**1. Introduction**

In the context of a Spark Ignition engine, the inherent complexity of premixed combustion is exacerbated by a range of engine variables that render the process highly transient in nature and not fully predictable. The present work aims to contribute to the continuous research effort to better understand the details of combustion and be able to model the process in gasoline SI engines. Coexisting fossil fuels depletion and environmental concerns, along with an alarming connection between traditional internal combustion engines emissions and human health degradation [1], have in recent years driven a strong research interest upon premixed SI combustion of energy sources alternative to gasoline, including liquid alcohols like ethanol, and gaseous fuels like hydrogen. However, the advancements enjoyed by gasoline-related technology and infrastructure in the last 40 years have eroded the potential advantages in efficiency and emissions offered by alternative fuels [2], and the SI engine running on gasoline continues to be the most common type of power unit used in passenger cars (Port-Fuel Injection gasoline engines accounted for the vast majority (91%) of all light-duty vehicle engines produced for the USA market in 2010 [3]).

The characteristics which make the gasoline engine well suited to light-weight applications includerelativelyhighpowertoweightratio,acceptableperformanceoverawiderangeofengine speeds, the vast infrastructure for gasoline and lower manufacturing costs when compared to diesel or more modern hybrid technologies [4]. The continuing exploitation of spark ignition engines reflects a history of successful development and innovation. These have included the electronic fuel injection system, exhaust emissions after-treatment, Exhaust Gas Recirculation and, increasingly, the use of some form of variable actuation valve train system. The modern SI engine, addressed to as high-degree-of-freedom engine by Prucka et al. [5], may also feature flexible fuel technology, typically to allow running on ethanol-gasoline blended fuels.

© 2013 Bonatesta; licensee InTech. This is an open access article distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/by/3.0), which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited. © 2013 Bonatesta; licensee InTech. This is a paper distributed under the terms of the Creative Commons Attribution License (http://creativecommons.org/licenses/by/3.0), which permits unrestricted use, distribution, and reproduction in any medium, provided the original work is properly cited.

As the technology advances, the number of engine actuators increases and so does the number of variables that may potentially modify the combustion process. Methods of combustion control based on look-up tables may well be implemented in *high-degree-of-freedom* engines, for example to set optimal spark timing and phase combustion appropriately across Top Dead Centre, but are not well-suited during transient operation, when the boundary conditions are changing on a cycle-to-cycle basis. Whilst controlling the combustion process in highly complex engine architectures becomes more challenging, the development of straightforward modelling approaches, which allow reliable inclusion within real-time feed-forward engine controllers become essential to ensure improved performance and fuel efficiency also during transient or variable operation.

**2.1. Overview of flame propagation mechanism**

flame front, which diffuses outwards with laminar flame speed [6].

*b*

r

r

*d*

t

The second one considers the rate of mass entrainment into the flame-front:

*e*

*d*

t

formation of burned products:

Gillespie and co-workers provide a useful review of those aspects of laminar and turbulent flame propagation, which are relevant to SI engines combustion [8]. Similarly to laminar-like combustion taking place in a quiescent environment, two main definitions of time-based combustion rate can be proposed for turbulent combustion. The first one relates to the rate of

> *u fb dm A S*

*u fe dm A S*

In the above fundamental expressions of mass continuity, *ρ<sup>u</sup>* is the unburned gas density, *Af* is a reference reaction-front surface area and *Sb* (or *Se*) is the turbulent burning (or entrainment)

<sup>=</sup> (1)

<sup>=</sup> (2)

Detailed observations of development and structure of the flame in SI engines can be made by using direct photographs or other methods such as Schlieren and shadowgraph photography techniques [6,7].Theinitial stageofthecombustionprocess is thedevelopmentofaflamekernel, centred close to the spark-plug electrodes, that grows from the spark discharge with quasispherical, low-irregular surface; its outer boundary corresponds to a thin sheet-like develop‐ ing reaction front that separates burned and unburned gases. Engine combustion takes place in a turbulent environment produced by shear flows set up during the induction stroke and then modified during compression. Initially, the flame kernel is too small to incorporate most of the turbulence length scales available and, therefore, it is virtually not aware of the velocity fluctuations[8].Onlythesmallestscalesofturbulencemayinfluencethegrowingkernel,whereas bigger scales are presumed to only convect the flame-ball bodily; the initial burning character‐ istics are similarto those foundinaquiescent environment(a laminar-like combustiondevelop‐ ment). As the kernel expands, it progressively experiences larger turbulent structures and the reaction front becomes increasingly wrinkled. During the main combustion stage, the thin reaction sheet becomes highly wrinkled and convoluted and the reaction zone, which sepa‐ rates burned and unburned gases, has been described as a *thick turbulent flame brush*. While the thickness of the initial sheet-like reaction front is of the order of 0.1 mm, the overall thickness of this turbulent flame brush can reach several millimetres; this would depend on type of fuel, equivalence ratio and level of turbulence. The turbulent flow field, in particular velocity fluctuations, determines a conspicuous rate of entrainment in the reaction zone, which has been described [9, 10] as being composed of many small pockets and isolated island of unburned gas withinhighlymarkedwrinkles that characterize a thinmulti-connectedreactionsheet.Theories have been advanced that describe the local boundary layer of this region as a quasi-spherical

Premixed Combustion in Spark Ignition Engines and the Influence of Operating Variables

http://dx.doi.org/10.5772/55495

5

The premixed, homogeneous charge gasoline combustion process in SI engines is influenced by the thermo-chemical state of the cylinder charge. Significant factors are local temperature and pressure, stoichiometry and the contents of burned gas within the combustible mixture; these quantities affect rate of burning and consequent in-cylinder pressure development. The combustion process is also greatly influenced by cylinder bulk motion and micro-scale turbulence. Understanding the connection between charge burn characteristics and relevant engines operating variables in the context of modern technologies is extremely useful to enable and support engine design innovation and the diagnosis of performance. The present chapter explores the evolution of the combustion process in modern-design gasoline engines, as indicated by the cylinder charge Mass Fraction Burned variation and combustion duration, and the most relevant factors influencing these. It also explores the use, accuracy and limita‐ tions of recently-proposed empirical, non-dimensional (or simplified thermodynamic) combustion models which respond to the requirements of fast execution within model-based control algorithms, and discusses relevant results, which entail the use of Variable Valve Timing systems. An exemplar simplified quasi-dimensional models is also presented at the end of the chapter, along with some relevant results concerning an application to flexible fuel, gasoline/ethanol operation. All the experimental data and models discussed here refer and are applicable to stable combustion, typically identified by a Coefficient of Variability of the Indicated Mean Effective Pressure (CoV of IMEP) smaller or equal to 6% [6]. Although the importance of cycle-by-cycle variability is acknowledged, as this may arise from highly diluted combustion, the topic of unstable combustion has not been the focus of the present work.

## **2. Premixed combustion in SI engines**

The present section reviews important features of the premixed combustion process in SI engines, introducing basic terms and definitions of relevant variables and combustion indicators. Ample space is dedicated to the working principles of VVT systems and how these may fundamentally affect the combustion process. This section ultimately provides definitions and methods of determination of in-cylinder charge diluent fraction, as the one most influential variable on combustion strength, duration and stability, in the case of engines fitted with a VVT system.

#### **2.1. Overview of flame propagation mechanism**

As the technology advances, the number of engine actuators increases and so does the number of variables that may potentially modify the combustion process. Methods of combustion control based on look-up tables may well be implemented in *high-degree-of-freedom* engines, for example to set optimal spark timing and phase combustion appropriately across Top Dead Centre, but are not well-suited during transient operation, when the boundary conditions are changing on a cycle-to-cycle basis. Whilst controlling the combustion process in highly complex engine architectures becomes more challenging, the development of straightforward modelling approaches, which allow reliable inclusion within real-time feed-forward engine controllers become essential to ensure improved performance and fuel efficiency also during

The premixed, homogeneous charge gasoline combustion process in SI engines is influenced by the thermo-chemical state of the cylinder charge. Significant factors are local temperature and pressure, stoichiometry and the contents of burned gas within the combustible mixture; these quantities affect rate of burning and consequent in-cylinder pressure development. The combustion process is also greatly influenced by cylinder bulk motion and micro-scale turbulence. Understanding the connection between charge burn characteristics and relevant engines operating variables in the context of modern technologies is extremely useful to enable and support engine design innovation and the diagnosis of performance. The present chapter explores the evolution of the combustion process in modern-design gasoline engines, as indicated by the cylinder charge Mass Fraction Burned variation and combustion duration, and the most relevant factors influencing these. It also explores the use, accuracy and limita‐ tions of recently-proposed empirical, non-dimensional (or simplified thermodynamic) combustion models which respond to the requirements of fast execution within model-based control algorithms, and discusses relevant results, which entail the use of Variable Valve Timing systems. An exemplar simplified quasi-dimensional models is also presented at the end of the chapter, along with some relevant results concerning an application to flexible fuel, gasoline/ethanol operation. All the experimental data and models discussed here refer and are applicable to stable combustion, typically identified by a Coefficient of Variability of the Indicated Mean Effective Pressure (CoV of IMEP) smaller or equal to 6% [6]. Although the importance of cycle-by-cycle variability is acknowledged, as this may arise from highly diluted combustion, the topic of unstable combustion has not been the focus of the present work.

