A computational study of direct injection gasoline HCCI engine with secondary injection

Fuel 85 (2006) 1831–1841 www.fuelfirst.com

A computational study of direct injection gasoline HCCI engine with secondary injection Zhi Wang *, Shi-Jin Shuai, Jian-Xin Wang, Guo-Hong Tian State Key Laboratory of Automotive Safety and Energy, Department of Automative Engineering, Tsinghua University, Beijing 100084, China Received 15 August 2005; received in revised form 20 February 2006; accepted 21 February 2006 Available online 23 March 2006

Abstract The detailed intake, spray, combustion and pollution formation processes of compression ignition engine with high-octane fuel are studied by coupling multi-dimensional computational fluid dynamic (CFD) code with detailed chemical kinetics. An extended hydrocarbon oxidation reaction mechanism used for high-octane fuel was constructed and a modeling strategy of 3D-CFD/chemistry coupling for engine simulation is introduced to meet the requirements of execution time acceptable to simulate the whole engine physicochemical process including intake, compression, spray and combustion process. The improved 3D CFD/chemistry model was validated using the experimental data from HCCI engine with direct injection. Then, the CFD/chemistry model has been employed to simulate the intake, spray, combustion and pollution formation process of gasoline direct injection HCCI engine with two-stage injection strategy. The models account for intake flow structure, spray atomization, droplet evaporation and gas phase chemistry in complex multi-dimensional geometries. The calculated results show that the periphery of fuel-rich zone formed by the second injection ignited first, then the fuel-rich zone ignited and worked as an initiation to ignite the surrounding lean mixture zone formed by the first injection. The two-zone HCCI leads to sequential combustion, this makes ignition timing and combustion rate controllable. In addition, HCCI load range can be extended. However, the periphery of fuel-rich zone leads to fierce burning, which results in slightly high NOx emissions. q 2006 Elsevier Ltd. All rights reserved. Keywords: HCCI; Direct injection; CFD

1. Introduction Homogeneous charge compression ignition (HCCI) has advantages in high thermal efficiency and low emissions and possibly become a promising combustion method in internal combustion engines. Recent researches have shown that HCCI engine fueled with high-octane number (ON) fuel has more advantages in fuel economy than that of low-octane fuel [1–8]. For the HCCI combustion with low-octane fuel, like diesel, ignition occurs in two stages (low-temperature heat release and high-temperature heat release). Therefore, ignition timing and combustion phase are difficult to optimize. Moreover, the lowtemperature heat release begins at a temperature of about 800 K. This ignition temperature limits the compression ratio to 13:1 and the efficiency to less than a conventional diesel engine. Generally, diesel HCCI engine only has advantage in NOx and PM emission, but lower thermal efficiency than * Corresponding author. Tel.: C86 10 62794876; fax: C86 10 62772515. E-mail address: [email protected] (Z. Wang).

0016-2361/$ - see front matter q 2006 Elsevier Ltd. All rights reserved. doi:10.1016/j.fuel.2006.02.013

conventional diesel engine. In contrast, the high-octane fuel allows compression to higher temperatures with ignition occurring in a single stage at above 1000 K. This temperature permits higher compression ratios, which lead to higher efficiencies. HCCI engine tests have shown that using highoctane fuel, such as iso-octane or gasoline, and diesel-like compression ratios, HCCI engines can achieve diesel-like efficiencies and low NO emissions [2–4] with stable ignition timing. The researches of HCCI engine with high-octane fuel existed can be fall into two categories according to compression ratio (CR): low-CR engine with slight-diluted mixture and high-CR engine with ultra-diluted mixture. The low-CR HCCI engines have advantages in switching to conventional SI mode at high load [5]. The high-CR HCCI engine with ultra-diluted mixture can realize ultra-low NOx emission (1 ppm) and higher thermal efficiency (43%) at fixed operation point [4]. Therefore, high-octane fuel is suitable for stationary HCCI application. For high-CR engine fueled with high-octane fuel, the fuel/air equivalence ratios (f) vary from 0.1 to 0.5 and the range of CR from 16 to 21 of conventional diesel engine. For instance, GM [6], Sandia [7], and KTH [8],

