An Investigation into the Combustion Characteristics of a SI engine to Verify Simulation ModelsJack Lu

Abstract

Introduction

Background

Today’s automotive industry heavily relies upon the use of engine simulation software to

develop and design both race and conventional engines. Sophisticated one dimensional engine

Computational Fluid Dynamics (CFD) packages such as RICARDO WAVE® cost millions of

dollars but are capable of producing results within an error of 1-3% of dynamometer (DYNO)

results1. Like any other CFD or Finite Element Analysis (FEA) software, the results produced are

only as good as the data input.

UWAM has been utilising WAVE® to design their powertrain package since 2004 and over the

past three years have defined their model to within 10% of their DYNO torque curve. The

significance of acquiring such an accurate model for UWAM is so that the effects of powertrain

hardware can be simulated during the design stage. Without an accurate engine simulation,

multiple prototypes have to be manufactured and tested. In some cases large modifications to the

engine itself can deem the resultant hardware ineffective and hence ultimately cost the team a

large amount of time and finances.

Recent work produced by Lu, 2007, and advances in engine research within UWAM have

enabled one to retrieve pressure information from within a cylinder of the Honda CBR600R.

When this data is logged alongside with crank angle it is capable of producing a Pressure-

Volume (P-V) plot that is undoubtedly invaluable to the understanding of the thermodynamic

system of the engine.

Literature Review

Past UWAM theses and engineering projects extensively cover intake design and gas exchange

for a Formula SAE car particularly naming Kitsios (2002), Inkster (2004), Kawka (2005) and

Rogozinski (2006). Kitsios centred his work upon the theoretical fundamentals of gas dynamics

and produced a basis for engine simulations upon which Rogozinski built his CFD simulations

and experimental analysis upon. The core of Inkster and Kawka’s work revolved around the

design of a variable runner intake plenum and proved that educated decisions can be made from

full engine simulation models. Paget (2003) evaluated and utilised the Stanford Engine Simulation Software to design for and optimise engine valve train. His work included the

use of an in-cylinder pressure transducer but high levels of noise with secondary cyclic signals

superimposed upon the main signal were evident. These were concluded to be caused by the

existence resonance occurring within the 30mm long bore between the piezoelectric sensor and

the cylinder.

‘Design and Simulation of Four Stroke Engines’ by Prof. Gordon P. Blair covers

engine simulation techniques in high detail with extensive references to the fundamentals of

thermodynamics. Experimental research and case studies are provided to verify his models along

with a great deal of advice on increasing engine efficiency and performance.

‘Internal Combustion Engine Fundamentals’ by John B. Heywood is an extensive

review of the vast and complex mass of technical material that now exists on spark-ignition and

compressionignition engines. Heywood comprehensively covers all aspects of gas dynamics

related to the internal combustion engine by applying the laws of chemistry and

thermodynamics. A great deal of Heywood’s work is backed up by experimental results and

illustrations.

‘Measuring Absolute-Cylinder Pressure and Pressure Drop Across Intake Valves of Firing Engines’ by Paulius V. Puzinauskas, Joseph C. Eves and Nohr F. Tillman

is a technical paper describing a technique which can accurately measure firing-cylinder full-load

absolute pressure during intake events, thereby providing useful cylinder-pressure data for valve-

timing optimisation.

‘Spark Ignition Engines – Combustion Characteristics, Thermodynamics, and the Cylinder-

Pressure Card’ by Frederic A. Matekunas is a research paper covering the thermodynamics

theory behind combustion and discusses about the factors that are important to the timing of the

burn for maximum brake torque operation.

