CFD-assisted analysis of the characteristics of stratified-charge combustion inside a wall-guided gasoline direct injection engine

Accepted Manuscript CFD-assisted analysis of the characteristics of stratified-charge combustion inside a wall-guided gasoline direct injection engine Yu-Hsuan Su, Ting-Fu Kuo PII:

S0360-5442(19)30435-9

DOI:

https://doi.org/10.1016/j.energy.2019.03.031

Reference:

EGY 14867

To appear in:

Energy

Received Date: 24 September 2018 Revised Date:

19 January 2019

Accepted Date: 6 March 2019

Please cite this article as: Su Y-H, Kuo T-F, CFD-assisted analysis of the characteristics of stratifiedcharge combustion inside a wall-guided gasoline direct injection engine, Energy (2019), doi: https:// doi.org/10.1016/j.energy.2019.03.031. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.

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CFD-assisted Analysis of the Characteristics of Stratified-charge Combustion inside a Wall-guided Gasoline Direct Injection Engine Yu-Hsuan Su∗, Ting-Fu Kuo

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43, Keelung Rd., Sec. 4, Taipei, Taiwan, R.O.C.

Abstract

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The thermodynamic characteristics of stratified-charge combustion inside a wall-guided gasoline direct injection (GDI) engine are investigated experimentally and the effects of start of injection (SOI) on the combustion stability are studied in this work. Numerical simulation of the gas flow inside the GDI engine starting from the beginning of intake stroke to the point of start of spark ignition (SOS) are conducted. By coupling the heat release rate analysis with the geometry data of gas mixture composition obtained from numerical simulation, substantial insight into the combustion behaviors can be obtained. The effect of cycle-to-cycle variation on the combustion stability can be predicted by the equivalence ratio distribution obtained from numerical simulation. Results show that with the same amount of fuel injected and the same timing of ignition the stratified-charge combustion produces 10% larger imep than that produced by stoichiometric combustion at the expense of an insignificant combustion stability if proper injection timing is adopted. Emission data obtained from the exhaust gas analyzer indirectly confirmed the effects of SOI on the combustion behaviors inside the GDI engine. Keywords: gasoline direct-injection, stratified-charge, combustion stability, cycle-to-cycle variation.

∗

Corresponding author

Preprint submitted to Elsevier

January 19, 2019

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1. Introduction

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Driven by the stringent requirement of fuel efficiency and the Euro-5 particle matter emission regulation which came into force from 2009, gasoline direct injection (GDI) engines are embraced by the automotive industry with rapid and broad adoption since then. GDI engine achieves high fuel efficiency by switching between stratified-charge mode when running at part load or low speed and stoichiometric mode when running at high load or high speed. In stratified-charge mode, the throttle is wide open during the intake stroke thus the pumping loss can be minimized [1]. Gasoline is pressurized in the high pressure common-rail fuel line and is injected into the cylinder only at the later stage of compression stroke. This act not only prevents the occurrence of engine knock (thus the compression ratio of GDI engines can be as high as 12 [2]) but also limits the penetration depth of injected fuel due to the high in-cylinder back pressure and creates a relatively rich composition confined near the spark plug and ultra lean composition elsewhere in the cylinder. The composition is stratified and the overall air/fuel ratio can be as high as 40 as opposed to 14.6 in stoichiometric mode [1], [2], [3], [4]. According to the mechanism of introducing the stratified charge, GDI engines can be categorized into 3 types: spray-guided, air-guided, and wallguided. Spray-guided engines require that the fuel injectors must be installed in close proximity to the spark plug so that stratified mixture can be formed near spark plug regardless of the injection timing [5]. With the limited space available on the cylinder head, this may be a challenging design task. However, spark-fouling cases due to spark gaps locating in vicinity of fuel injectors have been noted [6]. Air-guided engines require that air conduit including intake manifold and intake valve position to be properly designed so that the injected fuel can be confined near the spark plug by the tumbling and/or swirl motions of intake air. Wall-guided engines make use of a specially shaped piston head to transport the inject fuel to the location of spark plug by the reverse tumbling motion of surrounding intake air flow [7]. However, fuel wetting of the piston may result in undesirable increase of hydrocarbon (HC) emissions [8]. While better fuel efficiency and higher power output can be achieved for GDI engines running in stratified-charge mode at part load [1], soot/NOx emissions increase dramatically at high loads [9],[10]. Since the three-way catalytic converter (TWC) is unable to remove NOx emissions produced under lean-burn condition effectively, this prevents GDI engines to run in stratified2