The present section reviews important features of the premixed combustion process in SI engines, introducing basic terms and definitions of relevant variables and combustion indicators. Ample space is dedicated to the working principles of VVT systems and how these may fundamentally affect the combustion process. This section ultimately provides definitions and methods of determination of in-cylinder charge diluent fraction, as the one most influential variable on combustion strength, duration and stability, in the case of engines fitted with a

transient or variable operation.

4 Advances in Internal Combustion Engines and Fuel Technologies

**2. Premixed combustion in SI engines**

VVT system.

Detailed observations of development and structure of the flame in SI engines can be made by using direct photographs or other methods such as Schlieren and shadowgraph photography techniques [6,7].Theinitial stageofthecombustionprocess is thedevelopmentofaflamekernel, centred close to the spark-plug electrodes, that grows from the spark discharge with quasispherical, low-irregular surface; its outer boundary corresponds to a thin sheet-like develop‐ ing reaction front that separates burned and unburned gases. Engine combustion takes place in a turbulent environment produced by shear flows set up during the induction stroke and then modified during compression. Initially, the flame kernel is too small to incorporate most of the turbulence length scales available and, therefore, it is virtually not aware of the velocity fluctuations[8].Onlythesmallestscalesofturbulencemayinfluencethegrowingkernel,whereas bigger scales are presumed to only convect the flame-ball bodily; the initial burning character‐ istics are similarto those foundinaquiescent environment(a laminar-like combustiondevelop‐ ment). As the kernel expands, it progressively experiences larger turbulent structures and the reaction front becomes increasingly wrinkled. During the main combustion stage, the thin reaction sheet becomes highly wrinkled and convoluted and the reaction zone, which sepa‐ rates burned and unburned gases, has been described as a *thick turbulent flame brush*. While the thickness of the initial sheet-like reaction front is of the order of 0.1 mm, the overall thickness of this turbulent flame brush can reach several millimetres; this would depend on type of fuel, equivalence ratio and level of turbulence. The turbulent flow field, in particular velocity fluctuations, determines a conspicuous rate of entrainment in the reaction zone, which has been described [9, 10] as being composed of many small pockets and isolated island of unburned gas withinhighlymarkedwrinkles that characterize a thinmulti-connectedreactionsheet.Theories have been advanced that describe the local boundary layer of this region as a quasi-spherical flame front, which diffuses outwards with laminar flame speed [6].

Gillespie and co-workers provide a useful review of those aspects of laminar and turbulent flame propagation, which are relevant to SI engines combustion [8]. Similarly to laminar-like combustion taking place in a quiescent environment, two main definitions of time-based combustion rate can be proposed for turbulent combustion. The first one relates to the rate of formation of burned products:

$$\frac{dm\_b}{d\tau} = \rho\_u \ A\_f \ S\_b \tag{1}$$

The second one considers the rate of mass entrainment into the flame-front:

$$\frac{dm\_e}{d\tau} = \rho\_u \; A\_f \; \mathcal{S}\_e \tag{2}$$

In the above fundamental expressions of mass continuity, *ρ<sup>u</sup>* is the unburned gas density, *Af* is a reference reaction-front surface area and *Sb* (or *Se*) is the turbulent burning (or entrainment) velocity. The dependence of the combustion rate on turbulence is embodied in the velocity term, which is fundamentally modelled as a function of turbulence intensity, *u* ', and laminar burning velocity, *SL* . The latter, loosely addressed to as laminar flame velocity in the context of simplified flame propagation models, has been demonstrated to retain a leading role even during turbulent combustion and depends strongly upon the thermodynamic conditions (namely pressure and temperature) and upon the chemical state (namely combustible mixture strength, i.e. stoichiometry, and burned gas diluent fraction) of the unburned mixture ap‐ proaching the burning zone.

sion (and at the start of combustion) would be dictated by the concurrent rates of turbulence production and natural viscous dissipation [17]. Although the literature is somewhat unclear on this specific topic, increased tumble ratio has been also reported to improve the cyclic

Premixed Combustion in Spark Ignition Engines and the Influence of Operating Variables

http://dx.doi.org/10.5772/55495

7

Two parameters are commonly used to describe the effects of turbulence on flame propagation: integral length scale *L* and turbulence intensity *u* '. The first one is a measure of the size of the large turbulent eddies and correlates with the available height of the combustion chamber; when the piston is at TDC of combustion, *L* is typically 2 mm [19]. The second parameter is defined as the root-mean-square of the velocity fluctuations. According to numerous experi‐ mental studies available in the literature, for example [12, 20, 21], the turbulence intensity, for given engine and running set-up, would depend primarily on engine speed (or mean piston speed). Computational Fluid Dynamics studies of the in-cylinder turbulence regime, per‐ formed by the Author [22] on a PFI, 4-valve/cylinder, pent-roof engine show that turbulence intensity (modelled using a conventional *k* −*ε* approach [6]) is characterised by a weakly decreasing trend during compression and up to TDC of combustion. In the range of engine speeds investigated, which were between 1250 and 2700 rev/min, the volume-averaged value of turbulence intensity, when piston is approaching TDC, can be approximated by the correlation: *u* '≈0.38*SP*, where *SP* is the mean piston speed (units of m/s), given by *SP* =2*SN* ,

Theories have been developed which ascribe importance to additional turbulence generated inside the unburned region ahead of the reaction-front, by the expanding flame. None of them has been confirmed by direct observations and their validity has been always inferred by means of comparisons between models predictions and experimental data. Tabaczynski and co-workers [23, 24] advance the so-called eddy rapid distortion theory according to which the individual turbulence eddies experience fast isentropic compression, in such a way that their angular momentum is conserved. They conclude that due to this interaction the turbulence intensity increases and the length scale reduces, respectively, during the combustion process. Hoult and Wong [25], in a theoretical study based on a cylindrical constant-volume combustion vessel, apply the same rapid distortion theory to conclude that the turbulence level of the unburned gas depends only on its initial value and the degree of compression due to the expanding flame. An interesting fit of experimental data to inferred combustion-generated

ThemostcommonlystatedreasonforintroducingVariableValveActuationsystemsinSIengines is to raise the engine brake torque and achieve improvements in its variation with engine speed, especiallyatlowspeed(includingidle conditions) andatthehighendofthe engine speedrange. A second coexistent reason is to reduce the exhaust emissions, especially nitrogen oxides, but also unburned hydrocarbons [26]. Today many modern engines are equipped with VVA technology because measurable improvements can be gained in fuel consumption and efficien‐ cy over wide ranges of operating conditions, including part-load conditions. Efficiency improvements are a direct consequence of a reduction in pumping (intake throttling) losses. At

stability and extend the running limits for lean or diluted mixtures [15, 18].

with *S* engine stroke (m) and *N* engine speed (rev/s).

turbulence intensity is due to Groff and Matekunas [12].

**2.3. Variable valve actuation mechanisms**

The difference between the two expressions of the combustion rate depends on the real, finite flame front thickness that at each moment in time would host a certain mass (*me* −*mb*), already entrained into the reaction zone but not yet burned. Several definitions can be used for the reference surface-area: the quantity *Af* identified above is the stretched cold flame-front, usually assumed to be smooth and approximately spherical, detectable with good approxi‐ mation using Schlieren images techniques and then traced with best-fit circles [11, 12]. A different approach considers the so-called burning surface *Ab*, defined as the surface of the volume *Vb* that contains just burned gas: the difference (*rf* −*rb*) between the correspondent radii would scale with the size of the wrinkles that characterise the real, thick reaction zone. When the burning velocities are calculated from experimental burning rates/pressure data (see below), the cold surface *Af* is often equated to the burning surface *Ab* [13], which assumes that the thickness of the reaction zone/front-sheet can be neglected.

Flame sheets, in real combustion processes, are subject to stretch, which shows a smoothing effect on the flame-front surface, and tends to reduce the burning velocities. When the flame is fully developed, incorporating most of the available turbulent spectrum, geometrical stretch is superseded by aerodynamic strain. The action of flame stretching in all stages of combustion reduces at increasing pressure, being low at engine-like operating conditions [8].

#### **2.2. In-cylinder motion field and effects on combustion**

Although the mean charge velocity in an engine cylinder may have an effect on the initial rate of combustion, by distorting the developing flame kernel and, possibly, by increasing the available burning surface [14], the main mechanism of combustion enhancement is turbulence.

Modern-design gasoline engines typically have 4 valves per cylinder, 2 intake and 2 exhaust valves. The use of two intake valves, which gives symmetry of the intake flow about the vertical axis, generates a mean cylinder motion called tumble, or vertical or barrel swirl, an organised rotation of the charge about an axis perpendicular to the cylinder axis. The strength of a tumbling flow is measured by means of a non-dimensional number called tumble ratio, defined as the ratio between the speed of the rotating bulk-flow and the rotational speed of the engine. The tumbling mean flow has been observed to promote combustion [15, 16] through turbulence production towards the end of the compression stroke. As the flow is compressed in a diminishing volume, the rotating vortices that make up the tumbling flow tend to break down into smaller structures and their kinetic energy is gradually and partially converted in turbulent kinetic energy. Whether the turbulence intensity is actually rising during compres‐

sion (and at the start of combustion) would be dictated by the concurrent rates of turbulence production and natural viscous dissipation [17]. Although the literature is somewhat unclear on this specific topic, increased tumble ratio has been also reported to improve the cyclic stability and extend the running limits for lean or diluted mixtures [15, 18].

