An Investigation into the
Combustion Characteristics of a SI
Engine to Verify Simulation Models
Jack Lu
Letter of Transmittal
Abstract
Table of Contents
Acknowledgements
1. Introduction
2. 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.
-2-
3. 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 sparkignition 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 cylinderpressure data for valve-timing optimisation.
-3-
‘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.
4. 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
seen in Figure 1. 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.
-4-
Figure 1 The 4 stroke engine cycle
4.1 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.
4.2 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.
4.3 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
-5-
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.
4.4 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.
5. Variables Effecting Combustion
5.1 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 2 illustrates the effects of combustion phasing by spark advance upon
cylinder pressure.
-6-
Figure 2 Combustion phasing by advancing spark timing
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. MBT phasing often
produces a peak pressure location within the range of 14-17deg ATDC. Matekunas
(1984)
5.2 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
-7-
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.
5.3 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 3) 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.
Figure 3 Tumbling of the intake charge within the cylinder
Measuring Swirl and Tumble
Swirl ratio
-8-
5.4 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 𝑣𝑜𝑙𝑢𝑚𝑒 𝑉𝑏𝑑𝑐 𝑉𝑠 + 𝑉𝑐
=
=
𝑚𝑖𝑛𝑖𝑚𝑢𝑚 𝑣𝑜𝑙𝑢𝑚𝑒 𝑉𝑡𝑑𝑐
𝑉𝑠
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.
6. 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 4 illustrates the difference normal and abnormal combustion as seen
from a pressure trace.
-9-
Figure 4 Pressure trace involving abnormal combustion
6.1 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.
- 10 -
6.2 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.
6.3 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 relationship𝑝𝑉 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.
- 11 -
7. Combustion Experimentation
7.1 Aim and Methodology
It is necessary to measure performance data in order to tune and validate simulation
models. The experiments discussed and outlined within this section include the
extraction of multiple pressure traces across the range of engine speeds from
3000RPM to 14000RPM, cylinder volume measurements and calculations, and the
measurement of port flow coefficients. The data obtained from these experiments are
then analysed to further develop the simulation models.
7.1.1 Recommended Data Measurements
All testing should be performed at wide opened throttle (WOT) to maximise the
volumetric efficiency within each RPM. Ideally the following list of performance
measurements at WOT should be available when tuning engine models:

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.
7.1.2 UWAM Dynamometer and Data Acquisition Apparatus
The engine test bed employed during testing consists of an engine, a dynamometer
(dyno) and dyno control system, sensor instrumentation, data logger, exhaust
extraction and cooling system for both engine and dyno. The dyno control system
allows the operator to hold a fairly constant engine speed by applying a torque to the
- 12 -
engine via eddy currents for steady state tuning while the data logger records
information from the test session.
Previous UWAM engine research allows for the majority of the recommended data
measurements listed in Section 7.1.1 to be logged by the MoTeC M800 ECU. The
experiments covered by this investigation utilises the Kistler Type 6005 piezoelectric
high pressure transducer with a pressure range of up to 1000bar and an operating
temperature range of -50oC to 200oC. Via a 10-32 to BNC gold wire, the pressure
transducer is connected to a charge amplifier that converts the charge into a voltage.
This signal, along with those produced by the cam pulse generator and crank angle
sensor, are then transmitted to the PCI-DAS 4020/12 high speed DAQ card and
logged by the TracerDAQ Pro software.
7.2 Acquiring the Pressure Trace
TracerDAQ Pro logs all inputs within the PCI-DAS 4020/12 high speed DAQ card
over a time domain. Triggered sampling is possible but did not produce a reading per
crank angle which shows a flaw in the programs ability to trigger at high speeds.
Therefore a specified sampling frequency must be specified by the user.
7.2.1 Sampling Frequency
To sample the pressure trace accurately, a significant amount of readings were needed
to be taken. The minimum amount of readings per cycle that is needed to develop an
accurate pressure trace is 360, one reading per degree. Inaccuracies then occur
because the real world steady state operation of the engine is not truly in steady state
as the engine dyno can only hold engine speeds to within +/- 50RPM accurately due
to dyno stability issues discussed by Bleechmore (2006) and therefore sampling the
trace with the minimum frequency will eventually lead to inaccuracies as the engine
speed phases out with the sampling frequency. Secondly upon logging the data with
the minimum frequency, each reading could be taken upon a non-5V point from the
crank encoder and therefore yield further inaccuracies therefore a higher resolution is
required. Table 1 below outlines the required frequencies that allow sampling to 4
readings per degree providing a high resolution sample. Sampling at a higher
frequency than those listed below will result in large files with more data to process
than necessary.