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charge mode at high loads. If the engine speed exceeds approximately 3000 rpm, there is simply not enough time for the fuel droplets to evaporate and to form the stratified charge. This sets an upper limit in engine speed for GDI engines running in stratified-charge mode [11]. GDI engines switch to stoichiometric mode when operating at high load and/or high speed. In stoichiometric mode, the fuel is injected at the beginning of intake stroke. The injected fuel has enough time to mix with the intake air during the intake and compression strokes to form a homogeneous charge. Meanwhile, evaporation of fuel helps lower down the cylinder temperature. Thus volumetric efficiency is increased. The air/fuel ratio is maintained at stoichiometric by the feedback control signal coming from a wide band oxygen sensor sent to the electronic control unit (ECU). Nowadays most GDI engines are equipped with turbocharger to facilitate the intake air flow and increase the volumetric efficiency when the engine is running at high load/speed [12]. Due to the time required for flame propagation and the ignition delay, the start of spark (SOS) is advanced before the piston reaches the top dead center so that the in-cylinder pressure is peaked around or slightly after the top dead center. Depending on the mixture quality in the vicinity of spark plug at the moment of SOS, the spark discharge may or may not initiate the combustion process successfully. It has been known for decades that random variations in combustion process on a cycle-by-cycle basis persist even if the spark ignition engine is operated under steady-state conditions apparently. Ozdor et al. [13] gave a very interesting review of cyclic variability in spark ignition engines. Cycle-by-cycle variations have long been known to be the major limiting factor for the stable operations of SI engines [14], [15]. The principal factors influencing cycle-by-cycle variations in combustion are cyclic cylinder charging, in-cylinder mixture motion, and the spatial inhomogeneity of mixture equivalence ratio [14], [16] ,[17]. Uncontrolled mixture formation is the dominant cause for the cyclic variability in GDI engines. Unfortunately, there is no way of knowing the exact distribution of mixture composition inside a GDI engine in practice. The only information we have regarding the composition is the overall air/fuel ratio obtained from the wide band oxygen sensor. Even for the homogeneous charge (the composition is presumed to be stoichiometric), the cycle-by-cycle variation may still be quite significant and it is particularly noticeably when engine is at low load or idle speed. All these phenomena complicates the understanding of combustion behaviors inside a GDI engine. Perhaps, a more 3

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interesting and important question to ask is the combustion stability of a GDI engine, which is essential to maintain the smooth engine operation. In this paper, numerical simulations of the mixing processes of in-cylinder air/fuel mixture are conducted to obtain the averaged distribution of equivalence ratio of air/fuel mixture at the time of spark. Since the random flow perturbation near the spark plug may result in the cycle-by-cycle variation in combustion, we may infer the combustion stability by the size of the region centered at spark plug gap with equivalence ratio that is within the flammability conditions (ϕ = 0.9 − 1.6, [18] ). These predictions will be checked against the experimental data. In addition early injection results in over-mixed lean mixture and leads to low combustion efficiency, while late injection results in under-mixed and locally rich mixture and soot emission may increase due to incomplete combustion. It is found that indicated mean effective pressure is increased as the injection timing was retarded [19]. However, late injection stratified-charge operation is susceptible to combustion instabilities [20]. Therefore, the success of GDI engine relies critically on the proper timing of fuel injection (start of injection, SOI). We will investigate the effect of injection timing on the combustion characteristics and stability of a wall-guided GDI engine experimentally. Combustion characteristics were studied by analyzing the in-cylinder pressure curves obtained at various injection timings. Combustion stability was assessed by the IMEP and the coefficient of variation (COV) of IMEP calculated. The ability to understand the in-cylinder flow and its effect on combustion behaviors is critical to this research. Furthermore, steady-state measurements of O2 , CO2 , CO, unburned hydrocarbons, and NOx were also examined to justify the validity of our experiments. 2. Experimental Setup The engine used in the present study is a 4-cylinder wall-guided GDI engine equipped with turbocharger. The detailed engine specifications are given in Table 1.

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Table 1: Engine Specifications

4 stroke/4 cylinder 98.0 kW @ 5000-5500 rpm 199 N-m @ 2500-3000 rpm turbocharged Not installed common-rail direct injection 150 bar 9.6 78 mm 78.4 mm 1498 c.c. 130 mm 30.5 mm 25 mm

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Type/Configuration Max. power Max. torque Air intake EGR Injection system Max. injection pressure Compression ratio Bore Stroke Swept volume Con. rod length Inlet valve diameter Outlet valve diameter

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BOSCH AVL Z-131

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BOSCH FUEL INJECTOR

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A schematic of the experimental setup is shown in Figure 1.

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ENCODER OMRON E6C2-CWZ3E

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OXYGEN SENSOR

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Figure 1: Schematic of experimental setup.

In order to measure the in-cylinder pressure, a spark plug with integrated pressure sensor (Bosch AVL Z131) is installed on the cylinder head. This piezoelectric sensor allows measurements of pressure up to 200 bars with a sensitivity of 12 pC/bar and a linearity of ±0.5%. Fuel injectors are side mounted on the cylinder head with the tip of the injector located inside the combustion chamber. Gasoline coming from the high-pressure common rail (5-150 bars) is directly injected into the combustion chamber (enclosed by the specially shaped piston head and the cylinder head, see Figure 2), through 7 orifices at the tip of injector, each with a nominal diameter of 0.150 mm and a nominal length of 0.45 mm. The injection timing and pulse width are determined by the ECU. The crank angle is determined by an Omron 6

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E6C2-CWZ3E 200P/E rotary encoder with a maximum resolution of 2000 pulses/rev. A FR 50 eddy current dynamometer with a maximum torque of 150 Nm is attached to the engine to control the engine speed/load with an accuracy of ±0.2%. Air/Fuel ratio is determined by an fast Air-to-Fuel Ratio analyzer (ECM AFRecorder 2000A) with an accuracy of ±0.9% mounted on the exhaust manifold.

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Figure 2: The specially shaped piston head.

