Diesel combustion uses the thermal energy created from the compression of air and fuel to create an ignition. This ignition leads to the burning of fuel injected within a combustion chamber of the engine. It differs from gasoline combustion, which uses combustion of air and fuel ignited by a spark plug. A diesel engine operates on either a two-stroke or four-stroke method and uses a fuel refined from petroleum, biomass, oil waste or other sources.
The air and fuel mixture is very important to making diesel combustion work efficiently. Air is compressed into a combustion chamber with approximately a 600 pounds per square inch (about 40 bar) pressure. Due to the sheer compression, the air within the chamber heats up to approximately 1000°F (roughly 550°C).
Diesel fuel is then pushed into the chamber with the compressed air using a fuel injector. The injector itself breaks down the fuel into tiny droplets, ensuring it is evenly distributed throughout the chamber. The heat vaporizes the droplets, causing combustion and the pressure pushes the piston outward, powering the crankshaft. This gives the diesel engine its typical “knocking” sound.
The benefit of diesel compression is the fact that the system operates without a separate ignition system, as is the case in gasoline engines. The level of compression can be increased within a diesel engine in order to increase fuel efficiency. This can occur without the threat of damage to the cylinder. In addition, the fact that only air is compressed before fuel is introduced means that there is no threat of premature detonation, again damaging the engine.
The diesel engine was invented in the late 1800s. Having a love for engine design, refrigerator engineer Rudolf Diesel began exploring the concept of the internal combustion engine in the late 1880s. Diesel developed the first engine that operated without a spark, and filed for a patent in 1894. Within three years, he successfully demonstrated the power and efficiency of diesel combustion. The patent was approved in 1898.
Diesel's discovery and subsequent invention was the first to deal with the laws of thermodynamics in an internal combustion engine. Diesel combustion uses the natural physical process of heat transference that was a very creative manner for the time period. In addition, the inventor had an overall sociological intent: Diesel wanted the engine to help independent industry to be able to compete with larger businesses.
Combustion in Diesel Engines
• Components of Combustion Process
• Heat Release Rates in DI Engines
• Three Phases of Diesel Combustion
• Conceptual Diesel Combustion Model
• Spray Formation
• Flame Lift-Off
• φ-T Maps
Components of Combustion Process
Diesel engines have an excellent reputation for their low fuel consumption, reliability, and durability characteristics. They are also known for their extremely low hydrocarbon and carbon monoxide emissions. However, they have also been rejected by many for their odorous and sooty exhaust that is also characterized with high nitric oxide and particulate matter emissions. Since performance, fuel consumption, and emitted pollutants result from the combustion process, it is necessary first to understand the mechanisms of combustion in diesel engines if we are to improve it.
Combustion in diesel engines is very complex and until recently, its detailed mechanisms were not well understood. For decades its complexity seemed to defy researchers’ attempts to unlock its many secrets despite the availability of modern tools such as high speed photography used in “transparent” engines, computational power of contemporary computers, and the many mathematical models designed to mimic combustion in diesel engines. The application of laser-sheet imaging to the conventional diesel combustion process in the 1990s was key to greatly increasing the understanding of this process. This paper will review the most established combustion model for the conventional diesel engine to help readers gain an appreciation of diesel combustion and how it impacts performance and emission formation.
The basic premise of diesel combustion is its unique way of releasing the chemical energy stored in the fuel. To perform this process, oxygen must be made available to the fuel in a specific manner to facilitate combustion. One of the most important aspects of this process is the mixing of fuel and air, which is a process often referred to as mixture preparation.
In diesel engines, fuel is often injected into the engine cylinder near the end of the compression stroke, just a few crank angle degrees before top dead center. The liquid fuel is usually injected at high velocity as one or more jets through small orifices or nozzles in the injector tip. It atomizes into small droplets and penetrates into the combustion chamber. The atomized fuel absorbs heat from the surrounding heated compressed air, vaporizes, and mixes with the surrounding high-temperature high-pressure air. As the piston continues to move closer to top dead center (TDC), the mixture (mostly air) temperature reaches the fuel’s ignition temperature. Instantaneous ignition of some premixed fuel and air occurs after the ignition delay period. This instantaneous ignition is considered the start of combustion (also the end of the ignition delay period) and is marked by a sharp cylinder pressure increase as combustion of the fuel-air mixture takes place. Increased pressure resulting from the premixed combustion compresses and heats the unburned portion of the charge and shortens the delay before its ignition. It also increases the evaporation rate of the remaining fuel. Atomization, vaporization, fuel vapor-air mixing, and combustion continue until all the injected fuel has combusted.
