CHAPTER 1
INTRODUCTION
1.1 BACKGROUND
The Internal Combustion (IC) engine has evolved over the years since the four stroke ‘OTTO’ cycle was developed in 1897 by Nikolaus August Otto. The basic design and working principle has not changed much and has been the heart of every automobile. Through continuous research and development on the IC engine, performance and efficiency have improved significantly over the years. However, global emissions regulations and the shortage of fossil fuel resources have provided an impetus for engine designers and engineers to provide more advanced technologies, producing cleaner and more efficient motors.
Several alternative technologies such as fuel cells and electric vehicles have been introduced in the market, but they come with associated problems that include high cost, changes required to the fuelling infrastructure and lack of development to support these technologies. Lately, various technologies like catalysts and intelligent engine management systems have contributed to the achievement of lower emissions and fuel consumption without compromising on output performance. Technologies such as Variable Valve Actuation (VVA), direct fuel injection, and cylinder activation have had a significant impact on enhancing overall performance of the automobile. The future of automotive research and development is about improving the bottom line, source of power, i.e. Combustion of fuel.
In recent times, alternative combustion technology such as Homogenous Charge Compression Ignition (HCCI) has been studied and results have been positive. HCCI combustion has the potential to reduce
fuel consumption and NOx emissions pertaining to the most stringent of legislation of both present and future. HCCI technology is attractive as there is no need for major modifications to the existing structure of IC engines and with significantly low NOx emissions, after treatment systems are not required.
1.1.1 About HCCI
HCCI combustion is achieved by premixing the air-fuel mixture, either in the manifold or by early direct injection, as in Spark Ignition (SI) engine, and compressing the mixture until the temperature inside the combustion chamber reaches the auto ignition point and ignites, as in Compression Ignition (CI) engine. It is also known by following terms –
Controlled Auto Ignition (CAI)
Active Thermo Atmosphere Combustion (ATAC)
Premixed Charge Compression Ignition (PCCI)
Homogenous Charge Diesel Combustion (HCDC)
Premixed Lean Diesel Combustion (PREDIC)
Compression Ignited Homogenous Charge (CIHC)
HCCI engine combines the advantages of Spark Ignition engine (Homogenous Charge) and Compression Ignition engine (increased efficiency) with reduced emissions. Auto ignition combustion can be described by the oxidation of the fuel dependant solely on chemical reactions, governed by chain-branching mechanisms. According to various researchers, auto ignition in an HCCI engine is a random phenomenon that starts throughout the combustion chamber, while others argue that is a more uniform process. Hence, further study and investigation of auto ignition process is required so as to gain an understanding on controlling HCCI combustion. The following figure depicts the differences in Spark Ignition, Compression Ignition and HCCI engine:
Figure 1 – Combustion difference between SI, CI and HCCI combustion [1]
1.1.2 Controlling Factors
There are several controlling factors that affect the performance of HCCI combustion engine and currently research is being carried out globally to understand the relationship between these controlling factors. The relationships will enable designers and engineers to have a better understanding of the combustion process and overcome the following challenges to a certain extent:
Ignition Timing Control
Combustion Rate Control for High Load Operation
Engine Cold Start
Development of Emissions Control Systems
Achieving Satisfactory Engine Transient Operation
Development of Engine Control Strategies and Systems
Development of Appropriate Fuel Systems
Overcoming Multi Cylinder Engine Effects
However, the major obstacle for HCCI engine to become a commercial success is combustion control. For this reason, several methods have been proposed for achieving HCCI engine control over the wide range of operating conditions required for typical transportation engine applications. Control technologies reported in the literature have demonstrated some degree of success, but further R&D efforts are required [2]. Some of the proposed methods include:
Variable Compression Ratio
Variable Valve Timing
Ignition Enhancing Additives
CHAPTER 2
LITERATURE REVIEW
Onishi et al. in 1979 [3] were amongst the first researchers to investigate the possibility of using auto ignition combustion as a combustion mode in an engine. They have applied auto ignition combustion using gasoline in a two-stroke gasoline engine and named this process ATAC. They showed that there was very small Cycle-By-Cycle Variations (CBCV) in the peak combustion pressure, the reaction occurred spontaneously at many points and combustion proceeded slowly. They investigated the significance of the hydroxyl, OH, hydrated carbon and diatomic carbon radicals and showed that their concentration was significantly higher and that the radicals had a longer life than in a SI engine (40° life compared to 25°). They suggested that to attain ATAC, the quantity of the mixture and the A/F ratio must be uniform from cycle to cycle, the temperature of the mixture must be suitable and the cyclic variability of the scavenging process must be kept to a minimum to ensure the correct conditions of the residual gases remaining in the combustion chamber. They obtained satisfactory combustion over a wide range of A/F from 11 to 22 and they concluded that ATAC reduces both fuel consumption and exhaust emissions over the whole of that range.
Around the same time, the auto ignition and energy release processes of CIHC combustion and what parameters affect them were investigated using a single-cylinder four-stroke cycle Waukesha Cooperative Fuel Research (CFR) engine with a pancake combustion chamber and a shrouded intake valve [4]. It was deduced that this controlled auto ignition/ combustion mode was not associated with knocking but a smooth energy release that could be controlled by proper use of temperature and species concentrations. In their experiments they controlled independently the intake charge temperature (600-810K) and the recirculated
exhaust products (35-55% EGR), which were evaluated using carbon dioxide measurements. They used three different fuels;
70% iso-octane and 30% n-heptane
60% iso-octane and 40% nheptane
60% iso-propylbenzene and 40% n-heptane From the above research, it was concluded that:
Chemical species in the EGR gases had no effect on the rate of energy release and therefore EGR was primarily used to control combustion by means of regulating the initial gas temperature
Delivery ratio affected the combustion process through changes in the concentrations of fuel and oxygen in the reacting mixture. Therefore, at high delivery ratios the energy release became violent and for a CR of 7.5:1, it was found that a delivery ratio of 45% was the maximum
Fuels with lower octane numbers were ignited more easily.
In 1989, Thring [5] investigated the possibility of autoignition combustion in a single-cylinder, four-stroke internal combustion engine by Labeco CLR and was the first to suggest using SI operation at high loads and HCCI at part load. Even though the term ATAC and CIHC were previously used to describe this auto ignition/combustion process, Thring used the term HCCI. Intake temperatures (up to 425°C), equivalence ratios (0.33-1.30), EGR rates (up to 33%) and both gasoline and diesel were used to explore the satisfactory operation regions of the engine. There were three regions of unsatisfactory operation labelled “misfire region”, “power-limited region” and “knock region.” In the misfire and knock region the mixture was too rich while in the power-limited region the mixture was too lean. It was concluded that, under favourable conditions, HCCI combustion exhibited low cyclic variability and produced fuel economy results comparable with a diesel engine. However, high EGR rates (in the range of 30%) and high intake temperatures (greater than 370°C) were required.
