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In the case of internal mixture formation, the fuel is directly injected into the cylinder. Two concepts, that of early and of late injection, are possible. Early injection is the mostly used method for HCCI diesel applications and uses a long ignition delay along with low temperatures to homogenize the fuel mixture. A part or even the total amount of the fuel is injected noticeably before top dead center, see Fig. 6.31. In the case of diesel injection, the poor volatility of the fuel and the low gas density of the air inside the cylinder can result in considerable wall wetting. For these cases, new and highly flexible injections systems have to be designed, which have to be specially adapted to the variation of combustion chamber geometry as well as in-cylinder pressure and temperature during injection. Despite the problems of the in-cylinder injection, direct injection is expected to become the preferred method for HCCI engines in the future [87].
Some researchers have already spent significant effort on developing lowpenetration injectors and minimizing wall impingement [24, 17, 34, 39, 73]. A suitable injection system must have a high degree of flexibility in order to allow an adaptation of the injection strategy to the varying boundary conditions during injection. High-pressure injection in combination with a large number of small nozzle holes is generally used in order to increase spray disintegration and to include the complete cylinder charge in the mixture formation process while avoiding wall film formation. A further adaptation of the spray penetration can be realized by splitting the injection event into several pulses having different durations. Figure 6.31 shows an example of such a high-pressure pulsed injection strategy. The shorter the pulse duration, the less the momentum of the liquid, resulting in reduced penetration. The area below the curves represents the fuel mass belonging to each pulse. The low gas density at the beginning of injection requires short pulses with reduced injection velocities, and the time interval between the pulses is relatively large. As the piston moves up, density and temperature in the cylinder increase and penetration is reduced. The pulse durations can be prolonged, while the time intervals between subsequent pulses are decreased. At the end of the pulsed injection the distance between nozzle and piston reduces significantly, and the mass injected per pulse must be reduced again in order to prevent fuel deposition on the piston. A piezo common rail injection system capable of performing these fully variable pulsed injections has been developed by Meyer et al. [54].
Fig. 6.31. Pulsed injection strategy for early in-cylinder injection
Fig. 6.32. Adaptation of spray angle for early in-cylinder injection
In the case of early in-cylinder injection, the spray direction must be adapted as well. In comparison to the conventional diesel injection near TDC, the volume between the nozzle and the piston is significantly larger. In order to achieve an adequate mixing of fuel and gas and to prevent fuel deposition on the cold liner, the angle between the spray plumes must be decreased, see Fig. 6.32. In order to realize both combustion modes in an engine (dual-mode), the use of a variable nozzle concept, Sect. 6.2.4, would be highly favorable.
Figure 6.33 shows an example of a numerical investigation of the mixture formation in a 2.0 liter single-cylinder HCCI diesel engine with compression ratio 14:1 and zero swirl. A 13-hole common rail injector (hole diameter: 0.12 mm, one central hole, two rows of six holes with different spray directions) is used. Between 110° BTDC and 30° BTDC, nine pulses with a total mass of 70 mg, as shown in Fig. 6.31, are injected. The black dots represent the liquid droplets, the shaded background represents the ratio of air and evaporated fuel.
Fig. 6.33. Numerical investigation of mixture formation for early in-cylinder injection
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Liquid deposition on the liner can be avoided because of the short pulses and the narrow spray angles. However, these smaller angles result in fuel deposition on the piston at the end of the pulsed injection, and the use of a variable nozzle geometry would again be highly favorable. Because the temperature of the piston is significantly higher than the one of the liner, fuel deposition on the piston is less critical. Near TDC, a very lean but not completely homogeneous mixture can be achieved. Although the NOx emissions are not influenced very much by some degree of inhomogeneity, partial burning may occur in the very lean regions, which can result in increased HC and CO emissions.
A similar narrow spray angle and multi-hole HCCI injection system for early direct injection has been developed and tested by Gatellier et al. [24]. Multi-stage injection with up to eight pulses per cycle is used in order to reduce spray penetration. Engine and injection system are developed for dual-mode operation.