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To investigate the combustion characteristics of this GDI engine under stratified-charge mode at part load, the engine speed is clamped at 1500 rpm by the dynamometer with wide open throttle. It is mentioned in [2] that fuel economy is not significantly improved when the air/fuel ratio exceeds 30. Therefore, a fixed amount of fuel is injected every cycle so that the monitored overall air/fuel ratios are nearly 30 (more exactly, 29.6–30.2) in our experiments. The fuel injection pressure was held at 150 bar. It is reported that the maximum brake torque (MBT) spark timing of a GDI engine running in stratified-charge mode at part load occurs before top dead center during compression stroke [3]. According to our experiments, the spark timing of 22◦ BTDC produces the maximum brake torque. Thus the spark timing (SOS) is clamped at 22◦ BTDC in this paper. To examine the effect of injection timing on the combustion stability, injection timing (SOI) is swept from 48◦ BTDC to 78◦ every 2◦ . 7

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The strategy of valve opening for stratified-charge mode is quite different from that for conventional spark ignition engine. For conventional spark ignition engine, valve overlap is favored for improving volumetric efficiency when the engine is running at high speed/load. For GDI engine running in stratified-charge mode, valve overlap may deteriorate the combustion stability due to possible variations in the residual gas mass when both valves open simultaneously. In present study, valve overlap is thus avoided. Consequently, exhaust valve closing (EVC) is set at 0◦ BTDC and inlet valve opening (IVO) is set at 30◦ ATDC. To provide more time for air filling and to improve the volumetric efficiency, inlet valve closing is set at 210◦ ATDC. In addition, exhaust valve opening (EVO) is set at 176◦ BTDC. For convenience, the detailed operating conditions are summarized in Table 2.

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Table 2: Stratified-charge mode operating conditions

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Operating Conditions 1500 rpm Engine speed Inlet pressure 1.01 bar (abs) Inlet temperature 297 K Fuel injection duration 1.2 ms (10.8◦ C.A.) Fuel mass injected 0.0636 g Air Fuel Ratio 30 Injection Timing (SOI) 78◦ – 48◦ BTDC Spark Timing (SOS) 22◦ BTDC IVO ATDC 30◦ IVC ATDC 210◦ EVO BTDC 176◦ EVC BTDC 0◦

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3. Numerical Simulation

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In order to understand the in-cylinder stratified charge distribution and early flame development, applying optical methods such as Rayleigh scattering [21] and laser-induced fluorescence (LIF) to obtain images of vaporized fuel inside a research engine via optical access through the piston crown has attracted the interests of engine investigators recently. A very interesting review paper of applying various optical methods to the research of gasoline engines is given by Drake and Haworth [22]. Still there is no way of obtaining the composition of air/fuel mixture inside the cylinder quantitatively, numerical simulation may be the only viable resort in studying the stratified air/fuel mixture formation inside a GDI engine. To simulate the extremely complicated gas flow inside the engine and the formation of the stratified charge, a realistic model of the volume enclosed by the specially shaped piston head, cylinder wall, and cylinder head is constructed. To account for the piston and valve motions, the model is discretized using sliding mesh. After grid independency test, the number of elements in the mesh used in current study is 776,480 when piston reaches top dead center(see Figure 3) and 2,815,171 when piston reaches bottom dead center. 9

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Figure 3: The finite element mesh when piston reaches the top dead center.

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In present study, air is modeled as an ideal gas. Gas composition at intake manifold is O2 (22.96 wt.%), CO2 (0.06 wt.%), H2 O (0.81 wt.%), and N2 (76.17 wt.%). Due to the strong vortex motion inside the cylinder, the Navier-Stokes equation governing the gas flow is solved using RNG k − ε turbulence model. Fuel injection of liquid gasoline is modeled via the plain-orifice atomizer model (a phenomenological model provided in ANSYS Fluent). The liquid droplets (discrete phase) dispersing in the air (continuous phase) may break up into smaller droplets or coalesce to form bigger droplets due to collisions. Discrete Phase Model (DPM) and KH-RT breakup model (both provided in ANSYS Fluent) are both enabled to simulate these processes in our numerical simulations. According to the experiments of Preussner [23], the injection of liquid gasoline with a rail pressure of 150 bars, an in-cylinder pressure of 5.6 bars, and an in-cylinder temperature of 453K will produce a droplet size distribution with a Sauter Mean Diameter (SMD) of 20 µm. Khan et al. [24] reported droplet sizes at 10 mm down the exit of a 6-hole injector in the range of 15-20 µm with an injection pressure of 100 bar as well as a chamber pressure of 1 bar. Similar droplet size ranges are reported by Marriott et al. [25] and Alieferis et al. [26]. These injection conditions are close to the operating conditions in our experiments. By setting the parameters B1 = 40, Cτ = 1, CL = 30, and CRT = 2.5 in the KH-RT breakup model, a droplet size distribution with a SMD of 20.6 µm is obtained in our numerical simulations. 10

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To justify the validity of our FEM calculation, 20 cycles of the in-cylinder pressure curves when engine was driven by motor at the speed of 200 rpm are plotted together with the averaged static pressure curve of the numerical result (red solid line) in Figure 4. For clarity, only the first revolution (0–360◦ C.A.) of the numerical result are plotted. Other than the early stage of intake stroke (0–90◦ C.A.), the numerical result agrees well with the in-cylinder pressure curves obtained by motoring. The difference in peak pressures are less than 1%. The discrepancy of the pressure curves in the early stage may be attributed to the fact that the highly non-uniform pressure distribution during intake stroke may significantly biased the averaged static pressure curve of the numerical result.