HCCI combustion was later on also tested in a production engine [6] by using a 1.6 litre VW engine which was converted to HCCI operation with preheated intake air. By using λ=2.27, a CR of 18.7:1 and preheating the intake air
up to 180°C, an increase in the part load efficiency 124 Advances in Internal Combustion Engines and Fuel Technologies from 14 to 34% was achieved. A NiCE-10 two-stroke SI engine with a CR of 6.0:1 was also used [7] to investigate this auto ignition phenomenon by measuring the radical luminescence in the combustion chamber using methanol and gasoline as fuels. Luminescence images were acquired using an image intensifier coupled with a Charge-Coupled Device (CCD) camera and the luminescence spectra of the radicals OH, CH and C2 were acquired by using a band-pass filter in front of the Ultra Violet (UV) lens. With conventional SI combustion, radical luminescence indicated a flame propagating from the centre of the spark plug towards the cylinder walls, while with ATAC combustion, radical luminescence appeared throughout the combustion chamber. The total luminescence intensity exhibited with ATAC combustion was less compared to SI combustion. Furthermore, with SI combustion OH radical species were formed 30° Crank Angle (CA) Before Top Dead Centre (BTDC) and assumed that it occurred at the same timing as the main combustion process, while in the case of ATAC combustion, OH radical species increased before the main combustion process as indicated by the rate of heat release.
The effect of CRs ranging from 10:1 to 28:1 on various fuels was extensively studied [8],[9]. VCR can be achieved using a modified cylinder head that its position can be altered during operation using a hydraulic system. NOx and smoke emissions were not affected by CR and were generally very low. However, an increased CR resulted in higher HC emissions and a decrease in combustion efficiency [8]. Others [10] reported that decreasing inlet temperatures and lambdas, higher CRs were need to maintain correct maximum brake torque and concluded that variable CR can be used instead of inlet heating to achieve HCCI combustion. Furthermore, the effect of CR on HCCI combustion in a direct-injection diesel engine was also investigated [11]. The CR could be varied from 7:5:1 to 17:1 by moving the head and cylinder liner assembly relative to the centreline of the crankshaft. Acceptable HCCI combustion was achieved with ignition timing occurring before TDC – with misfire being exhibited if ignition timing was further delayed – with CRs from 8:1 to14:1. However, with a knocking intensity of 4
(where audible knock occurs at 5 on a scale from zero to ten), the acceptable HCCI operation was limited at CRs from 8:1 to 11:1.
Supercharging (2bar boost pressure) was shown to increase the Indicated Mean Effective Pressure (IMEP) of an engine under HCCI combustion to 14bar [12]. Supercharging was used because of its capability to deliver increased density and pressure at all engine speeds while turbocharging depends on the speed of the engine. However, this resulted in lower efficiency due to the power used for supercharging. Supercharging resulted in greater emissions of CO and HC, greater cylinder pressure, longer combustion duration and lower NOx emissions. There were no combustion related problems in operating HCCI with supercharging and the maximum net indicated efficiency achieved was 59%. On the contrary, others [13] investigated the effect of turbo charging on HCCI combustion. A Brake Mean Effective Pressure (BMEP) of 16bar (compared to 6bar without turbo charging and 21bar with the unmodified diesel engine) and an efficiency of 41.2% (compared to 45.3% with the unmodified diesel engine) were achieved. Furthermore, CO and HC emissions decreased with increasing load, but NOx emissions increased. However, at higher loads, as the rate of pressure increased and the peak pressure approached their set limit (i.e. peak pressure greater than 200bar), ignition timing was retarded at the expense of combustion efficiency. Thus, in order to improve the combustion efficiency at high boost levels, cooled EGR rates was introduced [14], and it was shown that under those conditions, the combustion efficiency increased only slightly.
Even though EGR has been employed by various researchers, the results are not always consistent within the research community. Depending on the method of EGR used (trapped exhaust gases due to valve timing, or exhaust gases re-introduced in the manifold), the results can vary, since both the temperature and chemical species present might not be the same in all cases.
Both aforementioned methods were employed [15],[16] where the first method relied on trapping a set quantity of exhaust gas by closing the exhaust
valves relatively early, while in the second method, all the exhaust gases were expelled during the exhaust stroke, but during the intake stroke, both the inlet and exhaust valves opened simultaneously, to draw in the engine cylinder both fresh charge and exhaust gas. It was shown that HCCI combustion is possible with EGR and without preheating the inlet air and that increasing the quantity of exhaust gases advances the ignition timing. Furthermore it was concluded that HCCI can become reproducible and consistent by controlling the ignition timing by altering the EGR rate. Others achieved EGR [17],[18] by throttling the exhaust manifold, which increased the pumping work and reduced the overall efficiency. They concluded that:
With increasing EGR, and thus decreasing A/F ratio and slower chemical reactions, the inlet gas temperature must also be increased
With increasing amounts of EGR, the combustion process becomes slower, resulting in lower peak pressure and lower rate of heat release and therefore longer combustion rates.
Both the combustion and gross indicated efficiencies increased with increasing EGR.
Based on further work [19], it was concluded that EGR had both thermal and chemical effects on HCCI combustion and that active species in the exhaust gases promoted HCCI. Others [20] however, reported contradicting results, where varying the EGR had little effect on combustion timing, on gross IMEP, combustion efficiency and net indicated efficiency. However, in those cases, the EGR was taken from the exhaust pipe and through a secondary pipe re-introduced in the inlet pipe where it was mixed with the fresh air mixture. There was no indication of pipe insulation or of the temperature of the EGR gases. Therefore, if the temperature of the gases was lower or of the same order as the intake gas temperature, then the effect of the EGR might have been reduced to only dilution effects.
Others on the other hand [21], investigated the importance of EGR stratification on HCCI combustion. It was found that HCCI combustion started
near the centre of the combustion chamber at the boundary between the hot exhaust gases, situated at the centre due to poor scavenging characteristics of the valves, and the fresh intake charge. The importance of the mixing of the EGR and the fresh-air mixture was identified, since by controlling the EGR stratification, the combustion timing might also be controlled. The effect of homogeneous and inhomogeneous cooled EGR on HCCI combustion has also been investigated [22]. For the homogeneous case, the fresh air and EGR gases were mixed upstream of the intake port and thus well-mixed before the fuel injector.