- 13 -
Engine Speed
(RPM)
420 (Motored)
2500
3000
3500
4000
4500
5000
6000
7000
8000
9000
10000
11000
12000
13000
14000
15000
Minimum Freq
(Hz)
2520
15000
18000
21000
24000
27000
30000
36000
42000
48000
54000
60000
66000
72000
78000
84000
90000
Required Freq
(Hz)
10080
60000
72000
84000
96000
108000
120000
144000
168000
192000
216000
240000
264000
288000
312000
336000
360000
Table 1 Required sampling frequencies to log a pressure trace accurately
7.2.2 Referencing TDC
Referencing TDC to the pulses produced by the crank angle sensor is achieved by
locating the SYNC signal produced by the cam pulse generator which occurs at a
certain number of degrees BTDC depending upon the camshaft timing. MoTeC refers
to this signal as the Crank Reference Index Position (CRIP), which is a user input
variable found by employing a timing light that is connected to the ignition circuit
under cylinder 1 and matching the TDC timing markers on the crankshaft and cover.
The unit of CRIP is in deg BTDC and therefore is added onto the SYNC signal to
provide the position of TDC. This method of locating TDC is dependent upon the
user’s CRIP input and therefore may not locate “true TDC”.
A more accurate method of locating and validating “true TDC” can be found through
analysing a motored trace without the presence of ignition and fuel. To eliminate
losses of suck air within the combustion chamber the engine is cranked via the starter
motor at WOT and because the pressures experienced within the cylinder are highly
dependent upon heat losses, the motored trace must also be acquired at operating
temperatures (ET 80oC). Due to the factor of time and heat loss, TDC does not occur
- 14 -
at peak pressure as expected in a theoretical adiabatic process. In fact peak pressure
occurs at 1o BTDC due to the heat losses within the compressed charge. The change
in volume between 1o BTDC and TDC is insignificant to a noticeable increase in
pressure but within this time heat is transferred out of the compressed charge hence
slightly dropping the pressure. [REF]
7.2.3 Signal to Pressure Conversion
The Kistler Type 6005 pressure transducer provides a gauge reading and does not
provide absolute pressures. Therefore the readings taken upon from the experiments
need to be referenced to a pressure that is known during the cycle. The intake
Manifold Air Pressure (MAP) sensor provides absolute readings after calibrating the
sensor within MoTeC and upon the IVC event, we can assume that the pressures
within the cylinder is equivalent that of the manifold air pressure as the cylinder is
filled with air. Therefore the pressure data is corrected by matching these values to
provide absolute pressures.
7.3 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 5 shows the piston
restrained to move only in the y-axis while the crank is restricted to a rotational
degree of freedom.
- 15 -
Figure 5 Schematic representation of a piston within cylinder
Where
Vc
Vg
x
b
L
a
θ
is the clearance volume (clearance + combustion dome)
is the gasket volume
is the stroke
is the bore diameter
is the con rod length
is the crank radius
is the crank angle from TDC
The stroke x is calculated by
𝑥 = 𝑎 + 𝐿 − √𝐿2 − a2 sin2 𝜃 + acos 𝜃
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.
𝑉𝑐𝑦𝑙 = 𝑉𝑐 + 𝑉𝑔 + 𝑥𝑏 = 𝑉𝑐 + 𝑉𝑔 +
𝜋 ∙ 𝑏2
∙ (𝑎 + 𝐿 − [√𝐿2 − a2 sin2 𝜃 + acos 𝜃])
4
- 16 -
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 illustrated by FIGURE. The volume in which it retains is
measured and added to the volume occupied by the piston clearance of 1mm with a
67mm bore diameter to establish the clearance volume.
The combustion dome
volume within the experimental engine is found to be 7.6ml across all four cylinders
and therefore acquires a clearance volume Vc of 11,125.6mm3.
The gasket volume is simply the cylindrical volume within the cylinder that exists due
to the gasket that separates the head and block. This is measured by multiplying the
cross sectional area of the cylinder bore by the gasket thickness, which is found to be
0.72mm on the experimental engine and therefore producing a gasket volume of
2,538.5mm3. It is important to note that measurements are taken from a used gasket as
new gaskets are thicker prior being fitted.
The dimensions of the engine components are listed in
Table 2 producing a sinusoidal relationship between crank angle and cylinder volume
seen in Figure 6. The mathematical model produces a pressure ratio of 11.96:1 and
therefore can be deemed valid as the measured compression of the engine was
recorded to be 12:1 from a compression gauge. These values agree with the specified
compression ratio within the Honda CBR600RR service manual.