Combustion Process within the Four Stroke Cycle

An internal combustion engine gains its energy from the chemical energy released during the

combustion of the fuel/air mixture and therefore the combustion process dictates engine power,

efficiency and emissions. The combustion process of a four stroke spark ignition (SI) engine can

be divided into four distinct phases: spark ignition, early flame development, flame propagation

and flame termination. The four phases lie between the compression and power strokes of the

four stroke engine cycle (Figure). During the intake stroke the piston falls from top dead centre

(TDC) increasing the cylinder volume while the intake valve is open. A fresh charge of fuel/air

mix is inducted through the intake valve and into the cylinder mixing with the residual gas that

remains in the cylinder from the previous cycle. During the compression stroke all valves are

closed and the cylinder volume decreases as the piston moves up from bottom dead centre

(BDC) compressing the gas mix. The combustion process is initiated by the spark plug towards

the end of the compression stroke under normal operating conditions and continues through to

the early portion of the power stroke. At this point a turbulent flame develops and propagates

through the fuel/air/residual gas mix away from the spark plug and towards the chamber walls

before extinguishing. Upon the start of the power stroke the cylinder pressure increases

significantly and work is transferred to the piston pushing it down towards BDC ultimately

increasing cylinder volume. The exhaust valve opens before BDC and the exhaust stroke expels

the exhaust gases from the rising piston leaving some residual gases behind.

Spark Ignition

The ignition within an SI engine is provided by the discharge of the spark plug that is generally

controlled by an electronic control unit (ECU). The spark ignition initiates the combustion

process and therefore controls the burn.

Early Flame Development

The early fame development (EFD) stage comprises of the flame development process from the

spark discharge which initiates the combustion process to a point where a small but measurable

fraction of the charge has burned or fuel energy released. In industry it is common to indicate the

end of the EFD stage when 10% of the charge mass has been burnt. Other figures such as 1% and

5% have been used also.

Flame Propagation

The flame propagation stage comprises of the rapid burning of the charge. During this stage each

element of fuel/air burns and its density decreases by a factor of four. The expansion of the

combustion product gas compresses the mixture ahead of the flame and displaces it towards the

chamber walls. At the same time the already burnt gas behind the propagating flame is

compressed and displaced towards the spark plug. Elements of the unburnt gas are of different

temperatures and pressures just prior to combustion and are at different states after combustion

and their condition is determined by the conservation of mass and energy.

Flame Termination

The flame termination stage comprises of the propagating flame reaching the chamber walls and

extinguishing. At this point the combustion process has ended and a large portion if not all of the

fuel energy has been released to produce work onto the piston. The amount of fuel energy

released is dependent upon the efficiency of the expansion in burn.

Variables Effecting Combustion

Combustion Phasing

Combustion events can be phased by advancing or retarding spark before top dead centre

(BTDC). The phasing of the combustion event influences the magnitude and location of peak

cylinder pressure by changing the rate of pressure rise within the chamber. FIGURE illustrates

the effects of combustion phasing by spark advance upon cylinder pressure.

By phasing the combustion so that the 50% mass burned point is closer to TDC allows complete

combustion at TDC and therefore increases the compression stroke work transfer from piston to

cylinder gases resulting in higher cylinder peak pressure. Ultimately this leads to increased work

transfer from the cylinder gas to piston upon the power stroke increasing the brake torque output.

Matekunas, 1984, introduces the idea of “phase loss” defined as the loss in efficiency as the 50%

mass burned point is moved away from TDC. The optimum phasing that provides maximum

brake torque (MBT) is known as the MBT point and any timing advanced or retarded from this

point increases the “phasing loss” and produces lower torque.

Cylinder Turbulence

The combustion process in a SI engine occurs in a turbulent flow field. This flow field is

produced by the high shear flows generated by the intake jet and flow pattern. In turbulent flows,

the rate of transfer and mixing are several times greater than the rates due to molecular diffusion

[Heywood (1988)]. One method of adding turbulence within the combustion chamber is known

as squish action and this is caused by the geometry of the combustion chamber as the piston rises

and compresses the gas. Squish characteristics in SI engines are relatively moderate compared to

that of a compression-ignition engine. Another method of promoting turbulence is through swirl

and tumble caused by the intake geometry.

Swirl and Tumble

The terms ‘swirling’ and ‘tumbling' are used to describe the rotating of flow within the cylinder.

Swirl is defined as the controlled rotary motion of the charge about the cylinders axis whereas

tumble (FIGURE) is in cylinder flow at right angles to the cylinder axis. They are created by

providing an initial angular momentum to the charge as it enters the cylinder through the intake

ports. Swirl and tumble can assist in speeding up the combustion process within SI engines and

hence achieve higher thermal efficiency.