For the inhomogeneous case, EGR gases were introduced downstream the fuel injector and therefore there was no time for proper mixing. With inhomogeneous EGR supply, auto ignition timing was advanced (due to local hot spots of fresh air-fuel mixture) but the overall combustion was slower (due to local cold spots of exhaust gas-fuel mixture), than with homogeneous EGR supply.
Fuel injection strategies is one of the most important topics under research for HCCI combustion, as it can be easily controlled, compared to VCR, multiple fuel injection, etc, to alter HCCI combustion, by varying the injection timing and duration, and the injector location and type. It was shown that even injector nozzle optimizations can be employed to alter the fuel spray and affect HCCI combustion [23]. Injector location was also investigated [24] by using both port injection – to create a premixed fuel-air mixture – and direct injection – to control the timing of auto ignition. Others [25], focused on different fuel injection strategies; injecting the fuel in a 20 litre mixing tank before the engine intake port and injecting the fuel just outside the engine intake port. The first treatment resulted in a homogeneous mixture, while the second treatment resulted in a mixture with fluctuations of the order of 4 to 6mm. Regardless of the preparation method; however, combustion was inhomogeneous with very large spatial fluctuations. Furthermore, the local combustion kernels did not have a tendency to be more frequent in the central part of the combustion chamber, where the temperature was assumed to be higher than
in the vicinity of the walls. They were unable though to identify the process that caused the very inhomogeneous combustion initiation.
Others also investigated the effect of various injection patterns and their combination on HCCI combustion. In particular [26], the following three fuel injection patterns were investigated: (i) Injection during the negative valve overlap interval to cause fuel reformation, (ii) injection during the intake stroke to form a homogeneous mixture and (iii) injection during the compression stroke to form a stratified mixture. It was found that with fuel reformation, the operating range of HCCI combustion was extended without an increase in the NOx emissions. Furthermore, limited operation was observed with late injection timing that also led to high NOx emissions. Two other injection systems were also employed [27]: (i) a premixed injection injector in the intake manifold to create a homogeneous charge and (ii) a DI injector to create a stratified charge. Thus by varying the amount of fuel injected through the DI injector (from 0 to 100%) and varying the injection timing of the DI injector (from 300 to 30°CA BTDC) different stratification levels were achieved. It was found that HCCI combustion was improved at the lean limit with charge stratification, while CO and HC emissions decreased. On the contrary, at the richer limit, a decrease in combustion efficiency was evident at certain conditions. It was concluded that charge stratification causes locally richer regions that, in the lean limit, improved combustion efficiency by raising the in-cylinder temperature during the early stages of the auto ignition process, while at the rich limit, the change in the in-cylinder temperature does not affect the combustion efficiency to such an extent.
The possibility of using a Gasoline Direct Injection (GDI) injector and varying the injection timing to control HCCI combustion has also been investigated [28]. It was concluded that the most homogeneous mixture was formed with early injection timings, while fuel inhomogeneities (and thus regions with richer fuel concentration) were present with retarded injection timings. With retarded injection timing and thus increased fuel inhomogeneity, combustion of locally richer mixtures caused an increase in the combustion temperature that as a
result, caused higher combustion efficiency, an increase in NOx emissions but a decrease in CO and HC emissions. Furthermore, with late retarded injection timings, a decrease in the combustion efficiency (and increase in the CO and HC emissions) was observed due to fuel impingement on the piston surface. It was concluded that fuel stratification can be used to improve HCCI combustion under very lean conditions but that great care is needed to avoid the formation of NOx due to locally near-stoichiometric fuel concentrations.
CHAPTER 3
MATHEMATICAL MODELLING
Mathematical modelling is a cheaper, faster and efficient way to gain an insight into the working of a system. It requires great deal of skills and resources to setup an experimental facility for HCCI combustion testing. However, with knowledge gained from various experiments conducted on HCCI combustion, it is possible to formulate a mathematical model and derive a set of relationships between various input parameters and output results. This model can further be fine tuned by validating against known experiment results or by conducting an experiment using exactly similar set of control parameters, as used in simulation model. The following section provides details on formulation of a mathematical model for the purpose of simulation of HCCI combustion. There are mainly two types of mathematical models that are used to describe the physical phenomena occurring within the engine cylinder:
3.1. PHENOMENOLOGICAL MODEL (BLACK BOX MODEL)
A phenomenological model or an empirical model is derived using experimental data only, using no prior information about the system, i.e. engine cylinder during HCCI combustion. Statistical principles are used to derive relationships among sensitive parameters affecting the final result. For such models, an experiment is set up where input parameters are controlled and output is measured. Input parameters or experimental setup is altered to study the corresponding effect on output result. From results of experiment, graphs are charted showing correlation between input parameters and measured results. Phenomenological models are more general and applicable to many different kinds of problems. However, they provide less insight into the problem or its possible solution and less predictive capability.
3.2. MECHANISTIC MODELS (GREY BOX MODEL)
A mechanistic model is developed using prior information about the system, i.e. results of previous work conducted on similar topic. They provide deeper understanding and more accurate prediction as compared to phenomenological models. Globally, research on HCCI combustion engine has been undertaken by almost every academic institution, automobile manufacturers, and consultants since last 3 decades. Also, from the detailed research done by the author, the information gathered so far on HCCI combustion, a glimpse of which is mentioned in Literature Review section, is sufficient to develop a mechanistic model on HCCI combustion.
Modeling, in science and engineering, may be generally regarded as the process of describing the physical phenomena in a particular system with the help of mathematical equations (subject to “reasonable” assumptions) and solving the same to understand more about the nature of such phenomena. Usually, engineering models help in designing better devices by understanding more about the fundamental physical processes occurring therein. Engine modeling activities, at least in recent decades, have largely been concentrated in the direction of designing better performing engines with lower emissions. In this regard, modeling of engine combustion processes assumes importance. The various engine combustion models that have been developed to date may be grouped into three categories:
Zero dimensional models
Quasi-dimensional models
Multi-dimensional models
In the above classification, although the level of detail and proximity to physical reality increases as one proceeds downward, so does the complexity of creating and using those models.
Zero dimensional models are the simplest and most suitable to observe the effects of empirical variations in the engine operating parameters on
overall heat release rates/cylinder pressure schedules. These models are zero dimensional in the sense that they do not involve any consideration of the flow field dimensions. Zero dimensional models are further sub-divided into:
Single zone models
Two zone models
Multi-zone models
In single zone models, the working fluid in the engine is assumed to be a thermodynamic system, which undergoes energy and/or mass exchange with the surroundings and the energy released during the combustion process is obtained by applying the first law of thermodynamics to the system.