Fig. 6.34. Effects of EGR, retarded injection timing (IT) and increased swirl on exhaust emissions and thermal efficiency, Nissan MK-concept [43]
This means that the same injection system can be used for HCCI combustion at part load and for conventional diesel combustion at full load. Another dual-mode HCCI diesel engine has been developed by Duret et al. [17]. The authors also utilized a specially adapted narrow-angle direct injection system (NADI) with 10 holes and angles of 25° to 35° between needle axis and hole axis. An engine with a compression ratio of 14:1 and zero swirl was used. The injection was split into four pulses, starting at about 120°–110° BTDC and ending at 60°–35° BTDC in the case of HCCI combustion. Ultra low NOx and soot emissions at part load are reported, while the engine also retained the high fuel economy during conventional combustion at full load.
The second direct injection strategy is late in-cylinder injection, which was chosen for example by Kimura et al. [43] (Nissan MK-concept, Fig. 6.34) and by Shimazaki et al. [76]. In the case of the MK-concept, the late direct injection of diesel fuel starts at about 3° ATDC. At this time the density and temperature of the cylinder charge are high, mixture formation is enhanced, and a deposition of fuel on the walls can be avoided. However, due to high pressure and temperature, ignition delay is short. Only with the use of heavy cooled EGR and a reduced compression ratio (16:1) can ignition delay be prolonged such that injection can be finished prior to ignition. This is absolutely necessary in order to avoid a diffusionlimited combustion and the formation of soot and NOx. In addition to this, a high swirl ratio (SR: 5.0) is used to further enhance mixture formation. The onset of heat release is clearly after TDC and thus significantly later than with the conventional combustion, resulting in a reduction of peak pressure and combustion noise, but also in a decrease of thermal efficiency. Although there is a clear inhomogeneity of the mixture, very low NOx emissions can be achieved. Because it is impossible to achieve a further increase of ignition delay in order to inject larger fuel quantities, the MK-concept is also limited to low loads.
Control of Ignition Timing and Heat Release Rate
Besides the problem of homogeneous mixture formation, the control of ignition timing, which determines the main combustion phasing and thus has a strong influence on efficiency and operating range, is a serious challenge. Premature combustion can result in heavy knock-like combustion that destroys the engine.
A stable combustion can be realized at low and partial load for lean fuel-air ratios and/or large amounts of EGR. An increase in load towards stoichiometric values results in a significant increase in heat release rates and in knock-like combustion. Furthermore, the emission benefits vanish (e.g. [7, 85, 87]). Unlike in spark ignition or conventional diesel engines, a direct control of ignition timing via the spark or the start of injection in combination with the very short and well-known ignition delay is not possible. The start of combustion is significantly influenced by the low temperature chemistry, which depends on the complete timetemperature history of the charge.
Most of the applications with diesel fuel suffer from premature ignition, and cooled EGR or reduced compression ratios are used in order to increase ignition delay.
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Fig. 6.35. Methods for controlling HCCI combustion phasing [87]
In the case of gasoline applications, the reduced ignitability of gasoline and the generally lower compression ratios of these engines often require measures like intake air heating or non-cooled internal EGR to achieve a reliable ignition. The main reason for the difficulty in controlling the start of combustion exactly is the long time interval during which the low temperature reactions occur. A slight variation of the boundary conditions can easily result in significant variations of the main combustion phasing. HCCI combustion phasing is affected by
ξ the auto-ignition properties of the fuel, ξ the fuel-air ratio,
ξ the volatility of the fuel,
ξ the EGR rate, the temperature and the reactivity of the recycled gas, ξ the mixture homogeneity,
ξ the compression ratio of the engine, ξ the intake temperature, and
ξ the heat transfer to the engine.
In order to control combustion phasing, two main groups of approaches can be distinguished [87], see Fig. 6.35. The first group are methods, the purpose of which is to alter the time-temperature history of the mixture. It includes fuel injection timing, variation of intake air temperature, variation of compression ratio (CR) and variable valve timing. The second group attempts to control the reactivity of the charge by varying the properties of the fuel, the fuel-air ratio or the amount of oxygen by EGR. A detailed description of the influence of each method on the HCCI combustion process is given in the following section.