Tumble

Measuring Swirl and Tumble

Swirl ratio

Compression Ratio

The compression ratio (CR) is defined as the ratio of maximum volume (when the piston is at

BDC) to minimum volume (when the piston is at TDC). At BDC the volume comprises of the

swept volume Vs and the clearance volume Vc whereas at TDC the minimum volume at which

combustion occurs consists of only the clearance volume Vc.

CR=maximum volumeminimum volume

=V bdc

V tdc=

V s+V c

V s

In the APPENDIX Blair (1999) proves that the highest thermal efficiency is achieved at the

highest compression ratio but if the compression ratio is too high, engine operation will exhibit

abnormal combustion which is an undesirable outcome.

Abnormal Combustion

Normal combustion is initiated by the discharge of the spark plug and develops a flame that

propagates to the chamber walls before extinguishing but there can be several factors to cause

abnormal co mbustion. These factors are fuel composition, engine design and operating

parameters and combustion chamber deposits (Heywood, 1988). The two most common forms of

abnormal combustion are identified as knock and surface ignition. Both of these reduce the

combustion efficiency and through persistence will destroy engine components by exceeding the

engines pressure design limits. FIGURE illustrates the difference normal and abnormal

combustion as seen from a pressure trace.

Knock

Knock is described as the sharp metallic noise caused by the auto-ignition of the fuel/air/residual

gas mix ahead of the propagating flame. During combustion the propagating flame compresses

and displaces the end gas ahead of the flame towards the chamber wall. This causes its pressure,

temperature and density to increase undergoing the chemical reactions prior to normal

combustion. When pressures and temperatures become excessive the end gas burns very rapidly

releasing a large amount of its energy at a rate five to twenty five times normal combustion

causing high frequency pressure oscillations within the chamber that exceed engine design

limits. These detonations are initiated by high pressures and temperatures and therefore can be

avoided by reducing the compression ratio, using a higher rating octane fuel, appropriate

calibration of the engines ignition timing and careful design of the engines cooling system.

Surface Ignition

Surface ignition is the uncontrolled ignition of the fuel/air/residual gas mix from overheated

valves, walls, spark plug or glowing deposits. There are two types of identifiable surface

ignition: pre-ignition and post-ignition. Pre-ignition can be identified from a pressure trace as the

combustion event is initiated before the targeted spark ignition time and causes the most severe

effects as the spark no longer controls the combustion process. Post-ignition occurs after the

spark ignition but can be difficult to distinguish from knock as they both portray the same

characteristics under a pressure trace.

Cyclic Variations

It is evident from observation of cylinder pressure versus crank angle measurements over

consecutive cycles within a sample that cyclic variation exists. For a motored pressure trace

cyclic variations are negligible and pressure measurements tend to follow closely to the

polytropic relationship pV n=constant . Therefore the pressure development is distinctively

related to the combustion process which is dependent upon different variables. These cyclic

variations are caused by variations in charge motion and mixture motion at the time of the spark,

the amount of fuel/air within the cylinder and the fuel/air ratio, and the mixing of the fresh

mixture with the residual gases remaining in the cylinder. Along with cylinder cyclic variations

there also exists cylinder to cylinder variance in multicylinder engines which are caused by the

same reasons.

It is important note that due to cyclic variations, the optimum combustion phasing will then be

different for each variation of combustion and that the MBT spark advance experimentally found

from engine tuning methods described by Bleechmore (2006) is set for the average cycle. Any

cycles faster than the average cycle effectively advances spark timing away from the MBT point

and any cycles slow than the average cycle effectively retards spark timing away from the MBT

point.

Combustion Characteristics

Combustion can be analysed from within an SI engine using a pressure trace acquired from

within the combustion chamber with relation its crank angle. When a sufficient number of cycles

are recorded, the data is capable of producing combustion characteristics such as the pressure

plot, P-V diagram, indicated mean effective pressure (IMEP), friction mean effective pressure

(FMEP), mass fraction burned (MFB), burn duration and the coefficient of variance (COV).

These characteristics are vital for describing the combustion process and its efficiency.