In two zone models, the working fluid is imagined to consist of two zones, an unburned zone and a burned zone. These zones are actually two distinct thermodynamic systems with energy and mass interactions between themselves and their common surroundings, the cylinder walls. The mass-burning rate (or the cylinder pressure), as a function of crank angle, is then numerically computed by solving the simplified equations resulting from applying the first law to the two zones.
Multi-zone models take this form of analysis one step further by considering energy and mass balances over several zones, thus obtaining results that are closer to reality.
3.2.1. Assumptions for a typical two zone model
The burned and unburned zones are ideal gases of different properties.
The unburned zone is assumed to consist of a premixed fuel-air mixture. Though this may not be exact for diesel combustion, it is more realistic for SI engine combustion.
The characteristic gas constants of the burned and unburned zones do not vary much with temperature and pressure; or if any such variations
exist, they can be suitably modeled using explicit relationships between gas constants and properties (T, P, etc.)
No heat transfer occurs from the burned to the unburned zone and vice versa.
Enthalpy associated with injected fuel is usually not significant and hence ignored.
Crevice losses may be significant but are not included.
Spatially averaged instantaneous heat transfer rates are adequate to estimate heat transfer to the cylinder walls.
Instantaneous pressure in both the zones is the same since the flame is a deflagration combustion wave.
The work required to transfer fluid from the unburned zone to the burned zone is negligible.
The single zone and two zone models have been traditionally used in two different directions –
In one way, both these models have been used to predict the in-cylinder pressure as a function of crank angle from an assumed energy release or mass burned profile (as a function of crank angle).
Another use of these models lies in determining the energy release/mass burning rate as a function of crank angle from experimentally obtained in-cylinder pressure data.
3.3. MODELLING PLATFORM
Various commercial packages have been developed and are available to solve engineering problems related to design and optimisation of internal combustion engines (ICEs). There are four primary engine simulation commercial packages used in the automotive industry today: Ricardo Wave (RW), Lotus Engine Simulation (LESoft), AVL fire, and GT-Power. These packages are similar in purpose and functionality. They require detailed input parameters to simulate the engine operation in an integrated manner rather than using different subsystems.
LESoft is an in-house code developed by Lotus Engineering. The package processes engine simulation in two modules, the data module and the solver module. The data module allows user to input the engine dimension data. The solver module is a built in combustion and heat transfer zero-dimensional equations and fuel/gas composition solver according to user input data in data module. The code is able to predict gas flow, combustion and overall performance of ICEs.
RW is an engine simulation package designed to analyze the dynamics of pressure waves, mass flows, energy losses in ducts, plenums, and manifolds of various systems and machines. RW provides simulation of time-dependent fluid dynamics and thermodynamics using two-zone model.
However, software costs generally prohibit use in small organizations, primarily making them industry-specific software packages. Moreover, commercial software packages are based on computation fluid dynamics (CFD) since they are designed to improve mechanical aspects of the engine. While mechanical models are used to calculate the moving parts of the engine in order to obtain the engine torque and acceleration, control models are used to allow calculations in feedback control schemes to optimise engine performance, such as variable valve timing, ignition timing, air to fuel ratio, and other variable engine geometries systems. Open source packages have been developed by small research groups to solve specific issues.
MATLab is widely used software used by majority of researchers to simulate the internal combustion engine. However, hardware requirements and training limitations, combined with project time constraint of the research work undertaken by the author, renders any of the above software packages difficult to use. While above simulation platforms are primarily based on computation fluid dynamics, simple mechanistic models, can be built and developed using any software package offering computation ability to solve engineering problems, program event based automation and
a graphic tool to visualize outcome, such as Visual Basic for Applications (VBA) with Excel spreadsheet.
3.4. SUBMODELS
Several submodels have been used or developed to describe all relevant incylinder processes and are vital for the main model to work. The submodels used for the HCCI combustion modeling are as per below – 3.4.1. Cylinder Geometry: Cylinder geometry includes bore diameter, stroke,
compression ratio and connecting rod length. These attributes determine the basic structure to the simulation model.
3.4.2. Piston Motion: This submodel is also known as crank-slider model. The instantaneous position of the piston in the cylinder is evaluated from this submodel. From the instantaneous position of the pistion, the instantaneous volume of the combustion chamber is also determined, as a function of crank angle.
3.4.3. Air and fuel properties: For the purpose of simulation, air fuel mixture is assumed to have same properties as air. Properties of air such as gas constant for air, ratio of specific heat capacities for air, and fuel properties such as Lower heating value of the fuel and stoichiometric air to fuel ratio are required to calculate pressure and temperature difference as piston moves inside the cylinder.
3.4.4. Engine Cycle: HCCI combustion works on Otto cycle as compression (heat addition) and expansion (heat release) happens at constant volume. Therefore, the changing chemical state of the air fuel mixture and changing thermodynamic state of the cylinder are depicted using Otto cycle equations. Based on previous knowledge of Otto cycle, certain points like peak pressure and start of combustion are fixed, in terms of crank angle degrees.
To maintain cycle to cycle consistency, EGR percentage would be varied using various degrees of negative valve overlap. Mass flows through open valves will be calculated by one dimensional compressible flow equations for flow through a restriction of filling and emptying models.
3.4.5. Heat Release Rate: The heat release rate is the amount of heat released from the chemical reaction with respect to crank angle degree. Pressure and corresponding values of temperature are calculated for unit increment in crank angle degrees. Heat release rate with respect to crank angle degree is calculated using the difference in instantaneous temperature and temperature at fixed points in the engine cycle. The heat release rate calculated is the gross value and when divided by the specific content of the fuel, combustion reaction rate is obtained.
3.4.6. Heat Transfer: Heat transfer occurs through conduction, convection and radiation from hot burned gases to piston head, inlet and exhaust valves, cylinder walls, cylinder liners and coolant. For physical testing, water is circulated through the cooling channels as engine runs on the test rig. The difference between inlet temperature and outlet temperature is used to calculate heat transfer, i.e. heat energy dissipated and could not be used to extract work.
3.4.7. Valve motion: This submodel describes the effect of valve motion on the final output of the combustion model. For the purpose of HCCI combustion modelling, internal EGR is used using Negative Valve Overlap (NVO). To vary the percentage of EGR, the duration of NVO is varied and practically this is possible via variable valve technologies. An effective variable valvetrain enables the engine to breathe smoothly to increase the volumetric efficiency, while allowing the engine to operate on lean mode at low load conditions.