Two parameters are commonly used to describe the effects of turbulence on flame propagation: integral length scale *L* and turbulence intensity *u* '. The first one is a measure of the size of the large turbulent eddies and correlates with the available height of the combustion chamber; when the piston is at TDC of combustion, *L* is typically 2 mm [19]. The second parameter is defined as the root-mean-square of the velocity fluctuations. According to numerous experi‐ mental studies available in the literature, for example [12, 20, 21], the turbulence intensity, for given engine and running set-up, would depend primarily on engine speed (or mean piston speed). Computational Fluid Dynamics studies of the in-cylinder turbulence regime, per‐ formed by the Author [22] on a PFI, 4-valve/cylinder, pent-roof engine show that turbulence intensity (modelled using a conventional *k* −*ε* approach [6]) is characterised by a weakly decreasing trend during compression and up to TDC of combustion. In the range of engine speeds investigated, which were between 1250 and 2700 rev/min, the volume-averaged value of turbulence intensity, when piston is approaching TDC, can be approximated by the correlation: *u* '≈0.38*SP*, where *SP* is the mean piston speed (units of m/s), given by *SP* =2*SN* , with *S* engine stroke (m) and *N* engine speed (rev/s).

Theories have been developed which ascribe importance to additional turbulence generated inside the unburned region ahead of the reaction-front, by the expanding flame. None of them has been confirmed by direct observations and their validity has been always inferred by means of comparisons between models predictions and experimental data. Tabaczynski and co-workers [23, 24] advance the so-called eddy rapid distortion theory according to which the individual turbulence eddies experience fast isentropic compression, in such a way that their angular momentum is conserved. They conclude that due to this interaction the turbulence intensity increases and the length scale reduces, respectively, during the combustion process. Hoult and Wong [25], in a theoretical study based on a cylindrical constant-volume combustion vessel, apply the same rapid distortion theory to conclude that the turbulence level of the unburned gas depends only on its initial value and the degree of compression due to the expanding flame. An interesting fit of experimental data to inferred combustion-generated turbulence intensity is due to Groff and Matekunas [12].

#### **2.3. Variable valve actuation mechanisms**

velocity. The dependence of the combustion rate on turbulence is embodied in the velocity term, which is fundamentally modelled as a function of turbulence intensity, *u* ', and laminar burning velocity, *SL* . The latter, loosely addressed to as laminar flame velocity in the context of simplified flame propagation models, has been demonstrated to retain a leading role even during turbulent combustion and depends strongly upon the thermodynamic conditions (namely pressure and temperature) and upon the chemical state (namely combustible mixture strength, i.e. stoichiometry, and burned gas diluent fraction) of the unburned mixture ap‐

The difference between the two expressions of the combustion rate depends on the real, finite flame front thickness that at each moment in time would host a certain mass (*me* −*mb*), already entrained into the reaction zone but not yet burned. Several definitions can be used for the reference surface-area: the quantity *Af* identified above is the stretched cold flame-front, usually assumed to be smooth and approximately spherical, detectable with good approxi‐ mation using Schlieren images techniques and then traced with best-fit circles [11, 12]. A different approach considers the so-called burning surface *Ab*, defined as the surface of the volume *Vb* that contains just burned gas: the difference (*rf* −*rb*) between the correspondent radii would scale with the size of the wrinkles that characterise the real, thick reaction zone. When the burning velocities are calculated from experimental burning rates/pressure data (see below), the cold surface *Af* is often equated to the burning surface *Ab* [13], which assumes that

Flame sheets, in real combustion processes, are subject to stretch, which shows a smoothing effect on the flame-front surface, and tends to reduce the burning velocities. When the flame is fully developed, incorporating most of the available turbulent spectrum, geometrical stretch is superseded by aerodynamic strain. The action of flame stretching in all stages of combustion

Although the mean charge velocity in an engine cylinder may have an effect on the initial rate of combustion, by distorting the developing flame kernel and, possibly, by increasing the available burning surface [14], the main mechanism of combustion enhancement is turbulence. Modern-design gasoline engines typically have 4 valves per cylinder, 2 intake and 2 exhaust valves. The use of two intake valves, which gives symmetry of the intake flow about the vertical axis, generates a mean cylinder motion called tumble, or vertical or barrel swirl, an organised rotation of the charge about an axis perpendicular to the cylinder axis. The strength of a tumbling flow is measured by means of a non-dimensional number called tumble ratio, defined as the ratio between the speed of the rotating bulk-flow and the rotational speed of the engine. The tumbling mean flow has been observed to promote combustion [15, 16] through turbulence production towards the end of the compression stroke. As the flow is compressed in a diminishing volume, the rotating vortices that make up the tumbling flow tend to break down into smaller structures and their kinetic energy is gradually and partially converted in turbulent kinetic energy. Whether the turbulence intensity is actually rising during compres‐

reduces at increasing pressure, being low at engine-like operating conditions [8].

the thickness of the reaction zone/front-sheet can be neglected.

**2.2. In-cylinder motion field and effects on combustion**

proaching the burning zone.

6 Advances in Internal Combustion Engines and Fuel Technologies

ThemostcommonlystatedreasonforintroducingVariableValveActuationsystemsinSIengines is to raise the engine brake torque and achieve improvements in its variation with engine speed, especiallyatlowspeed(includingidle conditions) andatthehighendofthe engine speedrange. A second coexistent reason is to reduce the exhaust emissions, especially nitrogen oxides, but also unburned hydrocarbons [26]. Today many modern engines are equipped with VVA technology because measurable improvements can be gained in fuel consumption and efficien‐ cy over wide ranges of operating conditions, including part-load conditions. Efficiency improvements are a direct consequence of a reduction in pumping (intake throttling) losses. At low to medium load, variable valve strategy, in particular the extension of the valve overlap interval(betweentheIntakeValveOpeningandExhaustValveClosing),exertsastronginfluence upon the amount of burned gas recirculated from one engine cycle to the following one. This amount,ormore specificallythe so-calleddilutionmass fraction,has aprofoundinfluenceupon combustion rates and duration. Combustion control strategies which aim at improved efficien‐ cyacross thewhole rangeof engine speeds andloadsmust carefullyconsiderthe extenttowhich the burning characteristics may be modified by VVA.

speed and load operating conditions. Early Intake Valve Opening timings produce large valve overlap interval and increase charge dilution with burned gas. Late IVO timings lead to increased pumping work, but may show an opposite effect at high engine speed where volumetric efficiency gains can be achieved by exploiting the intake system ram effects [6]. If the valve motion profiles are fixed, changes to IVO are reproduced by those to IVC, with significant effects on mass of fresh charge trapped, hence on engine load, and measurable changes in pumping losses. Early IVC controls engine load by closing the inlet valve when sufficient charge has been admitted into the cylinder. Reductions in Brake Specific Fuel Consumption of up to 10% have been observed with early IVC strategies [29, 30]. Recent studies by Fontana et al. [31] and by Cairns et al. [32] show similar reductions in fuel consumption, but explain these referring to the *displacement* of fresh air with combustion products during the valve overlap interval, which reduces the need forthrottling. The Exhaust Valve Opening strategy would be dictated by a compromise between the benefits of the exhaust blow-down (early EVO) and those associated to a greater expansion ratio (late EVO). At high speed and load conditions, late EVC exploits the benefits of the ram effect, which may assist in the combustion products scavenging process. The exhaust valve strategy also contributes to the process of *mixture preparation* at all engine conditions, by trapping burned gases in the cylinder (early EVC) or by backflow into the cylinder when intake and ex‐

Premixed Combustion in Spark Ignition Engines and the Influence of Operating Variables

http://dx.doi.org/10.5772/55495

9

Focusing on preparation of the combustible mixture and subsequent combustion process, the level of charge dilution by burned gas is the single most influential quantity, which is heavily varied using variable valve timing. Charge dilution tends to slow down the rate of combus‐ tion by increasing the charge heat capacity, ultimately reducing the adiabatic flame temper‐ ature. Charge dilution tends to increase with increasing valve overlap, particularly under light-load operating conditions when intake throttling produces a relatively high pressure differential between the exhaust and intake manifolds. This promotes a reverse flow of exhaust gas into the cylinder and intake ports. The recycled gas forms part of the trapped charge of the following engine cycle. There is a strong degree of interaction between the level of combustion products within the newly formed mixture and engine speed and load. Increasing speed shortens the duration of the valve overlap in real time, while increasing load raises the pressure-boundary of the intake system limiting the recirculating hot flows. At high speed and load conditions the increase of charge dilution with increasing valve

In the case of a gasoline engine fitted with VVT system, the dilution mass fraction is the sum of two different terms. The first one, properly named residual gas fraction, is associated with the amount of burned gas remaining inside the combustion chamber when the piston reaches the TDC of the exhaust stroke. If the exhaust valve closes before TDC, then the residual mass fraction would be given by the amount of burned gas trapped inside the cylinder at EVC. In

**2.4. Charge dilution mass fraction – Definitions and measurements**

symbols, the residual mass fraction is written as:

haust valves are overlapping (late EVC).

overlap is limited.