Figure 4: Validation of numerical calculation. To justify the validity of the FEM calculation, the 20 cycles of the in-cylinder pressure curves when engine was driven by motor at the speed of 200 rpm are plotted (from 0–720◦ C.A.) together with the numerical result (red solid line).

4. Experimental Results In this work, experiments were carried out by varying the start of injection (SOI) from 48◦ CA BTDC (late injection) to 80◦ CA BTDC (early injection) 11

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to investigate their effects on the lean burn combustion performance (air/fuel ratio all fixed at 30) of the GDI engine running in stratified-charge mode at 1500 rpm. For all cases, the same amount of fuel is injected in a period of 1.2 ms (10.8◦ C.A.). For the purpose of comparison, an additional experiment was carried out by injecting the same amount of fuel (0.00636 g) early at the intake stroke and maintaining the air/fuel ratio at stoichiometric so that the GDI engine is running in the homogeneous-charge mode.

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4.1. In-cylinder Pressure curves In-cylinder pressure data can be used a means of analyzing combustion characteristics and estimating the apparent heat release rate (AHRR). However, substantial cycle-by-cycle combustion variations can be observed from the in-cylinder pressure versus time measurements for successive cycles. Incylinder pressure curves for various cases with different injection timings are shown in Figure 5. Here for the purpose of clear demonstration, only one in-cylinder pressure curve out of every 30 cycles selected from 300 consecutive cycles is plotted . In addition, the pressure curves with maximum peak pressure and minimum peak pressure are also plotted on the same diagrams. Since the peak pressure for cold flow driven by motoring is about 19 bars for our GDI engine, this sets a lower envelope for the pressure curves. In other words, pressure curves with a peak pressure of 19 bars in Figure 5 are deemed to be misfire cases. It can be easily identified that no misfire occurs for the cases with SOI = 60–70◦ BTDC. In Figure 6, the in-cylinder pressure curves with maximum peak pressure (the best scenario cycle or the almost complete combustion cycle) for various injection timings are plotted together. Since the same amount of fuel is injected and the air/fuel ratio is kept nearly 30 for all cases (black) except for the homogeneous-charge case (red), these pressure curves are almost identical at the intake stroke (from 0◦ to 240◦ C.A.). Since the air/fuel ratio for homogeneous charge case is kept at stoichiometric (14.6), less air is pumped into the cylinder. The corresponding pressure curve (red) is obviously lower than all the other cases (black). The pressure curves diverge from one another starting from SOS (338◦ C.A.) due to combustion. All the pressure curves (black) converge toward one another at the late stage of expansion stroke, since there are almost the same amount of exhaust gas and roughly the same amount of heat released after completing combustion in each case. The corresponding p − V curves are plotted in the top plot of Figure 7. A close-up view of the pumping loops for various cases is shown 12

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in the bottom plot of Figure 7. It can be clearly observed that due to wide open throttle the pumping loops for stratified-charge combustion cases are negligibly small (yellow parts corresponding to exhaust blow-down and blue parts corresponding to the fresh air intake) compared to the pumping loop for the homogeneous charge combustion (red dotted curve). By numerical integration, we can calculate the work done during the best scenario cycle for various injection timings and these values (Wmax ) are also labelled in Figure 5. For best scenario cycles, the work done for the homogeneous charge mode (Wmax = 178.2J) is always lower than those for the stratified charge mode because excess air is pumped into engine cylinder for the stratified charge mode. The peak in-cylinder pressure pmax and the crank angle at which the peak pressure occurs θpmax vary on a cycle-by-cycle basis. The best scenario cycle does not give us a fair judgement of the ensemble combustion characteristics for various injection timings. To investigate the ensemble combustion characteristics for various injection timings, the histograms of the peak in-cylinder pressure pmax and the crank angle at which the peak pressure occurs θpmax are plotted in Figure 8 and 9, respectively. For the late injection cases such as SOI = 48◦ BTDC, lots of misfire cycles can be found probably due to insufficient time for proper mixing, see Figure 8. The ensemble mean in-cylinder peak pressure is 20.34 bars, which is very close to that of a cold flow case. The engine is barely operable under this partial-burn condition. In practice, there is no way of knowing why the mixture fails to ignite in a significant fraction of cycles. Numerical simulation provides us an ideal solution to this problem. Figure 10 shows 6 side views of equivalence ratio of the gas mixture inside engine cylinder cut through the spark gap at crank angle 37.2◦ BTDC (end of fuel injection), 33.5◦ BTDC, 31◦ BTDC, 28◦ BTDC, 25◦ BTDC, and 22◦ BTDC (SOS), respectively. The process of stratified charge formation was clearly captured. Gas mixture hits on the piston head at 37.2◦ BTDC and is deflected toward the spark plug at 33.5◦ BTDC. The tumbling motion brings the mixture to the cylinder head at 31◦ BTDC and the mixture is separated into two portions at 25◦ BTDC. At 22◦ BTDC, two stratified charges can be clearly identified. Figure 11 shows the corresponding 6 top views of equivalence ratio of the gas mixture inside engine cylinder cut through the spark gap. As shown in Figure 10, the gas mixture is divided into 2 parts due to the mixture motion. From Figure 12, the equivalence ratio of the stratified mixture at the spark gap is around 2.6 at the time of spark (SOS), which is too rich to ignite. For reliable flame propagation and secured com13