CHAPTER 4
MODELLING APPROACH
As mentioned earlier, internal combustion engine is a very complex system, and therefore it is fairly difficult to create a combustion model that is simple yet complete. The modelling approach has been to first create a simple model based on first principles, and then gradually add various submodels into the mix, thereby increasing the accuracy. Through previous researches conducted in HCCI, it is known that heat addition in HCCI occurs at constant volume [29]. Therefore, the HCCI combustion cycle can be modelled on Otto cycle.
4.1. STARTING PARAMETERS
The starting parameters are the basic values needed to create a simulation structure. In order to maintain simplicity, the simulation structure is first made in Microsoft Excel spreadsheet. Following table shows initial parameters, along with their values in SI units –
Table 1 – Engine Geometry Details
The engine geometry parameters define the physical size of the system to be modelled. Values of bore and stroke are assumed based on standard production engines. Compression ratio is assumed to be as per Otto cycle, which can be further modulated to study the effects of compression ratio on final outcome. Displaced volume is the total volume that can be filled by air fuel mixture. It is calculated from bore and stroke as per below –
Vd=π b
2
s 4
Clearance volume is the instantaneous volume in the cylinder when piston is at Top Dead Centre (TDC). Combustion process happens within clearance volume after which due to rapid increase in temperature and pressure, piston is pushed downward, i.e. mechanical work done by the system.
Table 2 – Constants (Ideal Gas)
The above table shows values of constants that are frequently used in the calculation. The ambient, i.e. atmospheric temperature and pressure are assumed to be 1 bar and 300 K. For the purpose of simplicity the fresh charge of air and fuel mixture is assumed to be having same properties as ideal gas. Hence, gas constant and specific heat at constant pressure and constant volume of reactant is same as air. The ratio of specific heat capacities (γ) is a dimensionless quantity, known as the isentropic expansion factor.
γ=Cp
Cv
The lower heating value of fuel is an indication of the energy stored in the fuel, which is extracted by combusting the fuel air mixture under high pressure. Every fuel has a characteristic stoichiometric air to fuel ratio (AFR), i.e. number of moles of air required to enable combustion to happen under standard temperature and pressure (STP) conditions for 1 mole of fuel. For practical purposes, the ratio is taken in terms of mass (Kg/Kg). Equivalence ratio ( ) is the ratio of Actual AFR to stoichiometric AFR. Inᴓ real life, equivalence ratio is always more than 1 so as to ensure complete combustion of fuel [29]. Self ignition temperature is the temperature at which fuel gets ignited without any external spark. At self ignition temperature, the kinetic energy of molecules is high enough to collide with molecules of air and initiate combustion. Mass of reactant, i.e. fresh charge inside the cylinder is calculated from standard gas equation –
PV =mRT
By rearranging terms in Eq. 3, expression for mass is obtained as per below -
m=PV
RT
Values of pressure and temperature are taken at ambient point, as mentioned in Table 2. From AFR, mass of air and mass of fuel present in the cylinder at the time of Inlet Valve Closing (IVC) are evaluated.
4.2. COMBUSTION CYCLE
As mentioned earlier, in HCCI heat addition occurs at constant volume, and therefore for HCCI combustion the ideal cycle would be Otto cycle. For Otto cycle following assumptions are made –
The engine operates in a closed loop
The working fluid is air
The air behaves as an ideal gas
All processes are reversible
The combustion process can be replaced by an external heat source
The exhaust process can be replaced by a heat rejection process
4.2.1.Properties at the State Points in the Cycle State 1
Taking the starting point for the cycle as the start of the compression process then from the inlet conditions:
in in T T p p 1 1
Assuming that the gas behaves as an ideal gas:
1 1 1 p RT v
Tthe mass in the cylinder is given by:
1 1 v V m State 2
From the definition of the compression ratio
r v v V V v v r 1 2 2 1 2 1
Since the process from 1 to 2 is an isentropic (reversible and adiabatic) compression process then:
r p v v p p v p v p 1 2 1 1 2 1 1 2 2
Thus the temperature can be found:
1 1 1 2 1 1 2 1 1 1 2 2 2 1 1 2 2 r T v v T T v v RT v v RT v p v p State 3
The process from 2 to 3 is a constant volume heat addition process, thus: r v v v 1 2 3
Applying the first law and noting that no work is done during this process:
v v v c q r T c q T T T T c q u u w q 23 1 1 23 2 3 2 3 23 2 3 23 23 Then assuming that the gas behaves as an ideal gas:
1 23 1 1 3 1 1 3 3 3 3 T c rq r p RT RT p r v rRT v RT p v State 4
At state 4 the piston has returned to its initial position thus: 1
4 v
The process from 3 to 4 is an isentropic (reversible and adiabatic) expansion process thus:
1 23 1 1 3 3 4 3 3 4 3 3 4 4 1 1 T c q r p r p r p v v p p v p v p v
Thus the temperature can be found:
v c q r T r T v v T T v v RT v v RT v p v p 23 1 1 1 3 1 4 3 3 4 3 3 3 4 4 4 3 3 4 4 4.2.2. Energy Exchanges Process 1 to 2
For the compression process from state 1 to state 2 the moving boundary work done is:
1
1 2
2 1
1 1 1 1 1 1 1 1 1 1 1 1 2 1 1 1 1 1 1 1 2 1 1 1 1 1 1 1 2 1 1 2 1 1 1 1 2 1 1 1 2 1 12 1 1 1 1 1 1 1 1 u u T T c T c r T c r RT r v p v v v v v p v v v v p v v v p v v p dv v v p pdv w v v v
Since the process is isentropic then no heat is transferred with the surroundings. Process 2 to 3
For the heat addition process from state 2 to 3 the heat transfer is specified and the work transfer is zero, since there is no change in volume during the process.