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investigated high-octane fuel HCCI combustion on direct injection engine, which was modified from heavy duty diesel engine. Although many researches on HCCI combustion with highoctane fuel have been carried out, some of the problems in ignition and combustion control are still to be solved. Unlike conventional combustion mode, HCCI combustion is not controlled, either by injection timings as in a CI engine, or by spark timings as in an SI engine. It is now generally agreed that HCCI combustion is a chemical-controlled combustion dominated by local chemical-kinetic reaction rates [9]. This view has been supported by spectroscopic data indicating that the order of radical formation in HCCI combustion corresponds to self-ignition rather than flame propagation [10,11]. Therefore, detailed chemical kinetics must be taken into account for HCCI modeling. Recent analytical developments also support this view and an analysis method based on this premise has had considerable success in predicting HCCI combustion and emissions [12]. However, even if a perfect homogeneous mixture exists at the time of combustion, turbulence has an effect to alter the temperature distribution and the boundary layer thickness within the cylinder. A little temperature differences in-cylinder have a considerable effect on combustion due to the sensitivity of chemical kinetics to temperature. As a result, heat and mass transfer are important in forming the condition of the charge prior to ignition. In this case, 3D-CFD with detailed chemistry is a suitable tool to capture relevant physicochemical processes and this powerful analytical tool constitutes a great advantage for HCCI engine research. This paper presents a ‘two-zone HCCI’ concept using secondary direct injection in middle of compression stroke for compression ignition and tries to solve the problems of ignition control. ‘Two-zone HCCI’ in this paper is defined as premixed ultra-lean compression ignition combining with premixed ultra-rich compression ignition. The objective of the work is to reveal ignition and combustion characteristics in direct injection compression Ignition engine with high-octane fuel, to study the ignition control mechanism in ‘two-zone HCCI’ combustion via analysis the auto-ignition spots distribution in combustion chamber. In order to resolve the problem between computational capabilities and simulation reliability, a modeling strategy for 3D-CFD with detailed chemistry has been introduced in this paper.

2. 3D model of HCCI engine 2.1. Chemically reactive flows with sprays An integrated numerical model for HCCI engine combustion computations has been developed in this study. HCCI combustion model with detailed chemistry are adopted in order to capture the combustion characteristics of HCCI engine. Three-dimensional model, in which calculations of complex chemical-kinetics/turbulent interaction are carried out using 3D-CFD code FIRE coupled with detailed chemistry, is used for providing an insight into the details of the physicochemical