Component
Crank radius (a)
Bore diameter (b)
Con rod Length (L)
Measured Dimensions
21.25mm
67mm
94mm
Table 2 Parameters measured from engine components
- 17 -
Crank Angle to Cylinder Volume
180000
Cylinder Volume [mm3]
160000
140000
120000
100000
80000
60000
40000
20000
0
0
45
90
135
180
225
270
315
360
Crank Angle [deg]
Figure 6 Cylinder volume vs. Crank Angle
7.4 Pressure Trace Data Validation
logP logV
7.5 Port Flow Testing
7.5.1 Definition and Background
The air flow characteristics within the combustion chamber are largely affected by the
intake and exhaust ports and, in turn, these air flow characteristics have a large impact
on combustion. There are two methods of measuring a port’s efficiency; steady-state
flow bench testing and computational fluid dynamics simulations. Of the two, this
investigation employed the use of a flow bench, which consists of the cylinder head of
interest being mounted onto a pipe that replaces the block and acts as a dummy
cylinder. Upstream of the port is a pressure box that is connected to a flow
straightener, a flow meter and a fan to draw/force air through the system. The flow
bench employed is similar to that of seen in Figure 7.
- 18 -
Figure 7 Typical flow bench layout (RICARDO)
The test consists of lifting the valve in increments of 1mm from very no lift (0mm) to
maximum lift (10mm) and forcing air through the ports. A pressure measurement is
then taken within the box and a pressure drop is recorded across the port and valve
geometry.
7.5.2 Flow Coefficient and Discharge Coefficient
The flow coefficient and discharge coefficient are both a measure of the ports
efficiency when compared to a theoretical unrestrictive port. These coefficients are
used interchangeably as the reference area differentiates the two. The flow coefficient
uses the circle of the valve throat (defined by the inner seat diameter) whereas the
discharge coefficient uses the valve curtain area. RICARDO. EQUATIONS are the
mathematical definitions of the flow and discharge coefficients where Aeff is the
effective throat area, D is the reference valve diameter and L is the valve lift. Valve
lift is typically specified by the non-dimensional L/D format.
𝐶𝑓 =
4 ∙ 𝐴𝑒𝑓𝑓
𝜋𝐷2
- 19 -
𝐶𝑑 =
𝐴𝑒𝑓𝑓
𝜋𝐷𝐿
With
𝑚̇√𝛾𝑅𝑇1
𝐴𝑒𝑓𝑓 =
1
𝛾−1
𝛾
𝑝 𝛾
2
𝑝
𝛾𝑝1 (𝑝2 ) √𝛾 − 1 [1 − (𝑝2 )
1
1
]
Where
𝑚̇
γ
R
T1
p1
p2
mass flow rate
is the ratio of specific heats
pressure
pressure
7.5.3 Test Results and Discussion
Any pressure loss due to temperature change should occur normally as the testing
described above is typically performed at atmospheric temperature and doesn’t
account for charge heating or cooling.
- 20 -
Intake Foward Port Flow Coefficients
0.8
0.7
Coefficient
0.6
0.5
0.4
0.3
Discharge Coefficients
0.2
Flow Coefficients
0.1
0
0.000
0.050
0.100
0.150
0.200
0.250
0.300
0.350
0.400
L/D
Exhaust Forward Flow Coefficients
0.8
0.7
Coefficients
0.6
0.5
0.4
0.3
Discharge Coefficient
0.2
Flow Coefficient
0.1
0
0
0.1
0.2
0.3
0.4
0.5
L/D
8. Combustion Characteristics & Results
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
- 21 -
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.
8.1 Pressure
8.1.1 Definition and Background
Cylinder pressure is the total pressure within the combustion chamber and is
commonly logged with crank angle 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.
𝑑𝑝 𝑝𝑖+1 − 𝑝𝑖−1
=
𝑑𝜃 𝜃𝑖+1 − 𝜃𝑖−1
Ricardo [REF] states that for maximum efficiency the pressure rise rate should be 2.3
bar/degree.
- 22 -
8.1.2 Test Results
8.2 Pressure-Volume
8.2.1 Definition and Background
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 8. 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 8 Typical Pressure-Volume diagram for a four stroke SI engine
- 23 -
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 8 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.
8.2.2 Test Results
8.3 Indicated Mean Effective Pressure
8.3.1 Definition and Background
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.
𝑜
720
𝑊𝑛
1
𝑁. 𝐼𝑀𝐸𝑃 =
=
∙∫
𝑝 ∙ 𝑑𝑉
𝑉𝑑 𝑉𝑑 0𝑜
- 24 -
𝑜
720
𝑊𝑔
1
𝐺. 𝐼𝑀𝐸𝑃 =
=
∙∫
𝑝 ∙ 𝑑𝑉
𝑉𝑑 𝑉𝑑 360𝑜
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 0 o 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.