#### *2.3.1. Overview of VVA mechanisms*

Thedevelopment ofVVAmechanisms startedinthe late 1960s andthe first systemwas released into production in 1982 for the USA market, prompted by tightening emissions legislation [26]. The mechanism was a simple two-position device, which reduced the valve overlap at idle conditions, improving combustion stability and hence reducing the noxious emissions. Very differentobjectives,inparticulartheincreaseofthebraketorqueoutputatbothendsoftheengine speed range, induced a second manufacturer to develop a VVA system for small-capacity motorcycle engines. Released also in the early 1980s, the system worked by simply deactivat‐ ing one inlet and one exhaust valve per cylinder at engine speeds below a fixed limit, achiev‐ ingbettermixingandgreaterin-cylinderturbulenceas theavailableinletflowareawas reduced. A better understanding of the potential advantages in fuel efficiency has prompted, in recent years, an increased interest in VVA technology and most major manufacturers now produce engines with some form of VVA. Most systems presently in use allow continuous variable camshaft phasing; some complicated mechanisms are capable of switching cams to gain the benefits of different valve lifting profiles. From 2001 at least one manufacturer incorporated a variable valve lift and phase control mechanism into the first production engine that featured throttle-less control of engine load [27]. The amount of fresh air trapped into the cylinder is controlled solely by appropriate Intake Valve Closing strategy, removing the need for throt‐ tling and the associated pumping losses. Variable lift serves as a means of controlling the air induction velocity and ultimately the level of in-cylinder turbulence.

Ahmad and co-workers [28] classify the VVA systems into five categories depending on their level of sophistication. The most complicated devises are classified in category 5, capable of varying valve lift, opening durations and phasing, independently of each other for both intake and exhaust valve trains. Despite the potential advantages, mechanical systems in category 5 tend to be expensive, physically bulky and complicated. The mechanism used by the Author for the experimental work reported in the following sections is classified in category 3, as it allows continuous and independent variable phasing of intake and exhaust valve opening intervals, with fixed valve lifting profiles. This system is usually called Twin Independent-Variable Valve Timing. The Twin Equal-VVT system represents a simplification of the TI-VVT, where both camshafts are phased simultaneously by equal amounts.

#### *2.3.2. VVT strategies and influence on charge diluent fraction*

By means of multiple combinations of intake and exhaust valve timings, the TI-VVT system allows the identification of optimal operating strategies across the whole range of engine speed and load operating conditions. Early Intake Valve Opening timings produce large valve overlap interval and increase charge dilution with burned gas. Late IVO timings lead to increased pumping work, but may show an opposite effect at high engine speed where volumetric efficiency gains can be achieved by exploiting the intake system ram effects [6]. If the valve motion profiles are fixed, changes to IVO are reproduced by those to IVC, with significant effects on mass of fresh charge trapped, hence on engine load, and measurable changes in pumping losses. Early IVC controls engine load by closing the inlet valve when sufficient charge has been admitted into the cylinder. Reductions in Brake Specific Fuel Consumption of up to 10% have been observed with early IVC strategies [29, 30]. Recent studies by Fontana et al. [31] and by Cairns et al. [32] show similar reductions in fuel consumption, but explain these referring to the *displacement* of fresh air with combustion products during the valve overlap interval, which reduces the need forthrottling. The Exhaust Valve Opening strategy would be dictated by a compromise between the benefits of the exhaust blow-down (early EVO) and those associated to a greater expansion ratio (late EVO). At high speed and load conditions, late EVC exploits the benefits of the ram effect, which may assist in the combustion products scavenging process. The exhaust valve strategy also contributes to the process of *mixture preparation* at all engine conditions, by trapping burned gases in the cylinder (early EVC) or by backflow into the cylinder when intake and ex‐ haust valves are overlapping (late EVC).

low to medium load, variable valve strategy, in particular the extension of the valve overlap interval(betweentheIntakeValveOpeningandExhaustValveClosing),exertsastronginfluence upon the amount of burned gas recirculated from one engine cycle to the following one. This amount,ormore specificallythe so-calleddilutionmass fraction,has aprofoundinfluenceupon combustion rates and duration. Combustion control strategies which aim at improved efficien‐ cyacross thewhole rangeof engine speeds andloadsmust carefullyconsiderthe extenttowhich

Thedevelopment ofVVAmechanisms startedinthe late 1960s andthe first systemwas released into production in 1982 for the USA market, prompted by tightening emissions legislation [26]. The mechanism was a simple two-position device, which reduced the valve overlap at idle conditions, improving combustion stability and hence reducing the noxious emissions. Very differentobjectives,inparticulartheincreaseofthebraketorqueoutputatbothendsoftheengine speed range, induced a second manufacturer to develop a VVA system for small-capacity motorcycle engines. Released also in the early 1980s, the system worked by simply deactivat‐ ing one inlet and one exhaust valve per cylinder at engine speeds below a fixed limit, achiev‐ ingbettermixingandgreaterin-cylinderturbulenceas theavailableinletflowareawas reduced. A better understanding of the potential advantages in fuel efficiency has prompted, in recent years, an increased interest in VVA technology and most major manufacturers now produce engines with some form of VVA. Most systems presently in use allow continuous variable camshaft phasing; some complicated mechanisms are capable of switching cams to gain the benefits of different valve lifting profiles. From 2001 at least one manufacturer incorporated a variable valve lift and phase control mechanism into the first production engine that featured throttle-less control of engine load [27]. The amount of fresh air trapped into the cylinder is controlled solely by appropriate Intake Valve Closing strategy, removing the need for throt‐ tling and the associated pumping losses. Variable lift serves as a means of controlling the air

Ahmad and co-workers [28] classify the VVA systems into five categories depending on their level of sophistication. The most complicated devises are classified in category 5, capable of varying valve lift, opening durations and phasing, independently of each other for both intake and exhaust valve trains. Despite the potential advantages, mechanical systems in category 5 tend to be expensive, physically bulky and complicated. The mechanism used by the Author for the experimental work reported in the following sections is classified in category 3, as it allows continuous and independent variable phasing of intake and exhaust valve opening intervals, with fixed valve lifting profiles. This system is usually called Twin Independent-Variable Valve Timing. The Twin Equal-VVT system represents a simplification of the TI-VVT,

By means of multiple combinations of intake and exhaust valve timings, the TI-VVT system allows the identification of optimal operating strategies across the whole range of engine

the burning characteristics may be modified by VVA.

8 Advances in Internal Combustion Engines and Fuel Technologies

induction velocity and ultimately the level of in-cylinder turbulence.

where both camshafts are phased simultaneously by equal amounts.

*2.3.2. VVT strategies and influence on charge diluent fraction*

*2.3.1. Overview of VVA mechanisms*

Focusing on preparation of the combustible mixture and subsequent combustion process, the level of charge dilution by burned gas is the single most influential quantity, which is heavily varied using variable valve timing. Charge dilution tends to slow down the rate of combus‐ tion by increasing the charge heat capacity, ultimately reducing the adiabatic flame temper‐ ature. Charge dilution tends to increase with increasing valve overlap, particularly under light-load operating conditions when intake throttling produces a relatively high pressure differential between the exhaust and intake manifolds. This promotes a reverse flow of exhaust gas into the cylinder and intake ports. The recycled gas forms part of the trapped charge of the following engine cycle. There is a strong degree of interaction between the level of combustion products within the newly formed mixture and engine speed and load. Increasing speed shortens the duration of the valve overlap in real time, while increasing load raises the pressure-boundary of the intake system limiting the recirculating hot flows. At high speed and load conditions the increase of charge dilution with increasing valve overlap is limited.