processes during the intake, compression, spray, ignitioncombustion-emission. Based on the fundamental conservation principles, the CFD code solves the averaged transport equations of total mixture mass, momentum and enthalpy. Turbulence effects are accounted for by adoption of the standard two equations k–3 model to solve the turbulent flow in complex geometry. The governing equations in its non-orthogonal form are solved in a contracting/expanding coordinate-frame, so as to enable their application to body-fitted computational grids with moving boundaries. The partial differential/transport equations are discreted on the basis of a finite-volume method. Spray simulations involve two-phase flow phenomena and as such require the numerical solution of conservation equations for the gas and the liquid phase simultaneously. The spray module is based on a statistical method referred to as the discrete droplet method (DDM) [13]. The evaporation model is based on the characteristics of one-component fuels like n-heptane or iso-octane. The heat and mass transfer processes are described by an evaporation model originally derived by Dukowicz [14]. The effects of turbulence on the spray particles are modeled by Gosman and Loannidis [15] through adding a fluctuating velocity to the mean gas velocity. CFD code FIRE offers a generic species transport to implement a detailed kinetic models. The general species transport model provides the necessary transport equations for gas phase chemical species in the computational domain. An arbitrary number of chemical species with an arbitrary set of properties can be defined, which is available for the calculation of the physical properties of the chemical species and the mixture of gases. In General species transport equations, the mass source can be calculated by gas phase reactions by regarding each cell as single zone reactor at every time step. 2.2. Chemical reaction mechanism Chemical kinetics reaction mechanisms for primary, alternative fuels such as hydrogen, methane, ethanol, dimethyl ether, n-heptane and iso-octane are well developed and have already proven valuable for HCCI investigations [16–20]. The reaction mechanism adopted in this paper was obtained by adding NOx mechanism into iso-octane oxidation mechanism. The iso-octane reaction mechanism developed by Dr Valeri Golovichev at Chalmers University was used to simulate the ignition and combustion chemistry of high-octane gasoline. The oxidation reaction path depends sensitively on temperature-low-temperature regime, Intermediate temperature regime, and high temperature regime. This mechanism is able to accurately predict the ignition delays and burn rate for iso-octane mixtures in wide ranges of pressures [16]. For HCCI engine with gasoline direct injection, the regions of high-temperature lean-fuel, low-temperature lean-fuel and low-temperature rich-fuel are existed simultaneously in combustion chamber. Therefore, nitric oxide is formed by the following three chemical routes-Zeldovich mechanism, N2O-intermediate mechanism, and prompt mechanism.

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Thermal NO: (extended Zeldovich mechanism) O C N2 Z NO C N

(R1)

N C O2 Z NO C O

(R2)

N C OH Z NO C H

(R3)

This mechanism becomes important in high temperature (TO1800 K) lean-fuel conditions. N2O-intermediate mechanism: O C N2 C ðMÞ Z N2 O C ðMÞ

(R4)

H C N2 O Z NO C NH

(R5)

O C N2 O Z NO C NO

(R6)

This mechanism is important in premix, lean-fuel (f!0.8), low-temperature conditions [21]. This mechanism becomes important in NO control strategies that involve lean premixed combustion. Prompt mechanism: CH C N2 Z HCN C N

(R7)

C C N2 Z CN C N

(R8)

HCN C O Z NCO C H

(R9)

NCO C H Z NH C CO

(R10)

NH C H Z N C H2

(R11)

N C OH Z NO C H

(R12)

NO C HO2 Z NO2 C OH

(R13)

NO2 C H Z NO C OH

(R14)

NO2 C O Z NO C O2

(R15)

This mechanism is intimately linked to the combustion chemistry of hydrocarbons and becomes important in rich-fuel, low-temperature regions, especially for droplet evaporation and burning. In consideration of combustion in GDI chamber, the NOx sub-mechanism has been constructed by 14 reactions mentioned above from the three routines. The corresponding kinetic constants in NOx sub-mechanism are taken from GRI3.0 [17]. The constructed reaction mechanism with NOx formation for high-octane fuel oxidation consists of 89 species and 413 reactions. 3. Modeling strategy 3.1. Simulation approaches Since some investigations have shown that the simultaneous calculation of a detailed fluid mechanics code with a detailed

Fig. 1. Modeling partition of work process of HCCI engine.