Diesel combustion is characterized by lean overall A/F ratio. The lowest average A/F ratio is often found at peak torque conditions. To avoid excessive smoke formation, A/F ratio at peak torque is usually maintained above 25:1, well above the stoichiometric (chemically correct) equivalence ratio. In turbocharged diesel engines the A/F ratio at idle may exceed 160:1. Therefore, excess air present in the cylinder after the fuel has combusted continues to mix with burning and already burned gases throughout the combustion and expansion processes. At the opening of the exhaust valve, excess air along with the combustion products are exhausted, which explains the oxidizing nature of diesel exhaust. Although combustion occurs after vaporized fuel mixes with air, forms a locally rich but combustible mixture, and the proper ignition temperature is reached, the overall A/F ratio is lean. In other words, the majority of the air inducted into the cylinder of a diesel engine is compressed and heated, but never engages in the combustion process. Oxygen in the excess air helps oxidize gaseous hydrocarbons and carbon monoxide, reducing them to extremely small concentrations in the exhaust gas.
The following factors play a primary role in the diesel combustion process:
• The inducted charge air, its temperature, and its kinetic energy in several dimensions.
• The injected fuel’s atomization, spray penetration, temperature, and chemical characteristics.
While these two factors are most important, there are other parameters that may dramatically influence them and therefore play a secondary, but still important role in the combustion process. For instance:
• Intake port design, which has a strong influence on charge air motion (especially as it enters the cylinder) and ultimately the mixing rate in the combustion chamber. The intake port design may also influence charge air temperature. This may be accomplished by heat transfer from the water jacket to the charge air through the intake port surface area.
• Intake valve size, which controls the total mass of air inducted into the cylinder in a finite amount of time.
• Compression ratio, which influences fuel vaporization and consequently mixing rate and combustion quality.
• Injection pressure, which controls the injection duration for a given nozzle hole size.
• Nozzle hole geometry (length/diameter), which controls the spray penetration as well as atomization.
• Spray geometry, which directly impacts combustion quality through air utilization. For instance, a larger spray cone angle may place the fuel on top of the piston, and outside the combustion bowl in open chamber DI diesel engines. This condition would lead to excessive smoke (incomplete combustion) because of depriving the fuel of access to the air available in the combustion bowl (chamber). Wide cone angles may also cause the fuel to be sprayed on the cylinder walls, rather than inside the combustion bowl where it is required. Fuel sprayed on the cylinder wall will eventually be scraped downward to the oil sump where it will shorten the lube oil life. As the spray angle is one of the variables that impacts the rate of mixing of air into the fuel jet near the outlet of the injector, it can have a significant impact on the overall combustion process.
• Valve configuration, which controls the injector position. Two-valve systems force an inclined injector position, which implies uneven spray arrangement that leads to compromised fuel/air mixing. On the other hand, four-valve designs allow for vertical injector installation, symmetric fuel spray arrangement and equal access to the available air by each of the fuel sprays.
• Top piston ring position, which controls the dead space between the piston top land (area between top piston ring groove and the top of the piston crown), and the cylinder liner. This dead space/volume traps air that is compressed during the compression stroke and expands without ever engaging in the combustion process.
It is therefore important to realize that the combustion system of the diesel engine is not limited to the combustion bowl, injector sprays, and their immediate surroundings. Rather, it includes any part, component, or system that may affect the final outcome of the combustion process.
Heat Release Rates in DI Engines
Many researchers have studied cylinder pressure traces to determine heat release rates Pressure traces obtained in open-chamber diesel engines are usually analyzed for this purpose. Figure 1 shows an example of a rate of net heat release diagram (Qn) together with cylinder pressure (p), and fuel injection rate (mfi). Details of the work done to obtain these traces were given by Lyn .The initial sharp rise in the heat release rate results from burning the premixed portion of the fuel. During the ignition delay (I.D.) evaporation from the spray forms a fuel rich fuel-vapor-air mixture, first at the sides and then at the tip of the fuel jet. The complete evaporation of fuel droplets is rapid and takes place in much less time than the I.D. The negative heat release after the start of injection, at about 20° BTDC, is due mainly to the heat transfer from the hot air to the evaporating liquid fuel.