#### **2.4. Charge dilution mass fraction – Definitions and measurements**

In the case of a gasoline engine fitted with VVT system, the dilution mass fraction is the sum of two different terms. The first one, properly named residual gas fraction, is associated with the amount of burned gas remaining inside the combustion chamber when the piston reaches the TDC of the exhaust stroke. If the exhaust valve closes before TDC, then the residual mass fraction would be given by the amount of burned gas trapped inside the cylinder at EVC. In symbols, the residual mass fraction is written as:

$$
\lambda \propto\_r = \frac{m\_r}{m\_{tot}} \tag{3}
$$

same time. Dilution mass fraction is calculated exploiting the readings from the two analysers,

( ) ( ) ( ) ( ) 2 2 2 2


% % (7)

http://dx.doi.org/10.5772/55495

11

Premixed Combustion in Spark Ignition Engines and the Influence of Operating Variables

*CO compr CO air <sup>b</sup>*

% %

*m x x*

*m x x*

= = -

*tot CO exh CO air*

In equation (7), (*x*˜*CO*2)*compr* is mole fraction of *CO*<sup>2</sup> in the unburned mixture extracted during compression; (*x*˜*CO*2)*exh* is the mole fraction in the exhaust stream, and (*x*˜*CO*2)*air* refers to the fresh intake air (this can be assumed constant at 0.03%). A full derivation of equation (7), along with validation and experience of use of the cylinder charge sampling system, can be found

Since *CO*2 is normally measured in fully dried gas streams, the outputs from the analysers are dry mole fractions and need to be converted into wet mole fractions to obtain real measure‐

ments. Heywood [6] suggests using the following expression for the correction factor:

( ) ( ) ( ) \* \*\* \* 2

*i CO CO CO*

1 0.5 / 0.74

*x mn x x x*

1

<sup>+</sup> é ù + - ë û

hydrogen to carbon ratio of the gasoline molecule. The *K* factor assumes different values if calculated using in-cylinder samples or exhaust stream ones, hence separate calculations are necessary. Data collected during the present research work show that, independently of the running conditions, wet dilution levels are 11% to 13% greater than dry dilution levels.

In an engine fitted not only with VVT system, but also with External-Exhaust Gas Recirculation system, the total mass of spent gas trapped at IVC accounts for a further source

*r IEGR EEGR b*

*tot tot tot tot mm m m*

The externally recirculated gas is commonly expressed as a fraction (or percentage) of the intake manifold stream. The formulation which allows the calculation of this quantity is:

*mm m m*

% %% % (8)

\* indicates dry mole fractions of the i-th component, while (*m* / *n*)=1.87 is the

=+ + = (9)

*b*

*x*

*i*

*2.4.2. Dilution by external exhaust gas recirculation*

(*mb* =*mr* + *mIEGR* + *mEEGR*), and the total dilution is written as:

*b*

*x*

= =

*x K*

%

with the expression:

in [22] and in [33].

In equation (8), *x*˜*<sup>i</sup>*

The second term is the Internal-Exhaust Gas Recirculation, i.e. the amount of burned gas recirculated from the exhaust port to the intake while the valves are overlapping. The associ‐ ated gas fraction is:

$$\propto\_{IEGR} = \frac{m\_{IEGR}}{m\_{tot}} \tag{4}$$

Since there are no physical ways to distinguish between *mr* and *mIEGR*, the total mass of spent gas recycled from one engine cycle to the following one is simply referred to as burned mass, *mb*. The total dilution mass fraction assumes the form:

$$\alpha\_b = \frac{m\_r}{m\_{tot}} + \frac{m\_{IEGR}}{m\_{tot}} = \frac{m\_b}{m\_{tot}} \tag{5}$$

In the previous expressions, *mtot* is the total mass trapped inside the cylinder at IVC, given by the addition of all the single contributions to the total:

$$m\_{\text{tot}} = m\_{\text{fuel}} + m\_{\text{air}} + m\_b \tag{6}$$

The total cylinder mass should also account for a small but not negligible mass of atmospheric water vapor, which can be safely assumed to be a constant fraction of *mair*.

#### *2.4.1. Measurements of dilution*

Methods to measure the cylinder charge diluent fraction are usually divided into two main categories: invasive or *in situ* techniques, and non-invasive. Invasive techniques, such as Spontaneous Raman Spectroscopy and Laser Induced Fluorescence, require physical modifi‐ cations to the engine, likely interfering with the normal combustion process [33]. The experi‐ mental data presented in the following sections have been collected using a non-invasive incylinder sampling technique, which entails the extraction of a gas sample during the compression stroke of every engine cycle, between IVC and Spark Timing. The small extracted gas stream, controlled via a high-frequency valve, is passed through a first GFC IR analyser, which can work reliably at low flow rates, to yield carbon dioxide molar concentrations within the cylinder trapped mass. A second GFC IR analyser is used to measure exhaust *CO*2, at the same time. Dilution mass fraction is calculated exploiting the readings from the two analysers, with the expression:

$$\mathbf{x}\_{b} = \frac{m\_{b}}{m\_{tot}} = \frac{\left(\tilde{\mathbf{x}}\_{CO2}\right)\_{compr} - \left(\tilde{\mathbf{x}}\_{CO2}\right)\_{air}}{\left(\tilde{\mathbf{x}}\_{CO2}\right)\_{exh} - \left(\tilde{\mathbf{x}}\_{CO2}\right)\_{air}}\tag{7}$$

In equation (7), (*x*˜*CO*2)*compr* is mole fraction of *CO*<sup>2</sup> in the unburned mixture extracted during compression; (*x*˜*CO*2)*exh* is the mole fraction in the exhaust stream, and (*x*˜*CO*2)*air* refers to the fresh intake air (this can be assumed constant at 0.03%). A full derivation of equation (7), along with validation and experience of use of the cylinder charge sampling system, can be found in [22] and in [33].

Since *CO*2 is normally measured in fully dried gas streams, the outputs from the analysers are dry mole fractions and need to be converted into wet mole fractions to obtain real measure‐ ments. Heywood [6] suggests using the following expression for the correction factor:

$$K = \frac{\tilde{\mathbf{x}}\_i}{\tilde{\mathbf{x}}\_i^\*} = \frac{1}{1 + 0.5 \left[ \left( m/n \right) \left( \tilde{\mathbf{x}}\_{\text{CO2}}^\* + \tilde{\mathbf{x}}\_{\text{CO}}^\* \right) - 0.74 \left( \tilde{\mathbf{x}}\_{\text{CO}}^\* \right) \right]} \tag{8}$$

In equation (8), *x*˜*<sup>i</sup>* \* indicates dry mole fractions of the i-th component, while (*m* / *n*)=1.87 is the hydrogen to carbon ratio of the gasoline molecule. The *K* factor assumes different values if calculated using in-cylinder samples or exhaust stream ones, hence separate calculations are necessary. Data collected during the present research work show that, independently of the running conditions, wet dilution levels are 11% to 13% greater than dry dilution levels.

#### *2.4.2. Dilution by external exhaust gas recirculation*

*r*

*<sup>m</sup>*<sup>=</sup> (3)

*<sup>m</sup>* <sup>=</sup> (4)

=+ = (5)

*mm mm tot fuel air b* = ++ (6)

*tot m*

The second term is the Internal-Exhaust Gas Recirculation, i.e. the amount of burned gas recirculated from the exhaust port to the intake while the valves are overlapping. The associ‐

*IEGR*

*m*

*tot*

Since there are no physical ways to distinguish between *mr* and *mIEGR*, the total mass of spent gas recycled from one engine cycle to the following one is simply referred to as burned mass,

*r IEGR b*

*tot tot tot mm m*

In the previous expressions, *mtot* is the total mass trapped inside the cylinder at IVC, given by

The total cylinder mass should also account for a small but not negligible mass of atmospheric

Methods to measure the cylinder charge diluent fraction are usually divided into two main categories: invasive or *in situ* techniques, and non-invasive. Invasive techniques, such as Spontaneous Raman Spectroscopy and Laser Induced Fluorescence, require physical modifi‐ cations to the engine, likely interfering with the normal combustion process [33]. The experi‐ mental data presented in the following sections have been collected using a non-invasive incylinder sampling technique, which entails the extraction of a gas sample during the compression stroke of every engine cycle, between IVC and Spark Timing. The small extracted gas stream, controlled via a high-frequency valve, is passed through a first GFC IR analyser, which can work reliably at low flow rates, to yield carbon dioxide molar concentrations within the cylinder trapped mass. A second GFC IR analyser is used to measure exhaust *CO*2, at the

water vapor, which can be safely assumed to be a constant fraction of *mair*.

*mmm*

*r*

*x*

*IEGR*

*x*

*mb*. The total dilution mass fraction assumes the form:

10 Advances in Internal Combustion Engines and Fuel Technologies

the addition of all the single contributions to the total:

*2.4.1. Measurements of dilution*

*b*

*x*

ated gas fraction is:

In an engine fitted not only with VVT system, but also with External-Exhaust Gas Recirculation system, the total mass of spent gas trapped at IVC accounts for a further source (*mb* =*mr* + *mIEGR* + *mEEGR*), and the total dilution is written as:

$$\propto\_{b} = \frac{m\_r}{m\_{\text{tot}}} + \frac{m\_{IEGR}}{m\_{\text{tot}}} + \frac{m\_{EEGR}}{m\_{\text{tot}}} = \frac{m\_b}{m\_{\text{tot}}} \tag{9}$$

The externally recirculated gas is commonly expressed as a fraction (or percentage) of the intake manifold stream. The formulation which allows the calculation of this quantity is:

$$EEGR = \frac{\dot{m}\_{EEGR}}{\dot{m}\_{man}} = \frac{\left(\tilde{\chi}\_{CO2}\right)\_{man} - \left(\tilde{\chi}\_{CO2}\right)\_{air}}{\left(\tilde{\chi}\_{CO2}\right)\_{exh} - \left(\tilde{\chi}\_{CO2}\right)\_{air}}\tag{10}$$

Symbols in the above expression retain the same meaning as before; (*x*˜*CO*2)*man* is the molar concentration of carbon dioxide in the intake manifold stream. The correlation connecting the total in-cylinder dilution and the dilution from External-EGR can be easily derived:

$$\propto\_{EEGR} = \frac{m\_{EEGR}}{m\_{tot}} = \frac{\left(1 - \chi\_b\right) \cdot EEGR}{1 - EEGR} \tag{11}$$

## **3. Combustion evolution: The mass fraction burned profile**

The evolution of the combustion process as indicated by the MFB variation is considered in the present section. Two methods of deriving this variation from measurements of Crank Angle resolved in-cylinder pressure are normally used. These are the Rassweiler and Withrow method or its variants [34, 35] and the application of the First Law of the Thermodynamics. The two approaches have been shown to yield closely comparable results in the case of stable combustion [22]. The Rassweiler and Withrow method and its inherent limitations are the main focuses here. All the experimental data presented in this section and in the following ones refer to the same research engine, unless otherwise specified. Technical specifications of this engine are given in section 4.1.