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bustion, the equivalence ratio should be in the range ϕ = 0.77 · · · 1.33 [32], [33]. However, this numerical result comes from the simulation of a single engine cylinder. There are 4 cylinders in the real GDI engine. The amount of air fed into a cylinder varies cylinder-to-cylinder, variabilities from RAM effect, valve timing and injection nozzle variation, all these may lead to the combustion variability. As SOI increases, the number of misfire cycles decreases rapidly and the ensemble mean in-cylinder peak pressure increases. Meanwhile, the standard deviation decreases indicating that better combustion stability is attained. The highest ensemble mean in-cylinder peak pressure (pmax = 36.6 bars) and the smallest standard deviation in peak pressures (σpmax = 1.82 bars) both occur at SOI = 62◦ BTDC. From Figure 12, the gap of spark plug is surrounded by a large area with equivalence ratio around 0.8 · · · 1.3 and this explains the better combustion stability for the case SOI=62◦ . For early injection cases such as SOI=78◦ BTDC, misfire cases start to occur again. Due to the early injection, the fuel gas has sufficient time to mix with air and spread out. Therefore, the composition becomes less stratified or even nearly homogeneous but with lower equivalence ratio, see Figure 12. This lower equivalence ratio results in the slow burning cycles. For some cycles, the burning rate is so slow that combustion is not completed by the time the exhaust valve opens. These partial-burn cycles can be clearly observed from Figure 5 that the tails of some cycles does not approach the tail of the best scenario cycle. This can also be confirmed from the emission data that will be discussed later. Note that the spark plug is at the boundary between gas with low but barely combustible equivalence ratio and fresh air, see Figure 12. It is no doubt that the combustion stability is poor for the cases with SOI=78◦ .

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15 Figure 5: In-cylinder pressure versus crank angle curves for cases with various injection timings. For clarity, only one out of every 30 cycles selected from 300 consecutive cycles are plotted.

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Figure 6: In-cylinder pressure curves of best scenario cycles for stratified-charge combustion cases with various injection timings (black, solid) for the engine running at 1500 rpm with an averaged imep of 5 bars. For comparison, the case for homogeneous combustion is also plotted (red, dotted).

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Figure 7: Top plot: In-cylinder pressure versus volume curves of best scenario cycles for stratified-charge combustion cases with various injection timings for the engine running at 1500 rpm with an averaged imep of 5 bars. For comparison, the case for homogeneous combustion is also plotted (red, dotted). Bottom plot: Close-up of the pumping loops for various cases with different injection timings.

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Figure 8: Histograms of peak cylinder pressures with the engine running at 1500 rpm and an averaged imep of 5 bars.

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std. = 4.30 bar

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std. = 3.52 bar

mean = 31.70 bar

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mean = 36.37 bar

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std. = 2.52 bar

mean = 35.37 bar

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30

SOI = 68° BTDC

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std. = 1.82 bar

mean = 36.60 bar

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SOI = 62° BTDC

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std. = 2.11 bar

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mean = 35.39 bar

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SOI = 56° BTDC

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Figure 9: Histograms of crank angles at which peak in-cylinder pressures occurs with the engine running at 1500 rpm and an averaged imep of 5 bars.

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std. = 2.99 °

mean = 370.3 °

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370

θp max

(°)

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SOI = 78° BTDC

std. = 1.86 °

mean = 370.6 °

365

370

SOI = 72° BTDC

std. = 1.48 °

mean = 369.4 °

365

370

SOI = 66° BTDC

std. = 1.30 °

mean = 368.5 °

365

370

SOI = 60° BTDC

std. = 1.73 °

365

370

SOI = 54° BTDC

mean = 368.5 °

365

std. = 3.04 °

mean = 362.1 °

SOI = 48° BTDC

375

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SOI = 80° BTDC

std. = 1.83 °

mean = 370.9 °

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370

SOI = 74° BTDC

std. = 1.51 °

mean = 369.8 °

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std. = 2.57 °

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Homogeneous mean = 373.0 °

365

std. = 2.36 °

375

375

375

375

380

380

380

380

375

375

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std. = 1.55 °

370

SOI = 70° BTDC mean = 370.4 °

365

std. = 1.42 °

370

SOI = 64° BTDC

mean = 369.0 °

365

std. = 1.30 °

370

SOI = 58° BTDC mean = 368.4 °

365

std. = 2.52 °

370

SOI = 52° BTDC mean = 368.1 °

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std. = 1.31 °

mean = 368.9 °

365

370

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SOI = 62° BTDC

std. = 1.36 °

365

370

SOI = 56° BTDC

mean = 368.7 °

365

std. = 4.22 °

mean = 366.3 °

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0 360

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SOI = 50° BTDC

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Figure 10: Side views of the distributions of equivalence ratio of gas mixture (SOI=48◦ BTDC) cut through the spark gap at crank angle 37.2◦ , 33.5◦ , 31◦ , 28◦ , 25◦ , and 22◦ BTDC, respectively.

Figure 11: Top views of the distributions of equivalence ratio of gas mixture (SOI=48◦ BTDC) cut through the spark gap at crank angle 37.2◦ , 33.5◦ , 31◦ , 28◦ , 25◦ , and 22◦ BTDC, respectively.

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Figure 12: Top views of the distributions of equivalence ratio of gas mixture cut through the spark gap (’+’) at the time of spark (SOS) and at crank angle 48◦ , 62◦ , and 78◦ BTDC, respectively.