Ignition Delay
Ignition delay in diesel engine combustion is the time between the start of injection and the start of detectable combustion (a → b in Figure 2). Criteria used to quantify the start of combustion include:
• abrupt changes in cylinder pressure
• light emission from combustion reactions
• temperature rise due to combustion
• combustion of a defined amount of fuel
• a fixed point on the heat release rate curve
One common definition for ignition delay is the time between start-of-injection (SOI) and the time at which the net rate of heat release returns to zero. Net heat release rate is normally negative shortly after injection due to liquid fuel heating and evaporation (Figure 1). When sufficient heat has been released by the combustion process to produce a positive net heat release rate, the ignition delay period is considered to be over. Stated another way, it is the time when the integrated amount of heat released by reactions becomes equal to that absorbed by the evaporated fuel.
Depending on how start-of-injection is measured, ignition delay may include injector lag—the time taken between the injector receiving the signal driving it open and fuel exiting the injector nozzle into the combustion chamber. For example, in one study using a HEUI injector with a maximum injection pressure of 142 MPa, an injector lag of about 1.5 ms was measured .This was several times that of the actual ignition delay period measured from the time fuel exited the injector until detectable combustion commenced. Others have measured injection lags in common rail systems ranging from 0.30 to 0.75 ms .
The duration of the ignition delay is an important criterion. It has a significant impact on the combustion process, mechanical stresses, engine noise and exhaust emissions. In contrast with gasoline engines, where combustion starts by an electrically energized spark at one location, combustion in diesel engines starts by autoignition at numerous locations in the combustion chamber. Figure 3 gives a summary of the physical and chemical steps before and after autoignition.
The physical processes involved in the ignition delay period (ID) are:
• spray break up and droplet formation
• fuel and air mixing
• heating of the liquid fuel and evaporation
• mixing of the vapor and air to form a combustible mixture
These steps are often referred to as mixture preparation, although this term may include more than just the steps listed above. Included in the mixture preparation is the effect of air motion due to intake port design, heat transfer to the air flow through the intake port, air temperature and any other parameters that may affect the quality of the fuel-air mixture. Most, if not all, of these steps are physical in nature and are generally completed in an extremely short period of time.
The chemical processes that take place in the ignition delay period are:
• pre-ignition reactions that break-down the hydrocarbon fuel and generate radicals
• localized ignition that takes place in several areas within the combustion chamber.
While the chemical processes start after the fuel vapor makes contact with the air, in the very early stages of injection the mass of fuel vapor which undergoes chemical reaction is too small to cause any detectable combustion phenomena. The early stages of pre-ignition can be considered to be dominated by the physical processes that result in the formation of a combustible mixture and the later stages by the chemical changes which lead to autoignition.
While it is difficult to draw a distinct line separating the physical and chemical processes because they overlap, an estimate can often be made of the point at which the chemical process starts to dominate. Figure 4 breaks down the ignition delay period into two quantifiable periods. The period τ1 represents the period before exothermic chemical reactions have had a measurable effect on cylinder pressure. The linear drop in cylinder pressure is dominated by factors affecting the physical delay component of the ignition delay period. The period τ1 ends when the cylinder pressure for a reacting spray (solid line) separates from that of an identical spray injected into an inert nitrogen atmosphere (lower dotted line). The period τ2 represents the period when exothermic reactions take over and allow the cylinder pressure to recover from the heat absorbed by the evaporating fuel. The combination of these two periods, τ1 + τ2, represents the ignition delay period
In addition to the fuel type (chemical structure), temperature, and pressure conditions, ignition delay is also affected by injection pressure and injector nozzle orifice diameter.
Ignition delay is often expressed as an Arrhenius equation of the form:
(1)τid = A p-n exp(EA/RT)
where τid is the ignition delay time often expressed in ms, A and n are constants, EA is the apparent activation energy for the autoignition process, T is the absolute temperature of the fuel-air mixture and R is the universal gas constant. Values for the parameters in Equation (1) can be found in the literature.
Experiments with high levels of charge dilution have also shown that the O2 concentration has a significant impact on ignition delay.