chemical kinetics code is well beyond our current computational capabilities [12]. Instead of directly linking fluid mechanics with chemistry, Salvador M. Aceves et. al. [12] developed a segregated, sequential multi-zone methodology. In this method, KIVA-3V was used to obtain temperature profiles, which were used as inputs to the following 10-zone chemical kinetics model. Kong et al. (2001, 2002) [22,23] developed KIVA/CHEMKIN model to simulate the HCCI engine combustion, in which computations started from compression stroke to expansion stroke using a 2D mesh due to symmetry of their combustion system. In order to keep the requirements of execution time acceptable to simulate the whole HCCI engine 3D working process including intake, compress, spray and combustion process, a new modeling strategy based on previous work was adopted in this paper. There are three simulation approaches in this modeling strategy as follows. 3.1.1. Span partition The working process of HCCI engine can be divided into three stages shown in Fig. 1. The chemical reactions only occur during the combustion duration. Generally, the combustion duration of HCCI engine with high-octane fuel is very short (!20 8CA). For iso-octane/air rapid compression, the chemical reactions donot occur until temperature rises to 786 K [24]. The start of reaction is defined as the point when temperature of anyone cell reaches 786 K. The approach of span partition can be explained as follows. Firstly, the CFD model with the real geometry of the engine is used to simulate the fluid dynamic processes correctly and the results of the CFD calculations can be used as initial values for the next stage. Secondly, the fluid mechanism couples with chemistry during the combustion period. After that, the exhaust process is simulated by CFD only. Using this approach, the calculation time can be reduced to less than 1/10. 3.1.2. Mesh rezone Two meshes of different quantity at the same crank angle are prepared to import into one project. At specified crank angle, mesh rezoning maps flow variable data that is generated

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during the previous calculation from the first mesh to the second for next calculation. The reasons for mesh rezoning are unacceptable distortion of cells caused by mesh movement or the use of different grid resolutions to speed up the simulation. Because the combustion occurs at very near TDC and the volume at TDC is a small portion of the volume at BDC. After rezone, the cell number of TDC mesh is no more than 1/10 than BDC mesh. Meanwhile, the running time is proportional to square of the number of cells when CFD couple with chemistry calculation. So this approach reduces calculation time to 1/100. 3.1.3. Limited mechanism A detailed mechanism with limited size, which consists of 89 species, is adopted in this paper. Compared with the full isooctane mechanism (LLNL mechanism, nearly 1000 species), the running time reduces to 1/100 because it is proportional to square of the number of species. 3.2. Calculation flow The most significant step in the calculation was to combine the CHEMKIN developed by Sandia National Laboratory [25] and the AVL FIRE 3D-CFD code. The 3D-CFD code can calculate the temperature and concentration distribution within the cylinder. While the chemical kinetics terms in the conservation equations can calculate chemical heat release as a function of pressure, temperature and composition in each cell. In fact, each cell is a well-stirred reactor. Fig. 2 shows the calculation flow diagram of linking CFD and CHEMKIN. For spray and combustion in HCCI engine, the system of chemical species must be predefined. In FIRE the calculation of the source terms is fully open for user-access via compiling user routines in FORTRAN format. To account for the effects of both chemistry and flow turbulence, a chemistry solver is available in FIRE by linking CHEMKIN library such that homogeneous reaction can be solved. At the beginning of each time step, a single zone reactor model is called for each computational cell so that the detailed reaction mechanism is taken into account. A formatted ASCII representation of a chemical reaction mechanism can be interpreted to the binary file required by CFD initialization. This file contains information that contains all required data about the elements, species, and reactions in the user’s mechanism. The corresponding physical properties (density, specific heat, viscosity, thermal conductivity, diffusion coefficient) of each species and of the gas mixture are calculated based on parameters extracted from CHEMKIN database. Chemical reactions are only activated during the combustion duration defined in the previous sections. The main aim of the modeling strategy is to carry out a completed, detailed chemistry calculation over the combustion duration. In other stages, CFD calculations are performed to provide the detail temperature and concentration distribution in flow field.

Fig. 2. Calculation flow diagram.