The quantity so far addressed to as MFB, is a non-dimensional mass ratio that can be ex‐ pressed as:

$$\left[\left.\chi\_{MFB}\right\rangle\_{\mathbf{r}}=\frac{\left[\left.m\_{b}\right\rangle\_{\mathbf{r}}}{m\_{f\mathbf{c}}}\right]\tag{12}$$

end of this stage an amount of charge as small as 1% has burned. During the second phase, the chemical energy release, from a stronger rate of burning, gives rise to the firing-cycle pressure trace. After peak pressure, that falls in this case at 15 CA degrees ATDC, when there is already an extensive contact between flame surface and cylinder walls, the MFB approaches

**Figure 1.** In-cylinder pressure trace for a firing cycle (bold line) and corresponding MFB profile (fine line); operating condition: engine speed N = 1500 rev/min; engine torque output T = 30 Nm. Dashed line represents the pressure trace

334 342 350 358 366 374 382 390 398 406 414

 **- CA degree**

**Motoring pressure trace**

The MFB profile provides a convenient basis for *combustion characterisation*, which divides the combustion process in its significant intervals, flame development, rapid burning and combustion termination, in the CA domain. The initial region of the curve, from the spark discharge to the point where a small but identifiable fraction of the fuel has burned, represents the period of flame development. It is common to find the Flame Development Angle defined

> 10% *ST FDA* = - J

 J

FDA covers the transition between initial laminar-like development and the period of fast burning where the charge burns in quasi-steady conditions, i.e. with a fairly constant mass flow rate through the thick reaction-front [9]. An alternative definition of the FDA, as the interval between ST and 5% MFB, is also common. Other definitions which refer to a shorter development interval (e.g. ST to 1% MFB) suffer from inaccuracies due to the low gradient of

The following combustion interval, the Rapid Burning Angle, is typically defined as the CA

(13)

0

0.2

0.4

**Mass Fraction Burned**

13

0.6

0.8

**MFB profile**

Premixed Combustion in Spark Ignition Engines and the Influence of Operating Variables

1

http://dx.doi.org/10.5772/55495

100% with progressively decreasing slope.

**Firing pressure trace**

**TDC of combustion Spark**

**N= 1500rpm T= 30Nm**

0

for the motored cycle.

5

10

**In-cylinder Pressure -**

15

20

 **bar** 25

as the CA interval between ST and 10% MFB:

the MFB profile during the initial phase of the process.

interval during which the MFB rises between 10% and 90%:

Here, *mb <sup>τ</sup>* is the mass actually burned at any instant *τ* after combustion initiates and *mfc* is the mass of fresh charge, including air and fuel, trapped inside the engine cylinder at IVC. Plotted as function of CA, the MFB profile assumes a characteristic S-shape, from 0% at ST to 100% when combustion terminates. Figure 1 shows the in-cylinder pressure trace for a firing cycle, the corresponding MFB profile and the motored pressure trace collected at the same engine speed and throttle valve setting.

During the early flame development, that in the case of figure 1 begins with the spark discharge at 26 CA degrees BTDC, the energy release from the fuel that burns is so small that the pressure rise due to combustion is insignificant; firing and motoring pressure traces are, therefore, coincident. During this period, over about 13 CA degrees, the MFB rises very slowly. At the

( ) ( ) ( ) ( ) 2 2 2 2

& % % (10)


ë û é ù <sup>=</sup> ë û (12)

*EEGR CO man CO air man CO exh CO air*

Symbols in the above expression retain the same meaning as before; (*x*˜*CO*2)*man* is the molar concentration of carbon dioxide in the intake manifold stream. The correlation connecting the

> (1 ) 1 *EEGR b*

*m x EEGR*

*m EEGR*

The evolution of the combustion process as indicated by the MFB variation is considered in the present section. Two methods of deriving this variation from measurements of Crank Angle resolved in-cylinder pressure are normally used. These are the Rassweiler and Withrow method or its variants [34, 35] and the application of the First Law of the Thermodynamics. The two approaches have been shown to yield closely comparable results in the case of stable combustion [22]. The Rassweiler and Withrow method and its inherent limitations are the main focuses here. All the experimental data presented in this section and in the following ones refer to the same research engine, unless otherwise specified. Technical specifications of this engine

The quantity so far addressed to as MFB, is a non-dimensional mass ratio that can be ex‐

*b*

*m*

é ù

*m* t

Here, *mb <sup>τ</sup>* is the mass actually burned at any instant *τ* after combustion initiates and *mfc* is the mass of fresh charge, including air and fuel, trapped inside the engine cylinder at IVC. Plotted as function of CA, the MFB profile assumes a characteristic S-shape, from 0% at ST to 100% when combustion terminates. Figure 1 shows the in-cylinder pressure trace for a firing cycle, the corresponding MFB profile and the motored pressure trace collected at the same

During the early flame development, that in the case of figure 1 begins with the spark discharge at 26 CA degrees BTDC, the energy release from the fuel that burns is so small that the pressure rise due to combustion is insignificant; firing and motoring pressure traces are, therefore, coincident. During this period, over about 13 CA degrees, the MFB rises very slowly. At the

*fc*

*MFB*

t

*x*

*m x x*

*m x x* - = = - % % &

total in-cylinder dilution and the dilution from External-EGR can be easily derived:

*tot*

**3. Combustion evolution: The mass fraction burned profile**

*EEGR*

12 Advances in Internal Combustion Engines and Fuel Technologies

*EEGR*

*x*

are given in section 4.1.

engine speed and throttle valve setting.

pressed as:

**Figure 1.** In-cylinder pressure trace for a firing cycle (bold line) and corresponding MFB profile (fine line); operating condition: engine speed N = 1500 rev/min; engine torque output T = 30 Nm. Dashed line represents the pressure trace for the motored cycle.

end of this stage an amount of charge as small as 1% has burned. During the second phase, the chemical energy release, from a stronger rate of burning, gives rise to the firing-cycle pressure trace. After peak pressure, that falls in this case at 15 CA degrees ATDC, when there is already an extensive contact between flame surface and cylinder walls, the MFB approaches 100% with progressively decreasing slope.

The MFB profile provides a convenient basis for *combustion characterisation*, which divides the combustion process in its significant intervals, flame development, rapid burning and combustion termination, in the CA domain. The initial region of the curve, from the spark discharge to the point where a small but identifiable fraction of the fuel has burned, represents the period of flame development. It is common to find the Flame Development Angle defined as the CA interval between ST and 10% MFB:

$$FDA = \mathcal{G}\_{10\%} - \mathcal{G}\_{ST} \tag{13}$$

FDA covers the transition between initial laminar-like development and the period of fast burning where the charge burns in quasi-steady conditions, i.e. with a fairly constant mass flow rate through the thick reaction-front [9]. An alternative definition of the FDA, as the interval between ST and 5% MFB, is also common. Other definitions which refer to a shorter development interval (e.g. ST to 1% MFB) suffer from inaccuracies due to the low gradient of the MFB profile during the initial phase of the process.