(2) τid = 12.254 p-1 XO2-1.2 exp(3242.4/T)
where XO2 is mole fraction of O2 in the intake charge, p is in kPa, T in K and τ in ms. The dependency on O2 reflects the slowing down of chemical reactions in dilute mixtures and decreased temperature rise from early low temperature reactions due to the heat capacity of the diluent molecules.
Such correlations should be used with caution as they do not separate the physical and chemical processes contributing to ignition delay. While differences in engine design (such as that of the fuel injection system) may have less of an impact on chemical delay at a given temperature and pressure it can have a significant impact on the physical delay. Application to engine designs that are significantly different can result in significant errors in ignition delay estimated by these equations.
The cetane number of a fuel is commonly used to quantify the ignition delay characteristics of a fuel. Low cetane number fuels have a longer ignition delay and more fuel is injected before ignition occurs. This produces a large premixed burn and very rapid burning rates once combustion starts, with high rates of pressure rise and high peak pressures. Under extreme conditions when autoignition of most of the injected fuel occurs, an audible knocking sound (‘diesel knock’) occurs. Higher cetane number fuels have shorter ignition delays and ignition occurs before most of the fuel is injected. The rates of heat release and pressure rise are then controlled primarily by the rate of injection and fuel-air mixing, and smoother engine operation results.
Cetane number, as well as other methods developed to characterize the ignition quality of diesel fuels are described in the papers on diesel fuel and, in more detail, on ignition quality testing.
Premixed Combustion
The term premixed combustion refers to the rapid premixed combustion of a portion of the fuel injected during the ignition delay period. This period is indicated as b→c in Figure 2 [fig]. This portion of the fuel would have undergone atomization, evaporation, and the pre-ignition chemical reactions. It would also have mixed with air to form a fuel-rich mixture that is ready to ignite once the proper temperature (autoignition temperature) is reached. When autoignition occurs, the premixed fuel burns at a very high rate, producing high temperature and high rates of pressure rise in the combustion chamber. The rate of premixed burning is governed mainly by chemical kinetics.
The characteristic noise of the diesel engine is also associated with this premixed burn phase. It is generally accepted that the rate of pressure rise resulting from the premixed combustion is proportional to the noise intensity in diesel engines.
The remaining fuel that is not involved in the premixed combustion has not been injected, evaporated, or mixed with air and may be too lean or too rich to burn.
Engine speed, load and injection timing can all affect the proportion of fuel burned in this premixed phase. An examination of several correlations found that the mass of fuel burned in the premixed burn phase (mf-id) increased linearly with the product of engine speed (N) and ignition delay time.
(4)mf-id ~ N • τid
This can also be stated as:
(5)mf-id ~ θid
where θid is the ignition delay in crank angle degrees.
Rate-Controlled Combustion
The balance of the fuel that has not participated in the premixed combustion phase represents the bulk of the fuel consumed during a complete cycle. In the rate-controlled combustion phase, the consumption rate of this fuel is controlled by its rate of injection and subsequent mixing with air. This phase is characterized by a lower heat release peak than that reached in the premixed phase. The rate-controlled combustion phase is represented by the curve between c→d in Figure 2 [fig].
This phase is also referred to as mixing-controlled combustion or diffusion combustion. The latter term is not strictly correct as premixed burning is also part of this burning phase, as discussed below.
The combustion paths of three types of mixtures—rich, stoichiometric, and lean—are presented in Figure 5. For the stoichiometric ratio, combustion is complete and its products are generally water and carbon dioxide. For rich mixtures, there are two possibilities. If the mixture remains rich, combustion will be incomplete, and it will manifest this fact by producing soot. The second option is for the rich mixture to find a leaner mixture or additional air with which to mix, forming an overall stoichiometric mixture to produce a complete burn.
The third type of mixture treated in Figure 5 starts lean. This mixture will also have two potential paths. If it mixes with a leaner mixture or just air, it will not burn effectively, and will produce unburned hydrocarbons most likely in the gaseous phase. However, if the lean mixture mixes with a rich mixture or more fuel and reaches a stoichiometric condition, it will then burn completely.