Pressure

Cylinder pressure and crank angle are commonly logged and plotted against each other when

analysing combustion. Without ignition the combustion does not occur and the pressure recorded

describes the motoring pressure within the cylinder seen in FIGURE. This is the pressure that the

cylinder experiences from its change in volume. When ignition occurs, the charge mass burns

and the cylinder pressure significantly increases causing a higher transfer of work onto the

piston. Note that ignition begins BTDC and peak pressure occurs ATDC as seen in FIGURE

Parameters of interest include magnitude and crank angle of maximum pressure, and magnitude

and crank angle of the maximum pressure rise. The rate of pressure rise is calculated using the

simple numerical differentiation in EQUATION.

dpdθ

=pi+1−pi−1

θi+1−θ i−1

Ricardo [REF] states that for maximum efficiency the pressure rise rate should be 2.3 bar/degree.

Pressure-Volume Diagram

The thermal cycle of an SI four stroke engine can be illustrated by mapping the pressure-volume

(P-V) data extracted from a pressure trace. As the cylinder volume is a function of the crank

angle, it is possible to relate cylinder pressure to cylinder volume and hence construct a P-V

diagram as seen in FIGURE. Typical valve events such as intake valve open (IVO), intake valve

close (IVC), exhaust valve open (EVO) and exhaust valve close (EVC) are shown in the diagram

along with direction indicators to clarify the process.

Figure x.x: Typical P-V diagram for a four stroke SI engine

The area under the curve is the indicated work per cycle as given by EQUATION where p is

cylinder pressure and V is cylinder volume.

W /cycle=∫ p ∙dV

From FIGURE it can be seen that there are three distinctive areas known as Area A, Area B and

Area C. The integral over the exhaust and intake strokes (Area B + Area C) is the indicated work

done on the gas by the piston known as pumping indicated work whereas integral the over the

compression and power strokes (Area A + Area C) is the indicated work done onto the piston by

the gas known as gross indicated work. The work generated throughout the entire cycle is then

known as the net indicated work. Note that work out of the system is negative and work into the

system is positive.

Indicated Mean Effective Pressure

While the cylinder pressure and volume varies throughout the engine cycle, an imaginary

constant pressure difference can be substituted over the volume change to obtain the same net

work (Spencer, 2004). This pressure difference is known as the indicated mean effective pressure

(IMEP) and is used to assess combustion performance independent of the size of bore and stroke,

speed and number of cylinders in the engine. FIGURE shows a rectangle with a height that

represents the pressure difference that is IMEP and contains an equal area representing the

identical work done by the complex cycle shape.

In accordance to the definition of net and gross work, Elmqvist-Möller (2006) defines the net

IMEP (N.IMEP) and gross IMEP (G.IMEP) in EQUATION and EQUATION respectively.

N . IMEP=W n

V d= 1

V d∙ ∫

0o

720o

p ∙ dV

G . IMEP=W g

V d= 1

V d∙∫

360o

720o

p ∙ dV

Where Wn is the net indicated work, Vd is the swept cylinder volume and p is the cylinder

pressure. The work obtained is integrated between crank angles where 0o is TDC upon the intake

stroke and 720o is TDC upon the end of the exhaust stroke. The difference between net and gross

IMEP is known as pump mean effective pressure (PMEP) or pump loss given by EQUATION.

PMEP is the measure of work done by the engine expressed in units of pressure and therefore the

relationship between N.IMEP, G.IMEP and PMEP is seen in EQUATION.

N . IMEP=G . IMEP+PMEP

Brunt (1980) [14 in loughbrough thesis] outlines that errors in IMEP calculations are mainly

caused by thermal shock, crank angle phasing errors and transducer sensitivity. Minor errors are

caused by coarse crank angle resolution, incorrect con rod length, signal noise and integration

period error.

Friction Mean Effective Pressure

The friction mean effective pressure (FMEP) is the measure of frictional losses that contribute to

the lower brake torque experienced at the crankshaft output expressed in units of pressure. The

sum of N.IMEP and FMEP then result in the brake mean effective pressure (BMEP) measured at

the crankshaft output defined by EQUATION where τ is the brake torque and Vs is the swept

volume.

BMEP= τ ∙ 4 πV s

BMEP=N . IMEP−FMEP

FMEP is mathematically defined by the Chen Flynn (1965) model seen in EQUATION. This

experimentally derived model states that the total engine friction is a function of peak cylinder

pressure, mean piston speed and mean piston speed [GT Power].