The following combustion interval, the Rapid Burning Angle, is typically defined as the CA interval during which the MFB rises between 10% and 90%:

$$RBA = \mathcal{G}\_{90\%} - \mathcal{G}\_{10\%} \tag{14}$$

Constant-volume bomb experiments have also shown that the pressure increment due to combustion, the total mass being constant, is inversely proportional to volume. In order to draw a second analogy with engine combustion, the combustion pressure rise at each step, calculated as (*ΔP* −*ΔPV* ), is multiplied by a volume ratio which eliminates the effects of volume

1 1 *c V*

*P PP*

J J® +

*ref V*

é ù =D D ë û å å (19)

® + éù é ù D = D -D ëû ë û (18)

*V* J

Premixed Combustion in Spark Ignition Engines and the Influence of Operating Variables

http://dx.doi.org/10.5772/55495

15

J J

allows determining the pressure increments due to combustion as if they all occur into the same volume *Vref* . The reference volume is taken equal to the clearance volume, i.e. the combustion chamber volume when the piston is at TDC. The relation that gives the MFB as a

> / *ST ST*

The method discussed above provides a robust platform to extract combustion evolution information from sensors data which are routinely acquired. Nevertheless, its accuracy is questionable as necessary constrains such as the EOC are not easily identifiable, and because it accounts for heat losses to the cylinder walls only implicitly, by selecting appropriate

In theory, the polytropic index which figures in equation (17) should change continuously during combustion. However, this is not practical and an easier strategy of indices determi‐ nation must be adopted. In the work presented here, two different values of the polytropic index are used for intervals in the compression and power strokes, respectively. The evaluation of the MFB curve proceeds by successive iterations, until appropriate values of the polytropic indexes, in connection with the determination of EOC, are established. The sensitivity of pressure increments to these indices increases with pressure and then is emphasized after TDC, when the in-cylinder pressure reaches its maximum. While the sensitivity of the MFB profile to the compression index is relatively low, the selection of the expansion index is more important. During compression the unburned mixture roughly undergoes a polytropic process that begins at IVC. In this work the polytropic compression index is calculated as the negative of the slope of the experimental [log V, log P] diagram over 30 consecutive points before ST, and maintained unvaried up to TDC. During the expansion stroke the polytropic index varies due to several concurrent phenomena, including heat transfer, work exchange and turbulence variation. In theory, it increases approaching an asymptotic value just before EVO. As

*MFB <sup>c</sup> <sup>c</sup> x PP* J

EOC indicates the CA location of End Of Combustion, corresponding to 100% MFB.

J

J

*3.1.1. Polytropic indexes and EOC condition for MFB calculation*

polytropic indices for compression and expansion strokes.

*EOC*

 J

changes. The relation:

function of CA is finally obtained:

The selection of 90% MFB as limiting point is dictated by convenience since the final stage of combustion is difficult to identify. During the so-called combustion termination the chemical energy release from the fuel that burns is comparable to other heat transfer processes that occur at the same time; during this stage the MFB increases only slightly over a large number of CA degrees.

#### **3.1. The rassweiler and withrow method**

In the present section and in the following ones, the Rassweiler and Withrow method has been used for MFB calculations from ensemble-averaged experimental pressure records and volume variation data. The method is well established due to ease of implementation, which allows real-time processing and because it shows good intrinsic tolerance to pressure signal noise across wide ranges of engine operating conditions [35]. Its rationale comes from observations of constant-volume bomb explosions, where the fractional mass of burned charge has been seen to be approximately equal to the fractional pressure rise. If *Ptot* and *P <sup>τ</sup>* are, respectively, pressure at the end of combustion and at a generic time *τ*, this equality can be written as:

$$\left[\left.\chi\_{MFB}\right\rangle\_{\rm r} = \frac{\left[\left.m\_b\right\rangle\_{\rm r}\right]\_{\rm r}}{m\_{\rm f^c}} \approx \frac{\left[\left.P\right\rangle\_{\rm r}\right]}{P\_{\rm tot}}\tag{15}$$

More precisely, the pressure rise due to combustion is proportional to chemical heat release rather than to fractional burned mass, but MFB calculations using the above approximation are consistently in agreement with those from thermodynamic models [36].

In order to apply to engine-like conditions the analogy with constant-volume bombs, the total pressure rise measured across a small CA interval is divided into contributions due only to combustion and only to volume variation:

$$
\Delta P = \Delta P\_c + \Delta P\_V \tag{16}
$$

In each CA step, increments due to piston motion are calculated assuming that pressure undergoes a polytropic process:

$$\left[\left.\Delta P\_V\right]\_{\mathcal{g}\to\mathcal{g}+1} = P\_{\mathcal{g}}\left|\left(\frac{V\_{\mathcal{g}}}{V\_{\mathcal{g}+1}}\right)^n - 1\right|\right.\tag{17}$$

Constant-volume bomb experiments have also shown that the pressure increment due to combustion, the total mass being constant, is inversely proportional to volume. In order to draw a second analogy with engine combustion, the combustion pressure rise at each step, calculated as (*ΔP* −*ΔPV* ), is multiplied by a volume ratio which eliminates the effects of volume changes. The relation:

$$
\left[\begin{array}{c}
\Delta P\_c \end{array}\right]\_{\mathcal{S}\rightarrow\mathcal{S}+1} = \left[\begin{array}{c}
\Delta P-\Delta P\_V \end{array}\right]\_{\mathcal{S}\rightarrow\mathcal{S}+1} \frac{V\_{\mathcal{S}}}{V\_{ref}}\tag{18}
$$

allows determining the pressure increments due to combustion as if they all occur into the same volume *Vref* . The reference volume is taken equal to the clearance volume, i.e. the combustion chamber volume when the piston is at TDC. The relation that gives the MFB as a function of CA is finally obtained:

$$\left[\left.\propto\_{MFB}\right]\_{\mathcal{B}}\right]\_{\mathcal{B}} = \sum\_{\mathcal{B}\_{ST}}^{\mathcal{B}} \Delta P\_c \quad / \quad \sum\_{\mathcal{B}\_{ST}}^{EOC} \Delta P\_c \tag{19}$$

EOC indicates the CA location of End Of Combustion, corresponding to 100% MFB.

#### *3.1.1. Polytropic indexes and EOC condition for MFB calculation*

*RBA* 90% 10% = - J

degrees.

written as:

**3.1. The rassweiler and withrow method**

14 Advances in Internal Combustion Engines and Fuel Technologies

combustion and only to volume variation:

undergoes a polytropic process:

 J

The selection of 90% MFB as limiting point is dictated by convenience since the final stage of combustion is difficult to identify. During the so-called combustion termination the chemical energy release from the fuel that burns is comparable to other heat transfer processes that occur at the same time; during this stage the MFB increases only slightly over a large number of CA

In the present section and in the following ones, the Rassweiler and Withrow method has been used for MFB calculations from ensemble-averaged experimental pressure records and volume variation data. The method is well established due to ease of implementation, which allows real-time processing and because it shows good intrinsic tolerance to pressure signal noise across wide ranges of engine operating conditions [35]. Its rationale comes from observations of constant-volume bomb explosions, where the fractional mass of burned charge has been seen to be approximately equal to the fractional pressure rise. If *Ptot* and *P <sup>τ</sup>* are, respectively, pressure at the end of combustion and at a generic time *τ*, this equality can be

*b*

*fc tot*

 t

ë û ëû é ù = » ë û (15)

*c V* D =D +D *PP P* (16)

(17)

*m P*

é ù éù

*m P* t

More precisely, the pressure rise due to combustion is proportional to chemical heat release rather than to fractional burned mass, but MFB calculations using the above approximation

In order to apply to engine-like conditions the analogy with constant-volume bombs, the total pressure rise measured across a small CA interval is divided into contributions due only to

In each CA step, increments due to piston motion are calculated assuming that pressure

1

*V*

J

*V* J

1 *n*

*MFB*

t

are consistently in agreement with those from thermodynamic models [36].

1

J

 ® + <sup>+</sup> é ù æ ö D= - ê ú é ù ç ÷ ë û ê ú è ø ë û

*P P*

J J

*V*

*x*

(14)

The method discussed above provides a robust platform to extract combustion evolution information from sensors data which are routinely acquired. Nevertheless, its accuracy is questionable as necessary constrains such as the EOC are not easily identifiable, and because it accounts for heat losses to the cylinder walls only implicitly, by selecting appropriate polytropic indices for compression and expansion strokes.

In theory, the polytropic index which figures in equation (17) should change continuously during combustion. However, this is not practical and an easier strategy of indices determi‐ nation must be adopted. In the work presented here, two different values of the polytropic index are used for intervals in the compression and power strokes, respectively. The evaluation of the MFB curve proceeds by successive iterations, until appropriate values of the polytropic indexes, in connection with the determination of EOC, are established. The sensitivity of pressure increments to these indices increases with pressure and then is emphasized after TDC, when the in-cylinder pressure reaches its maximum. While the sensitivity of the MFB profile to the compression index is relatively low, the selection of the expansion index is more important. During compression the unburned mixture roughly undergoes a polytropic process that begins at IVC. In this work the polytropic compression index is calculated as the negative of the slope of the experimental [log V, log P] diagram over 30 consecutive points before ST, and maintained unvaried up to TDC. During the expansion stroke the polytropic index varies due to several concurrent phenomena, including heat transfer, work exchange and turbulence variation. In theory, it increases approaching an asymptotic value just before EVO. As suggested by Karim [37], the EOC associates the condition *ΔPc* =0 with an expansion index which settles to an almost constant value. Provided a reasonable condition is given to deter‐ mine the EOC, the correct expansion index would be the one that, when combustion is over, maintains the MFB profile steadily at 100% till EVO: *the zero combustion-pressure condition* [35]. In this work, the expansion index is estimated with an iterative procedure where, starting from a reference value (e.g. 1.3), the index is progressively adjusted together with the EOC, until the MFB profile acquires a *reasonable* S-shape, which meets the requirement of the zero combustion-pressure condition. Several methods are reported in the literature to determine the EOC; the first negative and the sum negative methods, for example, assume that EOC occurs when one or three consecutive negative values of *ΔPc* are found. In this work, the combustion process is supposed to terminate when *ΔPc* becomes a negligible fraction (within 0.2%) of the total pressure increment *ΔP* for 3 CA-steps consecutively.

