Download - 41802098 Sleeve Valve Report
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Cranfield University
Adriaan Moolman
Modelling of a 4-Stroke Sleeve
Valve Engine
School of Engineering
MSc Automotive Product Engineering
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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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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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Modelling of a 4-Stroke Sleeve Valve Engine August 2007
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.