While the combustion process has been treated as consisting of three distinct phases, a fourth phase can also be defined that describes the activity in the final stages after the end of injection and prior to opening the exhaust valve. In this final phase any remaining fuel that had not been consumed will continue to burn, perhaps at a much lower rate as shown in Figure 2 (d → e) [fig]. In addition, some of the fuel that may have burned in the rate-controlled phase could have formed carbon, having some energy yet to release if oxidized. Conditions in this phase are still non-uniform from the temperature distribution as well as chemical make up points of view. Therefore, as long as motion still exists inside the cylinder, mixing will continue to occur and provide opportunities for the fuel as well as partially-burned products to completely burn.
Conceptual Diesel Combustion Model
Evolution of Diesel Combustion Model
The detailed understanding of the phases of the “conventional” diesel combustion process discussed above advanced significantly in the 1990s with the application of laser-sheet visualization techniques. Before the application of laser-sheet imaging, it was assumed that the quasi-steady portion of diesel combustion from shortly after ignition and up to the end of injection was closely related to steady-state reacting jets such as those found in furnaces and gas turbines. These early models of diesel combustion, which described the diesel combustion process as a fuel spray with a fuel rich liquid core and a fuel distribution that dropped off in a Gaussian-like manner with increasing radius, had three important characteristics:
• The liquid phase penetrated well out from the injector with fuel droplets being present up to or within the combustion zone.
• After the premixed burn, combustion occurred solely in a diffusion flame and was confined to the peripheral region of the jet.
• Soot occurred mainly in the shell-like region around the jet periphery.
Early laser-sheet imaging studies showed features that conflicted with these early conceptual models such as:
• Soot is distributed throughout the cross section of the downstream portion of the reacting diesel jet.
• There are no liquid fuel droplets in the reacting jet.
• Soot particles in the upstream portion of the jet are much smaller than those in the head vortex region.
Further imaging studies provided additional details and as a result, the conceptual model of the conventional diesel combustion process changed significantly, Figure 6. Specifically:
• The belief that ignition occurred at a few locations around the periphery turned out not to be the case. Ignition actually occurs at multiple points across the downstream regions of the jet.
• Rather than penetrating to the end of the reacting jet, the length of the liquid portion of the jet core is actually very short in normal burning.
• Even after the end of the premixed burn phase, fuel is partially consumed through rich premixed burning before the fuel reaches the diffusion flame where burning is completed.
While many aspects of the older models were correct and these models may have captured unique features of the combustion process in older technology diesel engines that were common at the time the models were developed, three fundamental aspects about combustion in modern diesel engines were not fully appreciated:
1. Injection velocities are very high in modern engines which lead to a flame standing off from the injector and allowing significant air entrainment upstream of the diffusion flame zone.
2. Mixing rates upstream of the tip of the liquid spray must be very high since the mixture downstream of this location is relatively uniform.
3. The instantaneous picture can be very different from the average picture. Early measurements were time averaged which showed Gaussian distributions. Actual gradients can be much steeper.
Conceptual Model of Conventional Diesel Combustion
The conceptual model of the premixed portion of the conventional diesel combustion is summarized in Figure 7. Since the description is for a specific diesel engine design operating at a moderate load condition, the quantitative information applies only to the case being considered. Values for other engines and operating conditions may be different.
1. Liquid jet emerges from injector tip. Air is entrained forming a mixture of air and fuel droplets.
2. Fuel vapor/air mixture starts to develop along sides of jet.
3. Liquid jet reaches maximum extent - vaporization by entrained hot air.
4. Chemiluminescence appears.
5. Start of rapid pressure rise from premixed burn.
6. Rich fuel vapor/air (φ=2-4) region develops beyond liquid jet. Chemiluminescence continues but little evidence of fuel breakdown.
7. Fuel breakdown and PAH formation in leading portion of jet. Rapid rise in pressure from premixed burn.
8. Small soot particles appear. Diffusion flame starts to develop at periphery.
9. Liquid jet becomes shorter.
10. Diffusion flame surrounds downstream portion of jet. Small soot particles fill leading portion of jet. Larger soot particles appear just inside periphery.
11. Premixed burn complete.
12. Soot concentration increases with greatest increase and largest particles near tip.
13. Next largest soot particles remain near periphery.
14. Mixing controlled burn dominates entirely. Head vortex developed. Soot concentration highest in head vortex.
15. No soot upstream of this location. Soot particle size largest in head vortex, medium along rest of periphery and smallest in center.