6.4.5 Influence Parameters and Control of HCCI Combustion
The key challenge for operating an engine in HCCI auto-ignition mode is the control of the combustion process without direct access to the reactions. The start of chemical reactions is initiated by the thermodynamic conditions and the chemical composition of the cylinder charge. In the following, the important engine parameters influencing the start of combustion and the subsequent heat release are discussed in detail.
Compression Ratio
Being directly related to charge temperature, the compression ratio is important in determining the rate of heat release in HCCI combustion processes. An increase of compression ratio raises the end-of-compression temperature and the rate of low temperature reactions. This advances the overall ignition process and allows inlet charges of lower temperature to be successfully combusted (CAI process). In the case of HCCI diesel engines the conventional compression ratios usually must be reduced in order to delay the start of combustion and to prevent excessive heat release rates. Nevertheless, the reduction of compression ratios should be moderate, because a reduction of compression ratio decreases thermal efficiency.
Figure 6.36 from Ryan’s study of diesel fuelled HCCI engines [71] shows predicted values of start of combustion versus compression ratio. With a constant EGR rate of 40% ignition advances about 200 degrees if the compression ratio is changed from 8:1 to 12:1. Figure 6.37 shows similar results from Velji et al. [94] Excessive compression ratios result in knock-like combustion with high temperatures and increased NOx emissions.
A variable compression ratio by variable valve timing is an effective but expensive method for solving this conflict. This technology can be used for dual-mode engines to realize the high compression ratios needed for high thermal efficiency in the case of conventional combustion at full load and to enable lower compression ratios in the case of HCCI combustion.
Fig. 6.36. Start of combustion versus compression ratio [71]
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Fig. 6.37. Effect compression ratio on HCCI combustion [94]
Variable Valve Timing
In contrast to HCCI diesel applications, in which EGR-cooling is needed in order to suppress a premature ignition start, lean homogeneous gasoline-air mixtures need additional heating in order to auto-ignite. The required high temperatures and the charge dilution for decelerating the heat release are usually obtained by the application of non-cooled internal EGR, which can be provided efficiently only by variable valve trains. The residual gas can be supplied either by remaining in the combustion chamber due to early exhaust valve closing, or by redrawing from the exhaust system during the induction stroke, Fig. 6.38.
Although simple systems with cam phasers and valve lift shift tappets can fulfill the steady-state HCCI requirements, the dynamic behavior is limited. For fully transient operation a cycle-to-cycle and cylinder individual control of charge composition is advantageous. This is possible with a fully variable electro-hydraulic
Fig. 6.38. Internal EGR: residual gas supply for HCCI operation [23]
valve actuation system [17, 23, 6]. These systems also enable one to change the effective compression ratio and to modify the intake flow (swirl ratio), which is also beneficial for conventional operation (dual-mode engine).
Exhaust Gas Recirculation (EGR)
Exhaust gas recirculation is a well-known method for reducing combustion temperatures and NOx-emissions in conventional diesel engines. A part of the original cylinder charge is replaced by recycled burnt gases. In general, two different EGR-concepts are used today, the external EGR and the internal EGR. In the case of external EGR, a portion of the exhaust gases in the manifold is branched off, cooled if necessary, and then mixed with the fresh air in the suction part. The cooling offers an additional possibility to reduce combustion temperatures. In the case of internal EGR it depends on the inlet and outlet valve timing whether a portion of the hot exhaust gases of the previous cycle stays in the cylinder (e.g. [49, 102]) or is sucked back from the manifold (e.g. [6, 23, 49, 102]). In the case of HCCI combustion, EGR has four important effects:
Heat capacity effect: During compression and combustion, the inert burnt gases must be heated up together with the rest of the in-cylinder charge. Because the total heat capacity of the charge is higher with burnt gases due to the higher specific heat capacity values of carbon dioxide (CO2) and water vapor (H2O), lower end- of-compression and combustion temperatures are achieved, and heat release rates as well as maximum pressure rise (dp/dt)max are reduced [48, 104]. The heat capacity effect extends the combustion duration if large amounts of EGR are used.
Charge heating effect: if hot burnt gases are mixed with cooler inlet air, the temperature of the inlet charge increases. The heating effect is important for noncooled EGR applications and is mainly responsible for advanced auto-ignition timing. It also increases the heat release rate and the value of maximum pressure rise (dp/dt)max, and shortens the combustion duration [104].