4. Engine setup and computational details 4.1. Engine specifications The investigation was carried out on a test engine modified from a diesel engine. The engine specifications are listed in Table 1 and fuel properties of test fuel are listed in Table 2. 4.2. Combustion system The combustion system is shown in Fig. 3. The additional spark is used in spark ignition (SI) combustion mode at high load. In this system, high intake swirl can be kept in the toroidal Table 1 Engine specifications Type

2-Cylinder in-line, DI

Number of valves Bore Stroke Connecting rod Compression ratio Intake valve open Intake valve close Injector Manufacturer Flow rate of injector Head bottom Intake port Combustion chamber

2 Per cylinder 95 mm 115 mm 210 mm 17.8:1 K10 CAD 210 CAD High pressure swirl injector Mitsubishi 15.7 mm3/ms Flat Helical u

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Table 2 Fuel properties of test fuels Research octane number Motor octane number Antiknock index (RCM)/2 Density, 20 8C (g/cm3) Vapour pressure (kPa) Olefins % vol Aromatics % vol Oxygen % mass Sulfur % mass Carbon % mass Hydrogen % mass C/H A/F stoichiometric Lower heating value Distillation (8C) 0% 10% 50% 90%

94.4 83.0 88.7 0.7487 55.4 23.3 34.9 0.21 0.01910 86.55 13.49 6.42 14.6 44.0 34.0 54.5 103.7 163.5

combustion chamber promoting dispersion of the fuel injected in the compression stroke. Gasoline direct injection timing can vary in a wide range, from intake stroke to compression stroke. Early fuel injection is used for HCCI combustion mode due to sufficient mixing time available to obtain a homogeneous charge. The late injection in compression stroke can form a stratified mixture to obtained stratified charge compression ignition (SCCI) or stratified charge spark ignition (SCSI). 4.3. Computational grids The CAD model of the intake port was generated in Pro/Engineer. The geometry of flow field was obtained via CAD interface, STL, with the accuracy of 0.021 mm. The computational meshes of the HCCI engine are shown in Fig. 4. There are 117,842 cells with intake port at minimum valve lift and 9360 cells at combustion TDC. In engine operation, valves and the piston move, so the mesh should also move according to the real engine in order to simulate the change of valve and piston positions with crank angle.

Fig. 3. Combustion system layout.

Fig. 4. Computational grids.

In order to simulate the valve and piston position accurately at any crank angle, 13 groups of mesh were created from intake valve open (IVO) to exhaust valve open (EVO) at critical crank angles in this study. There are 12 rezones occur at 10, 20, 30, 50, 100, 180, 190, 210, 320, 340, 350, 380 and 400 8CA, respectively (Intake top dead center is defined as 0 8CA in this paper) for the sake of high mesh quality and saving calculation time. According to the experience obtained from rapid compression machine calculation [24], meshes with no more than 10,000 cells are acceptable for CFD/CHEMKIN calculation. 4.4. Boundary conditions The measured valve lift is given in Fig. 5 and temperature definition suggested by AVL [27] for each surfaces are listed in Table 3. Exhaust process is not the concern of this paper, so exhaust valve and port were not included in the computational mesh. Calculations begin at IVO and end at EVO. In the time steps within, which only flow is calculated, a time step of 0.2 8CA is used, while in the spray duration, a time step of 0.1 8CA is applied; while in the combustion duration, a time step of 0.5 8CA is applied. Each calculation case took about 80 h for

Fig. 5. Measured valve lift profile.

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Table 3 Temperature definition for surfaces Cylinder linear Cylinder head Piston top-land Piston bowl Intake valve Intake port

423 K 513 K 473 K 473 K 373 K 373 K

intake-spray-compression process, 60 h for combustion process on a PC workstation (CPU: P4 2.8 GHz). 5. Results and discussion Fig. 6 shows the temperature field at the end of the fuel spray in intake stroke. For spray/wall impingement, a walljet model [28] was adopted. It does not take into account the wallfilm physics. This allows spray/wall impingement calculations to be performed without using the wallfilm module, which is sufficient in a variety of practical applications where the wallfilm physics do not play an essential role within the wall interaction process. It can be seen that, the early fuel injection in intake stroke can utilize the heat of residual gas promote fuel vaporization. Moreover, there will be significantly long time available for mixture preparation. Therefore, it was found that the use of early injection timing achieves a homogeneous mixture near combustion TDC. Fig. 7a shows the streamline field at maximum valve lift. It can be seen that, a majority of gas feed into cylinder via helical segment of the intake port. The remaining feed directly into cylinder at a tangent of cylinder wall. Fig. 7b shows the velocity distribution in cylinder at intake valve closure (IVC). From Fig. 7b, it can be seen that the gas rotate around the axis of combustion chamber at IVC. The result can be attributed to the intake swirl due to the helical intake port and the toroidal combustion chamber. This is why some investigations approximately specify a swirl ratio at IVC for diesel engine combustion simulation. The calculated cylinder pressure histories using 3DCFD/Chemkin model and 0D-Senkin model [24] are compared with the measured pressure in Fig. 8. It can be seen that the predicted ignition timings agree well with different model since