Process 3 to 4
For the expansion process from state 3 to state 4 the moving boundary work done is:
3
3 4
4 3
1 3 1 3 1 3 3 1 3 1 3 1 4 1 3 3 3 1 3 1 4 1 3 3 3 1 3 1 4 3 3 4 3 1 3 3 4 3 3 3 4 3 34 1 1 1 1 1 1 1 1 u u T T c T c r T c r RT r v p v v v v v p v v v v p v v v p v v p dv v v p pdv w v v v
Since the process is isentropic then no heat is transferred with the surroundings. Process 4 to 1
For the heat rejection process from state 4 to 1 the work transfer is zero, since there is no change in volume during the process. Applying the first law:
23 1 41 23 1 1 1 4 1 41 4 1 41 41 q r q c q r T c T c T T c q u u w q v v v v CycleSince the start and end of the cycle are the same state point then there is no net gain or loss of energy by the system. Thus applying the first law
q23q41
w12w34
0The net work done by the system can be expressed in terms of the heat transfers as:
1 23 1 23 23 41 23 34 12 1 r q r q q q q w w wnetThe thermal efficiency of the cycle is given by:
1 23 1 r q wnet th
p T v 1 p1 pin T1 Tin 1 1 1 p RT v 2 p pr 1 2 1 1 2 Tr T r v v 1 2 3 3 3 3 v RT p v c q T T 23 2 3 3 2 v v 4 p p r 3 4 1 3 4 Tr T v4 rv3
Table 4 - State properties in terms of the previous state’s properties.
p T V 1 p1 pin T1 Tin 1 1 1 p RT v 2 p p r 1 2 1 1 2 r T T r v v 1 2 3 1 23 1 3 T c rq r p p v v c q r T T 1 23 1 3 r v v 1 3 4 1 23 1 1 4 1 T c q r p p v v c q r T T 23 1 1 4 v4 v1
Table 5 - State properties in terms of the properties at state 1.
In HCCI combustion, there is no direct way to initiate combustion; hence it is important to establish the controlling parameter that can be explicitly monitored and a closed loop control system can be designed around the controlling parameter. Now, for an internal combustion engine, it is desirable to have peak pressure after Top Dead Centre (ATDC) to avoid waste of energy in pushing against the direction of piston travel. As per previous research, peak pressure normally happens between 10 CAD to 30 CAD ATDC [29]. Therefore, the crank angle degree (CAD) at which peak pressure is achieved is first fixed to be within the above mentioned range. Therefore, every cycle input parameters have to be modulated in order to maintain cycle to cycle consistency with respect to peak pressure, and as a result mean effective pressure (mep), torque and power values are also consistent.
Also, peak pressure is obtained when combustion process is 90 % over [1], therefore the peak pressure point and 90 % combustion completion CAD should be within 5 to 10 CAD of each other. The burn duration is the time of combustion reaction, i.e. from start of combustion (SOC) to end of combustion (EOC). To model burn duration, Wiebe Function is used –
MFB (θ)=1−exp
[
−a(
θ−θ0∆ θ
)
m +1
]
[30]
MFB (θ) = Mass Fraction Burned at corresponding CAD (θ) θ = corresponding CAD
θ0 = CAD corresponding to SOC
∆θ = CAD corresponding to burn duration m = constant (typical value = 3)
a = constant (typical value = 5)
The values of MFB will be within the range from 0 to 1. A sample of Wiebe Function curve is shown below –
Figure 2 – Sample Wiebe Function Curve [30]
4.4. INTERDEPENDENCIES
While modelling any form of combustion it is important to note how different parameters are affecting each other. For example, during combustion phase in internal combustion engine, temperature increases due to progress of chemical reaction. As chemical reaction progress, fuel energy is released in the form of heat, due to which temperature increases rapidly. As a result, chemical reaction rate increases. Thus, temperature and chemical reaction rate are interdependent parameters. Similarly with increase in temperature, pressure also increases, which in turn contributes to increase in temperature.
For HCCI combustion to happen in a production engine, the most suitable technique is to use EGR with negative valve overlap (NVO) [31]. EGR makes the overall mixture more dilute and raises the temperature of reactant mixture. Due to this, the specific heat of mixture at the time of IVC is higher than SI or CI combustion. Now to achieve SOC point in every cycle within + 5 CAD of fixed SOC point, temperature of mixture, i.e. fresh charge + EGR at IVC has to be monitored. By varying EGR %, temperature at IVC can be maintained within a set tolerance. Again by varying NVO duration, EGR % can be increased or decreased
depending on previous cycle. It would be difficult to enable such control on a physical engine using conventional variable valve technology (VVT) of cam changing or cam phasing [32]. This calls for electromechanical control of valves opening and closing, which can be modelled on a simulation platform to further enhance the practicality of such system.
CHAPTER 5
RESULTS & DISCUSSIONS
The simulation structure was created on Microsoft Excel spreadsheet and as per derivations in 4.2. The table containing input parameters and calculation of pressure, temperature and specific volume at interval of 10 CAD is shown in Appendix 1. Following results have been obtained for each state –
r = 10 g = 1.4 q23 = 600000 J/k g cv = 717.75 J/k g K R = 287.1 J/k g K P T v State P T v u Pa K m3/k g Pa K m3/k g J/k g 100000 300.00 0.86130 1 100000 300.00 0.86130 0.0 2511886 753.57 0.08613 2 2511886 753.57 0.08613 325546.9 5298372 1589.51 0.08613 3 5298372 1589.51 0.08613 925546.9 210932 632.80 0.86130 4 210932 632.80 0.86130 238864.3 Process w q q-w uo-ui J/kg J/k g J/k g J/k g 12 -325547 0 325547 325547 23 0 600000 600000 600000 34 686682.6 0 -686683 -686683 41 0 -238864 -238864 -238864 wnet = 361135.7 J/k g nth = 0.601893
Using Table 1 Using Table 2
Table 6: Final Results
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1 1 10 100 1000 10000 100000 Specific Volume [m3/kg] Pressure [kPa]
Figure 3: Pressure vs Specific Volume
-180-120-60 0 60120180 1 10 100 1000 10000 100000 0 500 1000 1500 2000 2500 3000 3500 4000
Crank Angle atdc [deg]
Pressure [kPa] Temperature [K]
CHAPTER 6
CONCLUSION
HCCI combustion is a promising alternative to conventional SI and CI combustion as there is significant reduction of NOx emissions and increased power output. However, as in SI or CI combustion, there is no direct event to initiate combustion because reactant mixture spontaneously ignites when sufficient temperature and pressure conditions are reached. Therefore, several parameters have to be monitored together to make HCCI combustion work in a physical engine with consistent output cycle by cycle. This requires deeper analysis of chemical and physical kinematics. Through mechanistic modelling it is possible to further understand the correlation between various parameters and output.
The model presented in this report is of Otto cycle, which shows sensitivity towards several input parameters. For HCCI combustion, the Otto cycle simulation can be modulated by applying closed loop control for inlet temperature. Also, flow dynamics, which have been excluded for Otto cycle simulation, have to be included for modelling HCCI combustion, as EGR is primary parameter and the percentage of EGR is dependent upon air flow through outlet valve. Moreover, Arrhenius equation also would have to be included to depict the activation energy needed to overcome the reaction barrier.