FMEP=C+( PF ∙Pmax )+( MPSF ∙ Speedmp )+(MPSSF ∙ Speedmp2 )

Where

C is the constant part of FMEP

PF is the peak cylinder pressure factor

Pmax is the maximum cylinder pressure

MPSF is the mean piston speed factor

MPSSF is the mean piston speed squared factor

Speedmp is the mean piston speed

The two most common methods of measuring engine friction are motoring dyno testing and

comparing indicated torque (calculated from cylinder pressure) to brake torque. Motoring dyno

testing is recommended out of the two methods due to the difficulty of accurately measuring

cylinder pressure across the entire engine cycle and across the multiple cylinders. Unless

cylinder pressure measurement is taken from an average from several individual cylinders over

several engine cycles, cylinder-to-cylinder and cyclic variations can strongly effect the IMEP

measured when comparing to BMEP.

Mass Fraction Burned

The mass fraction burned (MFB) in an engine cylinder is a normalised quantity between a scale

of 0 and 1. It describes the chemical energy release as a function of crank angle as it measures

charge mass that has been burned during the combustion event. MFB plots are ‘S’ shaped as seen

in FIGURE and measures the fraction of charge mass which has burned within the cylinder at a

given crank angle. Additionally combustion duration and ignition delay are determined from

MFB curves. The ignition delay is the duration in crank angles between the start of combustion

and typically 1, 2 or 5% MFB and the burn duration of a cycle is simply calculated by the crank

angle duration from xb=0.1 and xb=0.9.

The MFB is most commonly estimated by the Rassweiler and Withrow method publicated in

1938 [REF] that is based upon the assumption that during engine combustion, the pressure rise

Δp consists of a pressure rise due to combustion Δpc and a pressure change due to a volume

change Δpv.

∆ p=∆ pc+∆ pv

At periods where there is no combustion, pressures at the start and end of interval Δθ are related

by the polytropic equation

pi V in=p j V j

n

Hence the pressure change due to a change in volume is given by

∆ pv=p j−p i=pi[( V i

V j )n

−1]And the pressure change due to combustion is given by

∆ pc=p j−pi( V i

V j)

n

Where i and j is the start and end of the interval respectively and n is referred to as the polytropic

constant. Since the combustion process does not occur at constant volume, a pressure rise during

combustion must be referred to a reference volume, such as that of the volume at TDC.

∆ pc¿=∆ pc ∙

V i

V TDC

Assuming that the pressure rise due to combustion is proportional to the mass of charge burned

within the interval Δθ then the MFB xb at the end of the ith interval is given by

xb(i)=mb(i)

mb (total)=∑

0

i

∆ pc¿

∑0

N

∆ pc¿

Where mb is the mass burned, 0 denotes the start of combustion and N is the total number of

crank intervals at the end of combustion. This method takes into assumption that the pressure rise

due to combustion is proportional to the amount of fuel chemical energy released rather than the

mass of mixture burned.

There are other methods of measuring MFB such as the Isermann and Muller approximation seen

in APPENDIX but the benefits of the Rassweiler and Withrow method is that no additional data

(besides pressure and crank angle) is needed. A functional form often used in engine simulation

to represent the mass fraction burned versus crank angle curve is the Wiebe function seen in

EQUATION.

xb=1−exp[−a (θ−θ0

∆ θ )m+1]

Where θ is the crank angle, θ0 is the start of combustion, Δθ is the total combustion duration

from xb=0 to xb=1, and a and m are adjustable parameters to change the shape of the curve to

fit.

Coefficient of Variance

Cycle by cycle variability can be measured by three means: pressure related parameters, burn-

rate related parameters and flame front positioning parameters. Pressure related parameters are

the easiest to determine and an important measure of cyclic variability that is derived from

pressure data is known as the coefficient of variance (COVimep). The COVimep is expressed as a

percentage and is defined by EQUATION where σimep is the standard deviation in IMEP and imep

is the mean IMEP.

