfer will take place and it was therefore decided not to simulate heat transfer in

the intake system. However, it was realised that a part of the two intake ducts leading

to the end inlet valves are directly in contact with the cylinder barrel which will be at a

considerably higher temperature as the ambient temperature and thus a significant

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amount of heat transfer will take place. It is hence imperative that the heat transfer in

these two ducts is simulated.

A first thing to notice is that the heat transfer multipliers as specified in Table 1 in the

previous section are all set to 1, even for the ambient temperature intake ducts. The

heat transfer calculated in that case is convective heat transfer between the fluid

stream in the duct and the boundary layer. Due to the friction between the boundary

layer and the duct wall it was decided to consider this convective heat transfer.

However, the conductive and radiation heat transfer of the ambient inlet ducts will be

ignored, but these heat transfer terms will be included in the analyses of the two ducts

which are in contact with the cylinder barrel.

A problem arises when attempting to activate the conduction and radiation heat

transfer to the two ducts which are in contact with the cylinder barrel. The problem is

that only one side of the duct is connected to the hot cylinder barrel and if the

geometries of theses ducts remain as they are specified in Table 1, excessive heat

transfer will take place due to the heat transfer area (the outside area of the duct)

being larger than the actual heat transfer area (only the one side). Thus, a way must

be found to decrease the heat transfer area without affecting the pressure loss and

mass flow rate through these ducts or their acoustic behaviour. In order to keep the

mass flow rate in tact the same diameters must be used as specified in the table. As

far as the pressure loss is concerned, altering the length of the ducts will not affect the

pressure loss, because the pressure loss of these ducts is already accounted for in the

discharge coefficients of the valves. Therefore, the length and thickness of these ducts

can be altered in order to accurately specify the heat transfer area. The thickness has

no affect on either the mass flow rate or pressure loss.

Unfortunately, altering the length of the pipe will affect the acoustic pressure wave in

the duct and ultimately the effective mass flow rate. It was subsequently decided to

divide each of the two intake ducts that lead to the end inlet valves into two separate

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ducts. The combined length of these two ducts will be the same as the geometrical

length of the duct, one of the ducts will model conduction and radiation heat transfer

while the other duct will model the convective heat transfer.

In order to calculate the input values the area and volume of the duct with conduction

and radiation must be equal to the area of the cylinder barrel that is in contacts with

the flow and the volume of that part of the barrel. According to the engine drawings,

that part of the barrel is roughly a block of 50 mm long, 38 mm high and 25 mm deep.

The contact area is only one face in the length and one face in the depth of the block.

Assuming that this duct will be placed adjacent to the valve, the diameter of the duct

will be 30.875 mm as presented in Table 1. Therefore,

and

thus

and

Solving these equations simultaneously leads to a duct length, , of 14.13 mm and a

thickness, , of 16.67 mm and a new length of the accompanying duct of 75.87 mm.

The new input values for these ducts are presented in Table 2 and the layout is

presented in Figure 33.

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Figure 33: WAVE Layout of Intake Including Heat Transfer Ducts

The cylinder barrel is a cast aluminium, air cooled cylinder block. (Incropera & De Witt,

1996) provides the following properties for cast aluminium:

Density = 2790 *kg/m+

Specific heat cp = 883 [J/kg.K]

Thermal conductivity k = 168 [W/m.K]

Emissivity 0.8

This leads to a heat capacity of roughly 2.46 x 106 [J/m.K]. The temperature of the

cylinder barrel was assumed to be 400K, but should be calibrated once experimental

data becomes available.

Table 2: Input Values for Intake Heat Transfer Ducts

DuctEI1 DuctHTEI1 DuctEI2 DuctHTEI2

Left Diameter

[mm] 33.53 30.875 33.53 30.875

Right Diameter

[mm] 30.875 30.875 30.875 30.875

Discretization

[mm] 15 14.13 15 14.13

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DuctEI1 DuctHTEI1 DuctEI2 DuctHTEI2

Overall Length

[mm] 75.87 14.13 75.87 14.13

Bend Angle [deg] 90 0 90 0

Friction Multiplier 0 0 0 0

Heat Transfer

Multiplier 1 0 1 0

Outer Wall

Thickness [mm] - 16.67 - 16.67

Heat Capacity

[J/m.K] - 2.46 x 106 - 2.46 x 106

Conductivity

[W/m.K] - 168 - 168

Convective Field

Temperature [K] - 400 - 400

Radiation Field

Temperature [K] - 400 - 400

Emissivity - 0.8 - 0.8

3.4.3 Junction

The modelling of the intake flow path consists of a Y-junction model that connects the

inlet pipe, following the carburettor throttle valve, and the three inlet manifold ducts.

A Y-junction element was used and specified with a diameter of 38 mm. The friction

and heat transfer multipliers were specified as 1 to account for the friction and

convection heat transfer, but because the junction does not contact any part of the

hot cylinder barrel, the conduction and radiation heat transfer were omitted. The

junction openings were set up as presented in Figure 34.