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4.2. Characteristics of combustion stability The characteristics of combustion stability for various injection timings are best illustrated in Figure 13. Here the peak cylinder pressure pmax and the crank angle at which the peak cylinder pressure occurs θpmax for each individual cycle is represented as a dot on the plane. For clarity, only the cases with SOI=48 (blue triangles), 62 (green dots), and 78◦ (black crosses) C.A. BTDC are plotted. For comparison, the case of homogeneous-charge combustion (red circles) is also plotted on the same diagram. For SOI=48◦ BTDC, the blue triangles are distributed into two separated regions: one dense cluster on the lower left corner labelled with misfire and the other minor cluster on the central part of the diagram. This indicates that the engine operation is predominately consisted of misfire cycles together with some sporadic cycles with successful ignitions. Thus the combustion stability is poor and the engine is barely operable for SOI=48◦ BTDC (late injection case). For SOI=62◦ BTDC, the green dots are closely packed within a cluster in the central part of the diagram. Obviously, the combustion stability is excellent and the cylinder pressures are peaked within a narrow window with θpmax = 369 ± 3◦ . By juxtaposing this case with the homogeneous-charge case (red circles) to awake our minds, the red circles are closely attached to a straight line and the cylinder pressures are peaked within a relatively wide window with θpmax = 373 ± 7◦ due to cycle-by-cycle variation. As noted in [28], the spark timing for homogeneous-charge combustion has to be set for average cycle due to the large cycle-to-cycle variation in combustion rate. Under this compromise the fastest burning cycles are likely to knock before spark discharge and the slowest burning cycles are most likely to burn incompletely. This situation is much alleviated for our current case with SOI=62◦ BTDC. For SOI=78◦ BTDC, the black crosses are scattered all over the diagram. Clearly, the combustion stability is poor. With the large combustion variability (black crosses), knock and partial burning cycles can be expected. Even without knowing the exact cause of cycle-to-cycle variation, the effects of cycle-to-cycle variation on the combustion stability can be inferred from the equivalence ratio distribution near spark plug at the time of spark discharge as will be explained in the next section.

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p

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θp

370

(o)

372

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max

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Figure 13: Individual-cycle peak pressure pmax versus crank angle at which it occurs. This plot demonstrates the stability of stratified charge combustion for SOI=48◦ (blue triangles), 62◦ (green dots), and 78◦ (black crosses) BTDC, respectively. For comparison, the case of homogeneous-charge combustion (red circles) is also plotted.

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pdV Vd

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4.3. Indicative mean effective pressure In practice, the net work delivered to the piston over the four strokes of the cycle determines how much power is available to the drive shaft of a vehicle. In present work, the net indicative mean effective pressure, imepn , the net work per cycle per unit displaced volume:

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is used to characterize the power output of the GDI engine. Here Vd is the displaced volume of the GDI engine. In addition, a quantitative measure of cyclic variability COVimep defined as the standard deviation σimep in indicated mean effective pressure (imep) divided by the ensemble averaged imep, µimep , [28]: COVimep ≡

σimep µimep

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is adopted in present study to characterize the stability of engine operation. From experience, vehicle driveability is poor when COVimep exceeds 10% [28]. The net imep of 300 consecutive cycles for various injection timings are plotted in Figure 14. For convenience, the ensemble averaged imep µimep and the cyclic variability COVimep for each injection timing are both labelled on the diagrams in Figure 14. It can be easily observed that for late injection such as SOI=48◦ BTDC a substantial fraction of cycles fail to ignite and thus result in zero imep. Only a small fraction of cycles do ignite and produce barely enough imep to sustain the engine operation intermittently. The ensemble average of imep is as low as 0.73 bar with a COVimep as high as 283%. Obviously, the engine is barely operable when SOI=48◦ BTDC. As indicated by the numerical simulation, the spark gap is located within a small region of too high equivalence ratio at the time of spark discharge when SOI=48◦ BTDC. Undoubtedly, misfire will occur predominantly in this case. Due to the cycle-by-cycle variations, the small region of high equivalence ratio may be carried away from the spark gap by the turbulent gas motion and this may lead to the sporadic success of ignition in a small fraction of cycles. Both the number of misfire cycles and cyclic variability COVimep decrease drastically as injection timing is advanced. The ensemble averaged imep µimep reaches its maximum value of 5.3 bar with a decently tight cyclic 24

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variability COVimep of 3.59% when SOI=62◦ BTDC. The result of numerical simulation showed that the spark gap is surrounded by a broad region with decent equivalence ratio at the time of discharge when SOI=62◦ BTDC. The region is so broad that the cycle-by-cycle variation in gas motion can not destabilize the ignition process but only perturb the combustion rate slightly. Consequently, no misfire cycle occurs and cyclic variability COVimep is small. As injection timing is further advanced, the gas mixture moves away from the spark gap toward the cylinder wall and becomes less stratified. Since there is enough time for the gas mixture to spread out before spark discharge, the gas mixture becomes lean and may result in the slow burn of fuel. In addition, cyclic variations becomes significant. For some cycles partial burn can be expected if the combustion is so slow that the exhaust valve opens before the combustion is complete. Again from the numerical simulation, we know the spark gap is on the boundary of gas mixture and fresh air at the time of spark discharge when SOI=78◦ BTDC. Therefore, cycle-by-cycle variation may lead to the misfire in some cycles. In this case, the ensemble averaged imep is 4.35 bar with a significant cyclic variability COVimep of 28.29%. As for the homogeneous charge case, the gas mixture is almost uniform with equivalence ratio of 1 everywhere inside the cylinder. Consequently, no misfire occurs. The ensemble averaged imep is 4.83 bar with a very small cyclic variability COVimep of 1.40%. The ensemble averaged imep of this homogeneous charge combustion is about 10% lower than that of stratifiedcharge combustion with SOI=62◦ BTDC due to the fact that excess air is pumped into the cylinder for the stratified-charge combustion case.