CHAPTER 7
SCOPE FOR FURTHER RESEARCH
There is an upper limit in terms of engine speed for HCCI combustion, after which detonation starts occurring, which is harmful for the engine. The major factor contributing to detonation is the temperature and pressure inside the cylinder, i.e. if HRR exceeds certain value, the flame speed will travel faster than the speed of sound and result in generation of shock wave. To prevent this generation of shock wave, practical HCCI engines have a limit of around 4000 rpm. There are several ways to approach the solution to this problem, wherein lies scope for further research.
One way is to model pressure rise rate (PRR) in terms of HRR and establish relationship between HRR and PRR at various mixture strengths and engine speeds. The other way is to study the process of deflagration to detonation and try to model the reverse process. Also, flame velocity, i.e. laminar and turbulent can further be studied to understand their effects on physical and chemical process.
APPENDIX 1
APPENDIX 2
HCCI COMBUSTION: MATHEMATICAL MODELLING APPROACH USING VISUAL BASIC FOR APPLICATIONS
VIVEK BHARAT PATTNI1, P. NAVEENCHANDRAN2, C.THAMOTHARAN3, R.
RAJASEKAR4
1M.Tech Scholar, Department of Automobile Engineering, Bharath Institute of Science and
Technology (BIST), Bharath University, Chennai, Tamil Nadu, India [email protected]
2Head of Department, Department of Automobile Engineering, Bharath Institute of Science and
Technology (BIST), Bharath University, Chennai, Tamil Nadu, India [email protected]
3Professor, Department of Automobile Engineering, Bharath Institute of Science and Technology
(BIST), Bharath University, Chennai, Tamil Nadu, India [email protected]
4 M.Tech Scholar, Department of Automobile Engineering, Bharath Institute of Science and
Technology (BIST), Bharath University, Chennai, Tamil Nadu, India [email protected]
ABSTRACT
In recent times, alternative combustion technology such as Homogenous Charge Compression Ignition (HCCI) has been studied and results have been positive. HCCI combustion has the potential to reduce fuel consumption and NOx emissions pertaining to the most stringent of legislation of both present and future. HCCI technology is attractive as there is no need for major modifications to the existing structure of IC engines and with significantly low NOx emissions, after treatment systems are not required. However, it is difficult to control the process and achieve constancy every cycle. Therefore, globally experts are studying HCCI combustion in depth to understand the associated idiosyncrasies. Through advent of modern computers, it has become possible to simulate HCCI combustion by creating a mathematical model that can solve complex equations within minutes. This paper details mathematical modelling approach to model HCCI combustion using Visual Basic for Applications (VBA), along with insight on different types of modelling techniques and submodels required to construct the simulation model.
KEYWORDS: HCCI, Auto Ignition, Mathematical Modelling, Mechanistic Model, Visual Basic for Applications (VBA)
INTRODUCTION
The current scenario of fossil fuel shortage, increase in prices and environmental problems due to vehicle emissions, has motivated engineers, scientists, technical education institutes and companies globally, to come up with an alternative technology to the conventional form of burning
the fuel and extracting energy out of it. Several approaches have been adapted to solve this problem such as:
Alternative Fuel: Use of fuel having better properties than gasoline or diesel in terms of self ignition temperature, volatility and calorific value
Alternative Combustion method: A combustion method that is more efficient than conventional methods in terms of work output and emissions.
Alternative Materials: Materials that can sustain higher forces, and therefore high pressure and temperature can be used inside the cylinder so as to extract more work
Alternative Engine Structure: Radically change the engine structure so as to reduce number of moving parts and hence, reduce frictional loss
Alternative working cycle: Explore the use of alternative working cycle to Otto or Diesel cycle that can provide more work output.
For the purpose of this paper, an alternative combustion method namely, Homogeneous Charge Compression Ignition (HCCI), is studied with the purpose of creating a simulation model.
HCCI
HCCI combustion is achieved by premixing the air-fuel mixture, either in the manifold or by early direct injection, as in Spark Ignition (SI) engine, and compressing the mixture until the temperature inside the combustion chamber reaches the auto ignition point and ignites, as in Compression Ignition (CI) engine. It is also known by following terms –
Controlled Auto Ignition
Active Thermo Atmosphere Combustion
Controlled Auto Ignition (CAI)
Active Thermo Atmosphere Combustion (ATAC)
Premixed Charge Compression Ignition (PCCI)
Homogenous Charge Diesel Combustion (HCDC)
Premixed Lean Diesel Combustion (PREDIC)
Compression Ignited Homogenous Charge (CIHC)
HCCI engine combines the advantages of Spark Ignition engine (Homogenous Charge) and Compression Ignition engine (increased efficiency) with reduced emissions. However, unlike SI and CI engine where start of combustion is controlled by spark timing and fuel injection timing respectively, there is no direct way to control the initiation of ignition, and as a result, it becomes innately difficult to control the process in order to extract maximum work from each cycle.
According to various researches done on HCCI engine, combustion happens simultaneously within the cylinder as opposed to flame front phenomenon in conventional SI or CI engine. Hence the combustion duration is comparatively lower resulting in lower
peak temperature. Therefore, harmful NOx emissions, which are temperature driven, are reduced considerably. This has motivated scientists and engineers throughout the globe to undertake research on HCCI combustion for better understanding and thereby exploring possibilities for commercial application. The following section highlights some of the research findings pertaining to HCCI.
LITERATURE REVIEW
Autoignition has been investigated from the start of automobile mass production era, albeit not as a process but to understand fuel properties such as autoignition temperature. Only more recently, autoignition process has been studied with the purpose of extracting positive work from the engine.
Onishi et al. in 1979 examined the possibility of running a two stroke engine on autoignition mode for their research. They concluded that there was little cycle by cycle variation with respect to peak combustion pressure and that the reaction happened spontaneously at several points within the cylinder. The importance of radicals was also studied and it was shown that their concentration was higher and had longer life than in an SI engine. They also suggested maintaining uniform quantity of mixture and the air to fuel ratio from cycle to cycle in order to attain HCCI. They obtained adequate combustion over a wide range of air to fuel ratios and concluded that HCCI reduces both exhaust emissions and fuel consumption for the entire range.
In 1983, Najt et al. studied in detail what parameters affect HCCI combustion using a single cylinder four stroke cycle engine with a pancake combustion chamber and a shrouded intake valve. They concluded that this autoignition phenomenon was not knocking but an even energy release process that can be controlled by manipulating temperature and mixture strength. They independently controlled inlet temperature and used EGR, simultaneously using different fuels for their experiments and following conclusions were drawn:
Chemical species in the EGR gases did not affect the Heat Release Rate (HRR), as a result EGR was used to modulate initial temperature of air fuel mixture so as to facilitate autoignition.