Figure 7. Conceptual Model of Conventional Diesel Combustion Process Through Premixed Burn Phase To Start Of Mixing Controlled Burn Phase
1° = 139 µs
As the liquid fuel jet emerges from the injector tip after the start-of-injection (ASI), it entrains air that has been heated by the compression process (~950 K) along its entire length (1.0-3.0°ASI). The entrained air heats the liquid fuel and causes it to start to evaporate. As this mixture of liquid fuel droplets continues to penetrate into the combustion chamber, it entrains more hot air and eventually a point is reached where the rate of energy supplied by the entrained air equals the energy required to evaporate all the fuel emerging from the injector. At this point, the liquid tip penetration stabilizes and reaches its maximum extent (Figure 13). This can happen before any significant heat release occurs and demonstrates that the liquid fuel penetration distance is not limited by combustion heating but by entrainment of compression heated air into the fuel jet.
Prior to the maximum liquid phase penetration, a vapor-fuel region forms along the sides of the liquid-fuel jet that is very thin or non-existent at the injector and that becomes progressively thicker downstream. The extent of the vapor phase fuel during this period is the same as that of the liquid phase (≤3.0°ASI). After the liquid phase reaches its maximum extent, the vapor phase continues to penetrate across the combustion chamber (>3.0°ASI) (Figure 17).
Prior to significant heat release (4.0-4.5°ASI), the fuel vapor is well mixed with entrained air to a fuel-air equivalence ratio of 2-4 just downstream of the liquid-fuel region. This rich but combustible mixture is present throughout the jet cross section and there are no regions of pure or almost pure fuel downstream of the liquid jet tip. This fairly uniform premixed region is separated from the surrounding air by a sharp well-defined boundary. A near-stoichiometric mixture occurs only in a very narrow region at the edges and contains only a very small portion of the premixed fuel.
Pre-ignition reactions (low temperature heat release) occur at multiple sites in the premixed vapor region and are uniformly distributed volumetrically. Chemiluminescence from these low temperature reactions first appears along the sides of the liquid jet well before ignition and before the vapor phase has penetrated beyond the liquid jet (3.5°ASI). They then appear in the premixed vapor region downstream of the liquid jet as it forms (4.5°ASI).
The breakdown of fuel throughout the premixed fuel-rich mixture about 70 µs (at 5.0°ASI) after the appearance of the chemiluminescence coincides with the initial rapid rise in the apparent heat release rate. This indicates that the initial premixed burn is fuel-rich. Soot precursor (PAH) formation starts almost simultaneously with the beginning of fuel break-down and soot formation starts ~140 µs later.
These first soot particles are very small and are detected throughout parts of the cross section of the premixed region. Then ~70 µs later (at 6.5°ASI), the soot particles are located throughout the entire cross-section of the downstream portion of the jet and larger soot particles are detected around the periphery of the premixed region. The appearance of the larger soot particles coincides with the formation of a diffusion flame at the same peripheral location—the soot volume concentration, however, remains similar to that at the center of the region where soot particles are smaller. The soot particles in the diffusion flame grow at a faster rate than those in the center. By the end of the premixed burn, the soot concentration in all regions of the reacting portion of the jet has increased substantially and the large-particle region at the periphery has become thicker due to transport. Large particles do however remain absent from the central part of the reacting region.
A diffusion flame starts to develop at the periphery of the fuel rich premixed portion of the reacting jet when temperatures and radical concentrations become sufficient (5.5-6.5°ASI). This flame spreads to engulf not only the downstream portion of the jet but to partially surround the region of fuel-air vapor just downstream of and around of the liquid jet. The diffusion flame is fed with partially reacted fuel, CO and H2 formed in the rich premixed burn zone on one side and surrounding air on the other. The diffusion burn commences shortly after ignition while premixed burning is still ongoing. Thus the premixed burn and the mixing limited diffusion burn do not proceed in a sequential manner. Rather, diffusion burning commences shortly after the premixed has started but long before it is over.
With the onset of the diffusion flame, local heating of the liquid fuel jet causes it to shorten slightly.
Throughout the remainder of the premixed burning phase, the jet continues to grow and soot concentration increases throughout the downstream portion of the jet with the greatest increase at the leading edge (8.0°ASI). The larger sized soot particles produced by the diffusion flame diffuse slightly inward but remain along the periphery. The soot particles at the head of the jet are even larger than those at the sides.