Dilution effect: The introduction of burnt gases into the cylinder replaces a part of the inlet air and causes a reduction of the oxygen concentration. This effect does not affect auto-ignition timing in the case of CAI combustion, but it extends combustion duration and also slows down the heat release rate if large amounts of EGR are used [104, 31]. Experiments of Tsurushima et al. [92] with diesel-like fuels (two-stage ignition) also show a decrease of oxidation rates, but the start of both, LTO and HTO, is delayed, and the time interval between both reactions is expanded.
Chemical effect: Finally, active combustion products present in the burnt gases can participate in the chemical reactions of the subsequent combustion cycle. External and cooled EGR usually provides chemically inert gases and thus does not contribute to this effect. However, in the case of internal EGR, the recycled burnt gases may contain short-lived chemically active components that result in advanced auto-ignition [38, 102, 48].
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Fig. 6.39. Successful HCCI operational range and IMEP map. Parameter: IMEP, compression ratio 18:1, intake temperature 30°C, n-heptane [65]
The influence of the EGR-effects mentioned above depends on the HCCI-sub- concept that is applied. In the case of HCCI diesel applications, external EGR with additional cooling is customarily used in order to prolong ignition delay, while in the case of CAI combustion (gasoline-HCCI) internal non-cooled EGR is mainly used in order to increase the temperature to the required auto-ignition level. A very extensive study of the HCCI-process on a four-stroke diesel engine (Η = 18:1, port injection system, external EGR, fuel: n-heptane) has been carried out by Peng et al. [65]. Similar investigations for a gasoline CAI process have been reported by Oakley et al. [61]. The investigations of Peng et al. give a good overview about the general effect of EGR on knock limit, engine load, combustion stability and emissions.
Figure 6.39 shows the successful HCCI operating region and the values of the indicated mean effective pressure (IMEP) as a function of EGR rate and the overall air-fuel ratio of the cylinder charge. At constant EGR-rate, the HCCI combustion is limited by its low load limit (lean mixture) and by the knocking combustion at higher loads (richer mixture). Starting from low IMEP and increasing load, the amount of fuel injected per cycle must be increased, resulting in lower air-fuel ratios (decreasing lambda values) and richer mixtures inside the cylinder. In the case of knocking combustion, too much heat is released during the very fast combustion which results in excessive peak pressures and can damage the engine. Near this limit, high local temperatures result in an increase of NOx-emissions [23, 24, 31]. An increase of EGR leads to a decrease of combustion temperature, later combustion timing and a reduced heat release rate.
Fig. 6.40. Effect of compression ratio on the operational range of HCCI diesel combustion [17]
This allows for an increase in the amount of fuel injected per cycle and thus in load and for a shift in the knock limit towards smaller lambda values. On the other hand, EGR also reduces the low load limit in terms of lambda: the more recycled burnt gases are present, the lower the combustion temperature. These low temperatures can result in partial burning, which again results in increased HC and CO emissions and in a reduction of efficiency. The third limit of the HCCIoperation region occurs at EGR-rates of about 70%. Reaction rates and ignition timing are so much reduced and retarded that misfiring occurs and HC-emissions increase again.
Similar investigations have been performed by Duret et al. [17]. Figure 6.40 shows the effects of a variation of compression ratio. The operational range of HCCI diesel combustion increases with increasing compression ratio and the misfire limit is shifted to higher EGR ratios. However, the maximum IMEP value is reduced due to earlier ignition, faster combustion and thus earlier onset of knocking combustion at leaner mixtures.
Inlet Air Temperature
Modulating the intake air temperature in order to control the start of ignition is the most popular method in laboratory experiments. Higher intake temperatures advance the start of combustion and vice-versa. Figure 6.41 shows results from Velji et al. [94] for a blend of 80% methane and 20% diesel and different air temperatures. Besides a large loss in volumetric efficiency in the case of high intake air temperatures, the range of crank angle, over which the ignition timing can be influenced, is quite limited. In the case of non-cooled EGR, the variation of the intake temperature is usually one of the multiple effects of a change of the EGRrate.