25

4

4

2

2

0

0

-2

-2 0

50

100

150

200

250

300

SOI = 54°, 5.05 bar, 13.20 %

6

SOI = 50°, 3.42 bar, 77.11 %

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4 2 0

-2

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50

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50

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250

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SOI = 60°, 5.23 bar, 4.16 %

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2 0 -2 0

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SOI = 66°, 5.31 bar, 4.09 %

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-2

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SOI = 62°, 5.30 bar, 3.59 %

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0 -2 0

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SOI = 68°, 5.28 bar, 4.58 %

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SOI = 74°, 4.99 bar, 9.68 %

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Cycle

Figure 14: Distributions of the individual-cycle IMEP for 300 consecutive cycles for the GDI engine running with various injection timings. For the purpose of comparison, the case of homogeneous-charge combustion is also plotted.

26

300

300

300

300

6

2

100

300

Homogeneous, 4.83 bar, 1.40 %

SOI = 80°, 3.44 bar, 48.99 %

4

50

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SOI = 76°, 4.72 bar, 16.35 %

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SOI = 70°, 5.27 bar, 5.42 %

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4

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SOI = 78°, 4.35 bar, 28.29 %

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SOI = 72°, 5.12 bar, 9.00 %

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SOI = 64°, 5.29 bar, 3.57 %

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SOI = 58°, 5.18 bar, 7.48 %

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SOI = 56°, 5.14 bar, 7.51 %

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SOI = 52°, 4.77 bar, 28.90 %

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4.4. Net Heat Release Rate The lean burn combustion characteristics of the stratified charge differs significantly from those of the homogeneous charge due to the difference in air/fuel mixture composition. The reaction rate of the rich mixture region near the spark plug is rapid and it is reduced when the flame propagates into the lean outer region [5]. The net heat release rate inside the GDI engine excluding the heat transfer via cylinder wall can be estimated based on the formula given in Heywood [28]: dQnet γ dV 1 dp AHRRnet = = p + V , (1) dt γ − 1 dt γ − 1 dt

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where γ is the specific heat ratio, p the in-cylinder pressure, V the volume of combustion chamber. In this work, a constant value of γ = 1.3 is set for all cases. However, we have to keep in mind that this value is just an approximate. Therefore, only qualitative conclusion can be drawn and quantitative comparison may be meaningless. Since the in-cylinder pressure obtained from the pressure sensor is noisy, dp in Eq. 1 may incur spurious oscillations in the calculation of net the term dt heat release rate. Therefore, the pressure curves are first plugging into Eq. 1 and filtered with forward and backward Butterworth filters to zero the phase lag due to smoothing. To demonstrate the cyclic variations in combustion characteristics for various injection timings SOI = 48◦ , 62◦ , and 78◦ BTDC, five typical net heat release rate profiles for each individual SOI are selected and plotted in Figure 15. For the purpose of easy comparison, the five net heat release rate profiles for homogeneous-charge combustion are also plotted here. Each profile is identified by the pair (pmax , θpmax ) given in the legend. By coupling the heat release rate analysis with the geometry data of gas mixture composition provided in Figure 12, substantial insight into the combustion behaviors can be obtained. For late injection case such as SOI=48◦ BTDC in Figure 15, following spark discharge (338◦ C.A.) there is a period during which the net heat release rate is indiscernible since the equivalence ratio at spark plug gap is too high at the moment of SOS. Thus misfire cycles dominate and sporadic inflammation is only possible as a result of cycle-to-cycle variation. In addition, the combustion behaviors of the sporadic cycles with successful ignition are characterized by a leading dominate peak of net heat release rate followed by a minor peak. For injection case SOI=62◦ BTDC in Figure 15, as can be seen in Figure 12 the spark plug gap 27

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is surrounded by a large region with equivalence ratio suitable for combustion at the moment of SOS. Immediate and stable ignitions can be expected. Gas mixture is apportioned into two parts by the gas motion past through the specially-shaped piston head as shown in Figure 12. The combustion behaviors are thus characterized by a double peak heat release rates. Combustion rates vary on a cycle-to-cycle basis. Fast burning cycles have a leading dominate peak, while slow burning cycles may have maximum net heat release rate at the second peak. For early injection case SOI=78◦ BTDC in Figure 15, the injected fuel has sufficiently enough time to mix with air and the equivalence ratio of gas mixture approaches nearly uniform as can be seen in Figure 12. Consequently, the combustion rates are in general slow and more or less flat. For some slow burning cycles (e.g. the cycle labelled as x5 in Figure 15), combustion is not completed by the time exhaust valves open. For the homogeneous-charge case, no double peaks in heat release rate can be found, see Figure 15). The net heat release rate for fast burning cycle is higher than that for slow burning cycle by 67%.