COV imep=σ imep

imep∙100

σ imep=√ 1N−1∑i

N

( pi−p )2

The COVimep is a measure of the cyclic variability in the indicated work per cycle and it is noted

by Heywood (1988) that vehicle drivability problems arise when COVimep exceeds 10 percent.

Acquiring the Pressure Trace

Recommended Data Measurements

After building an engine model it is necessary to measure performance data in order to tune the

model. Ideally the following list of performance measurements at wide open throttle should be

available when tuning the model:

Brake Power/Torque

Motoring Friction Power

Air flow, Fuel Flow, Air-Fuel Ratio

IMEP, BSFC, volumetric efficiency or Mass Air Flow (MAF)

Intake and Exhaust manifold Temperature and Pressure (Time averaged)

Cylinder Pressure and/or Combustion Rate

Dynamic Intake and Exhaust Port Pressures

Mean Temperatures at Exhaust Ports and Tertiary pipe

It is common to have inaccurate experimental data and measured data should be validated.

According to the GT Power manual [REF], one method of validating measured data is to

calculate BSFC from volumetric efficiency, fuel-air ratio and brake power/torque.

UWAM Dynamometer and Data Acquisition

The engine test bed employed during testing consists of an engine, dynamometer (dyno) and

control system, instrumentation, data logger, exhaust extraction and cooling system for both

engine and dyno. The dyno control system allows the operator to hold a constant engine speed by

applying a torque to the engine via eddy currents for steady state tuning while the data logger

records information from the test session.

Work accomplished by Rogozinski, 2007, and Bleechmore, 2007, provided UWAM with a good

foundation of dyno data acquisition hardware including a MAF sensor and exhaust temperature

sensors. Lu, 2007, covered the installation of a Kistler Type 6005 pressure transducer and

successfully extracted a pressure trace using crank angle sensor selected by McDermont, 2006.

Along with these advancements in engine research at UWAM, the data acquisition hardware

installed upon the engine test bed allows the majority of the recommended data measurements

listed in SECTION to be recorded into the MoTeC M800 ECU and high speed DAQ card.

Testing

Results

Data Validation

logP logV

Measuring Swirl

Measuring Tumble

Calculating Cylinder Volume from Crank Angle

Engine analysis requires acquired engine data referenced to cylinder volume and this is achieved

by data acquisition at known crank angles. Cylinder volume is directly related to crank angle

through a slider-crank mechanism. FIGURE shows the piston restrained to move only in the y-

axis while the crank is restricted to a rotational degree of freedom.

Where

Vc is the clearance volume

Vg is the gasket volume

x is the stroke

b is the bore diameter

L is the con rod length

a is the crank radius

θ is the crank angle from TDC

The stroke x is calculated by

x=a+ L−√ L2−a2 sin2 θ+acosθ

Therefore the total cylinder volume at crank angle θ is the sum of the clearance volume Vc and

the swept volume Vs, which is the multiplication of the stroke and the bore diameter.

V cyl=V c+V g+xb=V c+b ∙ (a+L−[√L2−a2sin2 θ+acos θ ] )

EQUATION is not valid for all engines as it does not take into account the wrist pin offset that

some engines use to reduce the effects of piston slap. For the Honda CBR600RR, the engine of

interest in this investigation, this method of cylinder volume calculation is valid as there is no

offset.

The clearance volume Vc is experimentally measured by filling the combustion dome with a

Shellsol solution as seen in FIGURE and measuring the volume in which it retains. The gasket

volume is simply the cylindrical volume within the cylinder that exists due to the gasket that

separates the head and block. This measured by multiplying the cross sectional area of the

cylinder by the gasket thickness and it is important to note that measurements are taken from a

good but used gasket as new gaskets are thicker before being fitted.

RICARDO WAVE

Flow Coefficients

Swirl Coefficients

Combustion Modelling

SI Wiebe

Friction Model

WAVE’s friction model is defined by the modified Chen Flynn (1965) correlation seen in

EQUATION. Where A, B, C, D are the Chen Flynn coefficients, Pmax is the peak cylinder

pressure, RPM is the engine speed and stroke is the cylinder stroke.

FMEP=A+B ∙ Pmax+C ∙ RPM ∙( stroke2 )+D ∙(RPM ∙ stroke

2 )2