This soot distribution suggests a history to the soot formation where soot initially forms as small particles at the upstream location of the reaction zone with additional formation and growth continuing as the soot moves down the jet into the head vortex. This pattern persists until the end of fuel injection—about half-way through the mixing controlled burn phase.
As most of the fuel-rich premixed fuel-air mixture is consumed, the burning fuel jet transitions from the premixed burn phase dominated by fuel-rich premixed burning to the mixing controlled burn phase (9.0°ASI) where mixing controlled burning dominates (inset Figure 7).
Figure 8 represents the conceptual model of mixing-controlled conventional diesel combustion after the premixed phase terminates. Temporally it follows the last image of Figure 7 and represents the mixing-controlled combustion process until the end of injection.
It is important to note that towards the end of the premixed burn, soot appears rather abruptly in a thin zone at the transition from the fuel-vapor/air mixture region to the reaction region. At this location, a standing fuel-rich premixed flame has developed that partially consumes the fuel-air mixture produced by the evaporating liquid jet. The products of this rich standing premixed flame are mainly: H2, water, CO2, CO, methane, soot precursors and smaller fuel fragments. The energy release in this thin flame zone is mainly due to the formation of water and it produces temperatures around 1600 K.
The thin diffusion flame sheath surrounding the burning jet extends upstream of the standing premixed flame. The distance between the furthest upstream extent of the diffusion flame sheath and the fuel injector nozzle is referred to as the lift-off length. The lift-off length—a critical parameter that determines the air-fuel ratio of the standing rich premixed flame—can have a profound impact on emissions formation in the combustion process. This is discussed further in the following sections.
Upstream of the lift-off length, the temperature of the liquid fuel jet rises by mixing with hot air from its injection temperature of ~350 K to a temperature of ~650 K. When it enters the region downstream of the lift-off length surrounded by the diffusion flame, re-circulated products of partial combustion entrained into the jet further increase the temperature to about 825 K.
After the end of injection, most of the soot formed earlier in the combustion process has been oxidized by OH radical attack at the periphery of the jet. Upon closing of the injector, the velocity of the last amount of fuel exiting the injector can decrease significantly and it may have trouble reaching the leading edge of the fuel jet. It may not atomize well and mixing may be poor—leading to a significant amount of soot formation and soot particle growth. This soot formed along the jet axis is still present at the end of heat release is thought to contribute preferentially to tailpipe soot emissions.
Spray Formation
Fuel Atomization
Figure 9 shows a spray formed by injecting fuel from a single hole in stagnant air. Upon leaving the nozzle hole, the jet becomes completely turbulent a very short distance from the point of discharge and mixes with the surrounding air. This entrained air is carried away by the jet and increases the mass-flow in the x-direction and causes the jet to spread out in the y-direction. Two factors lead to a decrease in the jet velocity: the conservation of momentum when air is entrained into the jet and frictional drag of the liquid droplets. Figure 9 gives the velocity distribution at two cross sections. The fuel velocity is highest at the centerline and decreases to zero at the interface between the zone of disintegration (or the conical envelope of the spray) and ambient air.
Primary Atomization. Near the injector nozzle, the continuous liquid jet disintegrates into filaments and drops through interaction with the gas in the cylinder. This initial break-up of the continuous liquid jet is referred to as primary atomization.
In general, the atomization of a jet can be divided into different regimes depending on the jet velocity.
• Rayleigh Regime. In this low jet velocity regime, breakup is due to the unstable growth of surface waves caused by surface tension and results in drops larger than the jet diameter.
• First Wind Induced Breakup Regime. In this medium jet velocity regime, forces due to the relative motion of the jet and the surrounding air augment the surface tension force, and lead to drop sizes of the order of the jet diameter.
• Second Wind-Induced Breakup Regime. In this high jet velocity regime breakup is characterized by divergence of the jet spray after an intact or undisturbed length downstream of the nozzle. The unstable growth of short-wavelength waves induced by the relative motion between the liquid and surrounding air produces droplets whose average size is much less than the jet diameter.
• Atomization Regime. At very high jet velocity, breakup of the outer surface of the jet occurs at, or before, the nozzle exit plane. The average droplet diameter is much smaller than the nozzle diameter. Aerodynamic interactions at the liquid/gas interface appear to be one major component of the atomization mechanism in this regime.