Fig. 6. Intake flow and spray.

Fig. 7. Flow structure in cylinder.

homogeneous charge was obtained due to the early injection in intake stork. The maximum rate of pressure rise and the peak pressure predicated by 3D-CFD/Chemkin model agree well with measured data than that by 0D-Senkin model. However, the measured pressure after 350 8CA is lower than

Fig. 8. Comparison of calculated and measured pressure curves.

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Fig. 9. Main species concentration in-cylinder.

the calculated pressure. This result can be attributed to the fact that leakage becomes significant when piston is near TDC. Fig. 9 shows the variation of the main combustion species with crank angle. As the mixture is compressed, the global temperature rises gradually. The species concentration incylinder start to vary after 318 8CA. Afterward, CO and formaldehyde begins to form gradually. As compression continues, iso-octane begins to decompose rapidly at 350 8CA. CO, CH2O, H2O2 form gradually. The reaction (H 2O 2CMZOHCOHCM), which produces two OH, becomes more and more important. The resulting OH radicals rapidly consume the fuel and initiate a branched thermal explosion and a following rapid increase in temperature. As a result, consumption of OH and fuel results in ignition. Since the ‘ignition’ point, are not well defined with HCCI, the start of combustion (SOC), can be taken to be the crank angle of sharply decreased H2O2, that is the crank angle of sharply increased OH. In this case, OH sharply increases at 358 8CA (see Fig. 9) and the corresponding temperature in-cylinder is 1150 K, this critical temperature is regarded as a criteria used to determine the ignition in this paper. The temperature rises to 1697 K rapidly at 360 8CA (TDC) due to the high exothermic reactions. The maximum rate of temperature rise is 300 K/8CA. Meanwhile, NO begins to form near TDC. Table 4 lists the comparison with different loads (equivalence ratios) when engine operates at 1800 rpm under the condition of early injection (100 oCA ATDC). It indicates that ignition advances and peak pressure increases as the increase of fuel/air equivalence ratio (f). The calculated the results agree well with experimental results at higher load. Fig. 10 shows the comparison of emissions data at the corresponding three cases. Both experimental results and simulated results show same