8.3.2 Test Results
8.4 Friction Mean Effective Pressure
8.4.1 Definition and Background
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 ∙ 4𝜋
𝑉𝑠
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
- 25 -
a function of peak cylinder pressure, mean piston speed and mean piston speed [GT
Power].
𝐹𝑀𝐸𝑃 = 𝐶 + (𝑃𝐹 ∙ 𝑃𝑚𝑎𝑥 ) + (𝑀𝑃𝑆𝐹 ∙ 𝑆𝑝𝑒𝑒𝑑𝑚𝑝 ) + (𝑀𝑃𝑆𝑆𝐹 ∙ 𝑆𝑝𝑒𝑒𝑑𝑚𝑝 2 )
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-tocylinder and cyclic variations can strongly effect the IMEP measured when comparing
to BMEP.
8.4.2 Test Results
8.5 Mass Fraction Burned
8.5.1 Definition and Background
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
- 26 -
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 essentially 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 𝑉𝑖 𝑛 = 𝑝𝑗 𝑉𝑗 𝑛
Hence the pressure change due to a change in volume is given by
𝑛
𝑉𝑖
∆𝑝𝑣 = 𝑝𝑗 − 𝑝𝑖 = 𝑝𝑖 [( ) − 1]
𝑉𝑗
And the pressure change due to combustion is given by
𝑛
𝑉𝑖
∆𝑝𝑐 = 𝑝𝑗 − 𝑝𝑖 ( )
𝑉𝑗
The subscript i and j denotes the start and end of the interval respectively. The
exponent n is referred to as the polytropic expansion/compression constant and is
found from the linear gradient constrained within the compression and power strokes
of the logP-logV diagram. The polytropic constant is commonly found to be 1.3
(±0.05) for both compression and expansion processes for conventional fuels.
Heywood (1988). 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.
- 27 -
∆𝑝𝑐 ∗ = ∆𝑝𝑐 ∙
𝑉𝑖
𝑉𝑇𝐷𝐶
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)
∑𝑖0 ∆𝑝𝑐 ∗
= 𝑁
𝑚𝑏(𝑡𝑜𝑡𝑎𝑙) ∑0 ∆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.
𝜃 − 𝜃0 𝑚+1
𝑥𝑏 = 1 − exp [−𝑎 (
)
]
∆𝜃
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.
8.5.2 Test Results
8.6 Coefficient of Variance
8.6.1 Definition and Background
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
- 28 -
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.
𝐶𝑂𝑉𝑖𝑚𝑒𝑝 =
𝜎𝑖𝑚𝑒𝑝
∙ 100
𝑖𝑚𝑒𝑝
̅̅̅̅̅̅̅
𝑁
𝜎𝑖𝑚𝑒𝑝
1
=√
∑(𝑝𝑖 − 𝑝̅)2
𝑁−1
𝑖
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.
8.6.2 Test Results
Measuring Swirl
Measuring Tumble
9. RICARDO WAVE
9.1 Port Modelling
Now that the data is collected experimentally, the pressure loss within the ports due to
the flow over the port surface and changes in flow momentum from the port geometry
can be modelled into WAVE. Either coefficients discussed in section 7.5.2 can be
entered into the flow coefficients profile but WAVE preferably uses Cd as internally,
if Cf is used, WAVE converts the values to Cd using the valve reference diameter
specified by the user within the valve lift editor. To accurately model the effects of
reverse flow, the reverse flow coefficient is recorded, calculated and entered by the
same means of achieving the forward flow coefficients with the flow direction
reversed.
- 29 -
To eliminate the effects of “duplicating” the pressure loss within the model, the ducts
and junctions used to simulate the ports should not contribute any additional pressure
losses and therefore all friction and pressure loss multipliers should be set to 0 and
discharge coefficients set to 1. Any duct bend angles should be set to 0 since the
pressure loss due to the ports geometries have been accounted for from the port flow
tests.
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 Acf, Bcf, Ccf, Qcf are the Chen Flynn coefficients inputted
by the user, Pmax is the peak cylinder pressure, Sfact is the speed factor, RPM is the
engine speed and stroke is the cylinder stroke.
𝑛𝑐𝑦𝑙
2
𝐹𝑀𝐸𝑃 = 𝐴𝑐𝑓 + ∑ [𝐵𝑐𝑓 (𝑃𝑚𝑎𝑥 )𝑖 + 𝐶𝑐𝑓 ∙ (𝑆𝑓𝑎𝑐𝑡 ) + 𝑄𝑐𝑓 ∙ (𝑆𝑓𝑎𝑐𝑡 ) ]
𝑖
𝑖=1
With
𝑆𝑓𝑎𝑐𝑡 = 𝑅𝑃𝑀 ∙
- 30 -
𝑠𝑡𝑟𝑜𝑘𝑒
2
𝑖