Initial break-up in diesel fuel jets generally occurs in the atomization regime. The dominant mechanisms driving this process are not entirely clear. Interdependent phenomena such as turbulence and collapse of cavitating bubbles may initiate velocity fluctuations in the flow within the nozzle of the injector that destabilize the exiting liquid jet. The unsteadiness of the injection velocity and drop shedding also play an important role.
For most diesel fuel injection systems, jet atomization at the nozzle exit plane occurs when:
(6)√(ρa/ρf) < 18.3/√A
where A is a function of the length/diameter (Lo/Do) ratio of the nozzle:
(7)A = 3.0 + 0.28 (Lo/Do)
Secondary Break-Up. After the initial disintegration of the liquid jet and the initial formation of droplets, aerodynamically induced droplet breakup further reduces the size of the droplets as they penetrate into the surrounding air. This secondary breakup combined with evaporation ensures that droplets continue to decrease in size as they move along the x-axis (see Figure 9).
Secondary break-up is assumed to be controlled by the droplet Weber Number (We) which is defined as the ratio of the inertia forces to the surface tension forces:
(8)We = ρa Dd urel2 / σf
where
ρa - ambient density
Dd - droplet diameter
urel - relative velocity between droplet and the ambient gases
σf - surface tension of fuel.
This secondary break-up can be classified into a number of different modes depending on Weber number, as shown in Table 1.
In modern diesel engines, the droplet Weber number are typically in excess of 100 indicating that stripping and catastrophic regimes are the most important modes of secondary breakup. Secondary break-up starts at a finite distance from the injector, on the order of several mm, and then stops about 15-20 mm from the injector. Further reductions in droplet size downstream of this distance can be attributed almost entirely to evaporation.
Droplets experience considerable deformation during break-up and are not in fact spherical. Droplet distortion, as measured by the ratio of the long axis diameter of an elongated drop to a spherical drop, can be about 5 under typical modern diesel injection conditions. This increases the surface area of the drop by a factor of 7-10 and has a profound effect on fuel vaporization. This deformation ensures that the fuel vaporization rate equals the injection rate shortly after the start of injection.
Spray Characterization
Sprays are often characterized by their mean droplet size. One common indicator of mean droplet size is the Sauter mean diameter (SMD). It is defined as:
(9)SMD = (∫Dd3 dn) / (∫Dn2 dn)
where dn is the number of drops having diameter Dd. The SMD can be viewed as the diameter of a droplet that has the same surface-to-volume ratio as that of the entire spray. In DI diesel engines equipped with modern fuel injection systems that provide high injection pressures, SMD can be as low as 6 to 15 microns.
Having reviewed the importance of spray atomization and penetration for the proper operation of DI diesel engines, it is also necessary to review the importance of their influence on heat transfer characteristics. A greater number of small size droplets leads to a greater surface area that facilitates heat transfer from the hot compressed air to the small fuel volume contained in those droplets. It is also important to note that the injection process is not a steady stabilized process during which pressure, effective spray hole geometry, and injection rate are fixed. In fact, they do vary from the start to the end of injection. Therefore, one might expect the droplet size distribution to vary during the injection process. To further complicate the combustion process, temperature inside the combustion chamber is non-uniform and this non-uniformity varies throughout combustion. Therefore, mixing rates vary according to many parameters such as droplet size distribution, temperature distribution within the combustion chamber, spray penetration and atomization, fuel quality and its evaporation rates, as well as many other parameters. The ability to model or to simply understand diesel combustion depends to a great extent on the ability to define each of the parameters involved in this process including droplet size distribution. Figure 10 is an example of the effect of injection pressure on droplet size as influenced by nozzle hole geometry (a) and nozzle hole diameter (b).
The droplet size distribution given in Figure 11 is for a fuel spray produced from a nozzle hole at different times from the start of injection. At 0.7 ms injection duration, Figure 11 indicates that small droplets are most prevalent. At later times, larger droplet diameters had greater frequency than small droplets. In other words, as the injection continues, the smaller droplet population decreases as the larger droplet population increases, as a percent of the total number of droplets. Injection systems that sustain high injection pressure over most of the injection duration will most likely support a larger number of small droplets. This in turn improves the overall surface-to-volume ratio of the droplets, thus increasing their evaporation and mixing rates.
This is all about Engine combustion from my side which I got from internet and bit my study too.