#### *3.1.2. Other methods of estimation of the expansion index*

Other methods have been proposed for the evaluation of *n*exp. One calculates the index as the slope of the log-log indicator diagram over narrow intervals before EVO. Although this approach avoids the EOC determination, experimental results show that the calculations are sensitive to the chosen interval and, in general, combustion duration is overestimated. As an improvement to this method, *n*exp has been calculated as the value that gives average *ΔPc* equal to zero after combustion terminates, satisfying the zero combustion-pressure condition [35]. Again, this approach seems to be sensitive to the interval over which the average *ΔPc* is evaluated, reflecting pressure measurements noise and the fact that often *n*exp does not settle properly before EVO. Figure 2 directly compares three different methods of expansion index determination for engine speed of 1900 rev/min and torque of 40 Nm (similar results are obtained at different operating conditions): with the view that the iterative method of *n*exp estimate yields accurate MFB characteristics (which, for stable combustion, are consistently similar to those from thermodynamics models [22]), the modification proposed in [35] tends to overestimate combustion duration during the rapid stage and especially during the termination stage, with the effect of delaying the EOC. The method for estimating the expan‐ sion index is crucially important as different methods may cause over 40% variation in the calculated RBA.

referencing to a known value. Since thermal-shock is driven by combustion, it would be preferable to perform cylinder pressure referencing when the artificial variability due to temperature changes is at a minimum, a circumstance which is likely to occur at the end of the intake stroke [40]. Nevertheless, the thermally induced drift persists throughout the whole engine cycle, assigning uncertainty to the experimental measurements. Payri et al. [39] account for a value of pressure accuracy of ±0.15 bar, estimated as maximum pressure difference at BDC of induction. Studies carried out by the Author [22] have shown that a value of intra-cycle pressure drift (calculated as difference between transducer BDC outputs at the beginning and at the end of single cycles) of ± 0.1 bar (with standard deviation of 0.055 bar) represents a

338 348 358 368 378 388 398 408 418 428 438 448 458 468

**Iterative method**

Premixed Combustion in Spark Ignition Engines and the Influence of Operating Variables

**SAE 900351 - interval 30 CA degree SAE 900351 - interval 10 CA degree**

http://dx.doi.org/10.5772/55495

17

 **- CA degree**

When MFB profiles are built applying the Rassweiler and Withrow method to ensembleaveraged in-cylinder pressure records, at least two sources of errors can be considered: pressure measurements inaccuracy, but also the consequential polytropic compression index variation. Expansion index and EOC location are also affected by pressure variation but, if the iterative optimisation technique described above is used, these cannot be enumerated among the causes of uncertainties. The compression index variation is a linear function of the pressure variation at BDC of induction, almost independently of engine speed and load. A variation of +10% in BDC pressure induces a reduction of the compression index of about -1.5% [22]. Further studies on the effects of pressure drift (used as an offset) on the MFB profile, have shown that the region mostly affected is the flame development interval between ST and 10% MFB. For a pressure offset of ±0.1 bar, typical values of MFB percentage variation are likely to be around ±6% at 10% MFB for low engine load (IMEP = 2.5 bar); the error reduces propor‐ tionately at increasing load (typically ±1.5% at 10% MFB for IMEP = 6 bar). After 10% MFB, the

realistic average estimate of the potential inaccuracy of the in-cylinder pressure.

**Figure 2.** MFB profiles built using three different methods of expansion index evaluation.

0

0.2

0.4

0.6

**Mass Fraction Burned**

0.8

1

**N= 1500 rpm T= 40 Nm**

**Spark**

**TDC of combustion**

**IVO= +06 CA BTDC EVC= + 06 CA ATDC**

#### **3.2. Estimated errors in the MFB profile**

The calculated burning characteristics of an engine, including the MFB profile, may be affected by measurements and calculation errors. Most of the potential inaccuracies are associated with the determination of the absolute in-cylinder pressure. The adoption of ensemble-averaged pressure trace, which as in the present work should be based on the acquisition of a minimum of 100 individual cycles [38], is beneficial to diminish the cyclical dispersion errors (inter-cycle pressure drift) [39] and signal noise. The major source of cylinder pressure error is indeed associated with thermal-shock and can be accounted for in terms of short-term or intra-cycle pressure drifts. Pressure sensors do not measure absolute pressure and the sensor signal need

**Figure 2.** MFB profiles built using three different methods of expansion index evaluation.

suggested by Karim [37], the EOC associates the condition *ΔPc* =0 with an expansion index which settles to an almost constant value. Provided a reasonable condition is given to deter‐ mine the EOC, the correct expansion index would be the one that, when combustion is over, maintains the MFB profile steadily at 100% till EVO: *the zero combustion-pressure condition* [35]. In this work, the expansion index is estimated with an iterative procedure where, starting from a reference value (e.g. 1.3), the index is progressively adjusted together with the EOC, until the MFB profile acquires a *reasonable* S-shape, which meets the requirement of the zero combustion-pressure condition. Several methods are reported in the literature to determine the EOC; the first negative and the sum negative methods, for example, assume that EOC occurs when one or three consecutive negative values of *ΔPc* are found. In this work, the combustion process is supposed to terminate when *ΔPc* becomes a negligible fraction (within

Other methods have been proposed for the evaluation of *n*exp. One calculates the index as the slope of the log-log indicator diagram over narrow intervals before EVO. Although this approach avoids the EOC determination, experimental results show that the calculations are sensitive to the chosen interval and, in general, combustion duration is overestimated. As an improvement to this method, *n*exp has been calculated as the value that gives average *ΔPc* equal to zero after combustion terminates, satisfying the zero combustion-pressure condition [35]. Again, this approach seems to be sensitive to the interval over which the average *ΔPc* is evaluated, reflecting pressure measurements noise and the fact that often *n*exp does not settle properly before EVO. Figure 2 directly compares three different methods of expansion index determination for engine speed of 1900 rev/min and torque of 40 Nm (similar results are obtained at different operating conditions): with the view that the iterative method of *n*exp estimate yields accurate MFB characteristics (which, for stable combustion, are consistently similar to those from thermodynamics models [22]), the modification proposed in [35] tends to overestimate combustion duration during the rapid stage and especially during the termination stage, with the effect of delaying the EOC. The method for estimating the expan‐ sion index is crucially important as different methods may cause over 40% variation in the

The calculated burning characteristics of an engine, including the MFB profile, may be affected by measurements and calculation errors. Most of the potential inaccuracies are associated with the determination of the absolute in-cylinder pressure. The adoption of ensemble-averaged pressure trace, which as in the present work should be based on the acquisition of a minimum of 100 individual cycles [38], is beneficial to diminish the cyclical dispersion errors (inter-cycle pressure drift) [39] and signal noise. The major source of cylinder pressure error is indeed associated with thermal-shock and can be accounted for in terms of short-term or intra-cycle pressure drifts. Pressure sensors do not measure absolute pressure and the sensor signal need

0.2%) of the total pressure increment *ΔP* for 3 CA-steps consecutively.

*3.1.2. Other methods of estimation of the expansion index*

16 Advances in Internal Combustion Engines and Fuel Technologies

calculated RBA.

**3.2. Estimated errors in the MFB profile**

referencing to a known value. Since thermal-shock is driven by combustion, it would be preferable to perform cylinder pressure referencing when the artificial variability due to temperature changes is at a minimum, a circumstance which is likely to occur at the end of the intake stroke [40]. Nevertheless, the thermally induced drift persists throughout the whole engine cycle, assigning uncertainty to the experimental measurements. Payri et al. [39] account for a value of pressure accuracy of ±0.15 bar, estimated as maximum pressure difference at BDC of induction. Studies carried out by the Author [22] have shown that a value of intra-cycle pressure drift (calculated as difference between transducer BDC outputs at the beginning and at the end of single cycles) of ± 0.1 bar (with standard deviation of 0.055 bar) represents a realistic average estimate of the potential inaccuracy of the in-cylinder pressure.

When MFB profiles are built applying the Rassweiler and Withrow method to ensembleaveraged in-cylinder pressure records, at least two sources of errors can be considered: pressure measurements inaccuracy, but also the consequential polytropic compression index variation. Expansion index and EOC location are also affected by pressure variation but, if the iterative optimisation technique described above is used, these cannot be enumerated among the causes of uncertainties. The compression index variation is a linear function of the pressure variation at BDC of induction, almost independently of engine speed and load. A variation of +10% in BDC pressure induces a reduction of the compression index of about -1.5% [22]. Further studies on the effects of pressure drift (used as an offset) on the MFB profile, have shown that the region mostly affected is the flame development interval between ST and 10% MFB. For a pressure offset of ±0.1 bar, typical values of MFB percentage variation are likely to be around ±6% at 10% MFB for low engine load (IMEP = 2.5 bar); the error reduces propor‐ tionately at increasing load (typically ±1.5% at 10% MFB for IMEP = 6 bar). After 10% MFB, the MFB variation reduces consistently, reaching very small values, perhaps 1% or 0.5%, at 90% MFB. The error study by Brunt et al. [41] shows similar nature and magnitude of errors.