1837

tendency in NOx, CO and CO2 emissions. It indicates that NOx emissions increase as the increase of f due to the increased temperature in cylinder. It also can be seen that CO emissions decrease as the increase of equivalence ratio (f), particularly at lower loads the peak burned gas temperatures (1194 K, fZ0.18) are too low to complete the CO–CO2. This results in a significant increase in CO emissions relative to conventional SI engines. The calculated CO emissions are always less than measured data. This can be explained by the fact that the crevices of piston ring were not taken into consideration in the model. As with homogeneous charge combustion systems, a significant portion of the in-cylinder fuel is stored in crevices during the compression stroke and escapes combustion. Unlike traditional SI engines, the burned gas temperatures are too low to consume much of this unburned fuel when it re-enters the cylinder during the expansion stroke [29]. The calculated NOx emissions are always less than measured data. This can be explained by the NOx mechanism. Since the peak temperature of lean HCCI case is less than 1800 K, NOx is very low. NOx does not mainly form via high temperature mechanism. Some of NOx possibly comes from Nitrogen in real fuel, which was not been taken into consideration. With the fuel/air equivalence ratio increase from 0.18 to 0.28, the peak temperature in cylinder increase from 1330 to 1835 K, the NOx calculated discrepancies decrease gradually. Another reason for that is more micro-inhomogeities exists in cylinder of experimental GDI engine than that of simulated case. That is, the 3D grid is not fine enough to distinguish micro-scale mixing. Since NOx emissions are sensitive to the temperature distribution, the calculated NOx emissions are less than the measured data. 6. Parametric study of injection strategy Two-stage injection strategy is more flexible than singlestage injection, and controls the distribution of mixture concentration in the cylinder more effectively. According to the injection strategy (single or two-stage, two cases selected as listed in Table 5). SOI1 and SOI2 stand for start of first injection and start of second injection, respectively. The corresponding mass of injection is gb1 and gb2. The trapped residual gas ratio and temperature are 5.6% and 800 K, respectively, in each case. Fig. 11 shows the mixture formation under the conditions of two-stage injection. Fig. 11a shows the distribution of equivalence ratio at crank angle of maximum intake valve

Table 4 Comparison of calculated data and experimental data at three cases f

0.28 0.24 0.18

SOC (8CA)

IMEP (MPa)

Peak pressure (MPa)

Calc.

Exp.

Error ( 8CA)

Calc.

Exp.

Error (%)

Calc.

Exp.

Error (%)

358 359 362

357 357 359

1 2 3

0.221 0.202 0.099

0.222 0.210 0.063

0.5 4 36

6.34 6.16 4.91

6.27 6.00 4.67

1 3 5

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Fig. 10. Comparison of emissions at three cases.

lift. It can be seen that, in the HCCI combustion system, the first stage fuel injection can utilize the high intake swirl to promote fuel mixing and vaporization. Fig. 11b shows the distribution of relative fuel/air ratio at the end of second fuel injection. It can be seen that, the second stage injection in compression stroke can form a weak stratified charge based on the first injection to realize a robust stratified charge compression ignition (SCCI) combustion. Fig.12 shows the comparison of temperature and concentration distribution in cylinder at crank angle prior to ignition with the two cases. Fig.13 shows the temperature, velocity and combustion species distribution in-cylinder at the onset of ignition. Fig.14 shows the temperature distribution in-cylinder Table 5 Operation cases with different injection strategy

Injection strategy SOI1 (8 ATDC) gb1 (mg) SOI2 (8 ATDC) gb2 (mg)

Case1

Case2

Single-stage 20 21.6 – –

Two-stage 20 18 270 3.6

Fig. 11. Mixture formation with two-stage injection strategy.

after ignition. Since the temperature in-cylinder is inhomogenenous, SOC was taken to be the crank angle at where local temperature exceeds 1150 K. Since CH2O formed during the low-temperature reactions of HCCI combustion [26] and OH is ignition radical during high-temperature reactions, the first Ignition spot can be found from the distribution of CH2O and OH concentration. In case1, homogeneous mixture was formed except boundary layers at the end of compression stroke. The maximum concentration of CH2O and OH occurs at the bottom of the cylinder head. It is obviously seen that the first ignition spot locates at this area and onset of ignition occurs at 357 8CA. After ignition, the average temperature in-cylinder increases rapidly to the maximum (1998 K, 361 8CA). The temperatures in the whole combustion chamber are almost same. This is due to a great amount of spots in the whole combustion chamber

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Fig. 12. Temperature and combustion species distribution prior to SOC.

are ignited simultaneously, which leads to rapid heat release. Since HCCI combustion has a characteristic of low-temperature and short duration, the calculated NO emissions are very low (12 ppm in case1, see Fig. 15).

While in case2, thermal and concentration stratified mixture was formed at the end of compression stroke due to the second fuel injection in compression stroke. The temperature incylinder decreases from boundary to center and the fuel

Fig. 13. Temperature and combustion species distribution at SOC.