The combustion process was sensitive to delivery ratio through changes in the concentrations of air and fuel in the fresh charge, i.e. at high delivery ratios the energy release became unstable.
It was easy to ignite fuels having lower octane numbers
Thring, in 1989, used a single cylinder four stroke internal combustion engine to examine the feasibility of HCCI combustion. He performed several experiments with intake temperature as high as 425 o C, equivalence ratios ranging from 0.33 to 1.3 and EGR
rates up to 33 % , using both diesel and gasoline to map the satisfactory operating regions. He observed that EGR was required in order to raise the intake temperature while there
was low cyclic variability and fuel economy results were comparable with diesel engine. He concluded that there were three unsatisfactory regions, dictated by mixture strength, i.e. equivalence ratio –
Mixture was too rich resulting in misfiring or knocking, labelled as ‘misfire region’ and ‘knock region’
Mixture was too lean resulting in low power production, labelled as ‘power limited region’ Lei Shi et al. investigated the effect of internal and external EGR on emissions and performance of four stroke HCCI engine running on diesel fuel. They observed that by injecting fuel before Top Dead Centre (TDC) of exhaust stroke, and employing negative valve overlap (NVO), the homogeneous mixture, when burned achieves low NOx and smoke emissions. It was also noticed that internal EGR benefited in the formation of homogeneous mixture, further reducing smoke emissions, however, high load limit of HCCI was affected negatively. Whereas cooled external EGR delayed the start of combustion, thereby helping to avoid knocking, this in turn expanded the high load limit of HCCI engine. Due to no fuel rich regions in the cylinder, smoke emissions were on lower side as compared to a conventional diesel engine.
Osbourne et al. conducted a study on evaluating HCCI combustion mode for future gasoline powertrains, with prime objective being to develop a greater understanding of in-cylinder processes. Based on experimental results they developed a 1D simulation model using Ricardo Wave software a CFD based 3D model to perform computations. As per their observation, there was 99% reduction in NOx emissions and an 8% reduction in ISFC compared with the baseline direct injection gasoline engine condition for a standard key point. Also, HC emissions for HCCI operation were comparable to other conventional gasoline engine modes of operation. Finally, they suggest the concept of two-stroke/four stroke switching HCCI engine, made possible by using camless, electro mechanical variable valve actuation.
HCCI MATHEMATICAL MODELLING
Mathematical modelling is a cheaper, faster and efficient way to gain an insight into the working of a system. It requires great deal of skills and resources to setup an experimental facility for HCCI combustion testing. However, with knowledge gained from various experiments conducted on HCCI combustion, it is possible to formulate a mathematical model and derive a set of relationships between various input parameters and output results. This model can further be fine tuned by validating against known experiment results or by conducting an experiment using exactly similar set of control parameters, as used in simulation model. The following section provides details on formulation of a mathematical model for the purpose of simulation of HCCI combustion.
There are mainly two types of mathematical models that are used to describe the physical phenomena occurring within the engine cylinder:
Phenomenological Model (Black box model)
A phenomenological model or an empirical model is derived using experimental data only, using no prior information about the system, i.e. engine cylinder during HCCI combustion. Statistical principles are used to derive relationships among sensitive parameters affecting the final result. For such models, an experiment is set up where input parameters are controlled and output is measured. Input parameters or experimental setup is altered to study the corresponding effect on output result. From results of experiment, graphs are charted showing correlation between input parameters and measured results. Phenomenological models are more general and applicable to many different kinds of problems. However, they provide less insight into the problem or its possible solution and less predictive capability.
Mechanistic Model (Grey box model)
A mechanistic model is developed using prior information about the system, i.e. results of previous work conducted on similar topic. They provide deeper understanding and more accurate prediction as compared to phenomenological models.
Globally, research on HCCI combustion engine has been undertaken by almost every academic institution, automobile manufacturers, and consultants since last 3 decades. Also, from the detailed research done by the author, the information gathered so far on HCCI combustion, a glimpse of which is mentioned in Literature Review section, is sufficient to develop a mechanistic model on HCCI combustion.
The various mechanistic engine combustion models that have been developed so far can by categorised as per below:
Zero dimensional models
Quasi-dimensional models
Multi-dimensional models
The level of detail and closeness to the real life physics increase as one moves from zero dimensional models to multi-dimensional models, along with the intricacy of models and using these models. There is always a trade-off between the usability of the model and the ability to accurately predict the outcome.
Zero dimensional models are the simplest and most suitable to model the effects of change of input parameters on heat release rate and pressure rise rate during HCCI combustion. Depending upon the assumptions made with respect to division of zones inside the cylinder, zero dimensional models can be further classified into –
Single zone: The entire cylinder is considered as a single zone and calculations are done by applying first law of thermodynamics. The combustible fluid mixture in the
engine is assumed to be a thermodynamic system that undergoes the process of energy exchange with the surroundings.
Two zones: The cylinder is divided into two zones, a burned and an unburned zone. These zones are two different thermodynamic systems with energy and mass exchange between them and common surrounding, i.e. the cylinder wall.
Multi zones: Multi zone models are used to evaluate energy and mass interactions through several zones, thus providing more accurate results, albeit with increasing level of complexity.
For HCCI combustion modelling, two zones model is used to simulate the unburned and burned zones.
MODELLING PLATFORM
Several packages have been developed to simulate the internal combustion engine system. Most notable simulation packages, currently used in automotive research are:
Ricardo Wave
Lotus Engine Simulation
AVL fire
GT-Power
KIVA
MATLAB Simulink
LabVIEW
However, software costs, hardware requirements and training limitations, combined with project time constraint of the research work undertaken by the author, renders any of the above software packages difficult to use. While above simulation platforms are primarily based on computation fluid dynamics, simple mechanistic models, can be built and developed using any software package offering computation ability to solve engineering problems, program event based automation and a graphic tool to visualize outcome, such as Visual Basic for Applications (VBA) with Excel spreadsheet.
SIMULATION STRUCTURE
The simulation structure is prepared based on several assumptions related to thermodynamic principles, which are listed below:
The gases present in burned and unburned zones are ideal gases with different properties
No heat transfer between unburned and burned zone
Instantaneous pressure in both the zones is the same because flame is a deflagration combustion wave
The characteristic gas constants within both zones do not vary much with pressure and temperature; in case of any variance, they can be modelled using thermodynamic relationships of gas constants with temperature and pressure
Since enthalpy connected with injected fuel is insignificant, it is ignored