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Fig. 14. Calculated temperature in-cylinder distribution after SOC.

concentration in-cylinder increases form boundary to center due to fuel evaporation of second injection and squish flow during compression stroke. On the one hand, the richer mixture at the center with the lower ratio of specific heats (k) which leads to lower-temperatures after the compression. On the other hand, the latent heat of second injected gasoline decrease the temperature of fuel rich zone at the center of cylinder. As a result, the temperature of lean zone exceeds 40 K than that of rich zone at the center of cylinder at 350 8CA. Therefore, the temperature and concentration are both high at the periphery of rich zone (fz1), this area is easy to ignite. From Fig. 13b, it can be seen that, the ignition spot first occurs at 355 8CA at the periphery of rich zone and then at the center, finally at the remained regions. Comparison of case1 and case 2 (split injection or not), the split injection makes ignition timing advance from 357 8CA to 355 8CA because of thermal and concentration stratification by late fuel injection. It indicates that second fuel injection in compression stroke can trigger the auto-ignition. Meanwhile, stratified burning slows the combustion rate. Therefore, twostage injection is flexible strategy to control ignition timing and combustion rate. However, temperature in-cylinder of case2 is inhomogeneous. Since many spots in fuel-rich zone are ignited

simultaneously, this leads to the local high-temperature at the center as shown in Fig.14. The local temperature increases rapidly to the maximum (2435 K, 360 8CA). As a result, the NOx is formed rapidly (298 ppm) as shown in Fig.15. It is inferred that SCCI can result in advanced ignition combined with slightly high NOx emissions (298 ppm). The calculated NOx emission agree with the experimental results (320 ppm) at case2 as shown in Fig.15. NOx emissions are still much low than that of SI combustion mode. The low NOx emissions attribute to two-zone HCCI. Since the severe knocking occurs at case1 in test engine, NOx emission is not given out. The stratified charge compression ignition (SCCI) formed by first injection at early intake stroke and second injection at middle compression stroke can realize two-zone HCCI, surrounding ultr-lean HCCI combining with center ultr-rich HCCI. Since the combustion during injection was avoided, no droplet evaporation and burning occur in combustion duration. Two-zone HCCI leads to slightly low NOx and soot. Since there is no flame propagation due to no spark ignition, and the slower heat release was due to the sequential heat release by HCCI of the different zones, which had different temperatures and compositions. Two-zone HCCI has advantage in avoiding knocking. Therefore, the HCCI operation range can be extended at higher load. 7. Conclusion

Fig. 15. NOx emissions at different injection strategy.

(1) The improved 3D-CFD/chemistry model is able to predict mixture formation, combustion and emission process under compression ignition conditions for high-octane fuel. The results provide a detailed insight into the processes governing combustion and pollutant formation in the HCCI engine. (2) For compression ignition engine fueled with gasoline-like fuel, homogeneous charge can be realized by single fuel injection in intake stroke and stratification charge can be obtained by the second fuel injection in compression

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(3)

(4)

(5)

(6)

stroke. The periphery of fuel-rich zone formed by the second injection ignited first, and then fuel-rich zone ignited and worked as an initiation to ignite the surrounding lean mixture zone formed by the first injection. The second fuel injection in middle of compression stroke creates stratified charge compression ignition (SCCI). Judging ignition spots distribution from the 3D simulation results, the SCCI can be regarded as two-zone HCCI, rich HCCI at the center and lean HCCI at the circumference. Two-zone HCCI creates advanced ignition and stratified combustion, this makes ignition timing and combustion rate controllable. Two-zone HCCI can improve load level. Meanwhile, the periphery of fuel-rich zone leads to fierce burning, which results in slightly high NOx emissions. The modeling strategy of 3D-CFD/chemistry coupling for engine simulation in this paper is able to meet the requirements of execution time acceptable to simulate the whole HCCI engine physicochemical process including intake, compression, spray and combustion process.

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