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o1 (34.03, 366.00) o2 (31.59, 369.80) o3 (27.66, 373.20) o4 (26.40, 377.00) o5 (22.33, 380.00)

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HRR (W)

2000

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Crank angle (°)

x1 (36.60, 370.00) x2 (32.08, 372.40) x3 (28.87, 376.80) x4 (24.14, 365.60) x5 (19.15, 366.80)

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HRR (W)

•1 (40.49, 367.40) •2 (32.57, 364.80) •3 (32.81, 368.40) •4 (33.85, 369.40) •5 (34.00, 372.20)

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HRR (W)

HRR (W)

SOI=62 o BTDC

∆1 (37.56, 370.60) ∆2 (31.10, 366.60) ∆3 (31.52, 372.60) ∆4 (27.65, 368.20) ∆5 (18.08, 360.40)

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(o) max

Figure 15: Calculated net heat release rate profiles for the GDI engine running with SOI = 48◦ , 62◦ , and 78◦ BTDC, respectively. For the purpose of comparison, the case of homogeneous-charge combustion is also plotted. The location map gives the exact location of each individual net heat release rate profile on the (θpmax , pmax ) plane.

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4.5. Emissions Emission data provides us an indirect way of examining the conditions of combustion processes. In present study, untreated emission data are obtained by inserting a Horiba MEXA-584L exhaust gas analyzer in the exhaust manifold shared by 4 cylinders instead of measuring the emission from a particular cylinder. Consequently, extra care must be exerted when drawing quantitative conclusion. Since gasoline is predominantly a blend of various hydrocarbon compounds, Heywood’s formula Cn H1.87n for gasoline with n = 8 is adopted in the present work [28]. With this formula, the overall combustion equation can be written as C8 H14.96 + 11.74 (O2 + 3.773N2 ) = 8CO2 + 7.48H2 O + 44.295N2

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and the stoichiometric air/fuel ratio is exactly 14.6. Therefore, for the homogeneous-charge combustion the volume fraction of carbon dioxide in 8 exhaust is ideally about = 13.38% vol if the combustion is 8 + 7.48 + 44.295 complete and the volume fraction of oxygen in exhaust should be 0. In this work air/fuel ratio is kept at 30 for the stratified-charge mode, the overall combustion equation should be modified as follows: C8 H14.96 + 24.12 (O2 + 3.773N2 ) = 8CO2 + 7.48H2 O + 91.00N2 + 12.38O2 .

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Therefore, for the stratified-charge combustion the volume fraction of carbon 8 = 6.73% vol and the dioxide in exhaust gas is about 8 + 7.48 + 91 + 12.38 12.38 volume fraction of oxygen should be = 10.42% vol. 8 + 7.48 + 91 + 12.38 CO2 concentration is used as a measure of the completeness of combustion in present work. Figure 16 shows the volume fractions of CO2 in exhaust gas reported by the exhaust analyzer are about 6.2 for SOI=60–70◦ , which are quiet close to the theoretical value 6.73% (bold dotted line) for stratifiedcharge complete combustion. The volume fractions of CO2 drop drastically for late injection due to significant portion of misfire cycles. For early injection, the volume fractions of CO2 drop less drastically due to significant amount of partial burning cycles. As can be observed from Figure 16, the fluctuation in CO2 concentration is decently small for SOI=54 · · · 70◦ BTDC, indicating stable combustion. 30

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CO 2 (% vol)

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5 4 3 2 1 0

48 50 52 54 56 58 60 62 64 66 68 70 72 74 76 78 80

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SOI (° BTDC)

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Figure 16: Volume fraction of carbon dioxide in exhaust gas for various injection timings.

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Figure 17 shows the volume fractions of O2 in exhaust gas are about 11.5 for SOI=60–70◦ , which are quiet close to the theoretical value 10.42% (bold dotted line) for stratified-charge complete combustion. For homogeneouscharge mode, the measured volume fraction of O2 is as low as 1.6%. Ideally, the volume fraction of O2 should be 0 if oxygen sensor functions normally. Obviously, the trend of the volume fractions of O2 is opposite to that of CO2 . Formation of NOx is determined dominantly by the peak temperature in cylinder. As can be seen from Figure 18, NOx emissions for stratified-charge combustion are significantly lower than that for homogeneous-charge combustion. This can be attributed to the fact that more air is pumped into the cylinder due to the higher air/fuel ratio (30) associated with stratified-charge requirement. Therefore, in-cylinder gas temperature for stratified-charge combustion is much lower than that for homogeneous-charge combustion. This in turn leads to the reduction in NOx generation. For stratified-charge combustion, NOx emission is peaked at SOI=64◦ . If we look at the equivalence ratio distribution for SOI=64◦ , the spark plug gap is surrounded by a large area with equivalence ratio ranging between 1.0 and 1.2. From litera31

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O2 (% vol)

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SOI (° BTDC)

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Figure 17: Volume fraction of oxygen in exhaust gas for various injection timings.

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ture [33], the combustion rate peaks at slightly rich mixture on the order of equivalence ratio ϕ = 1.1 · · · 1.25. This explains the highest NOx emission occurring at SOI=64◦ . The average temperatures in the cylinder for various cases range between 666 K and 1315 K and they are also plotted in Figure 18 (black crosses). HC emissions are caused by incomplete combustion resulted from oxygen deficiency. Consequently, rich mixtures with poor homogeneity resulted from late injection produce greater HC emissions. This is undesirable for PM suppression [34]. Excessively high unburnt hydrocarbon emission at late fuel injections such as SOI=48◦ can be seen in Figure 19. As with untreated HC emissions, CO emissions are results of incomplete combustion. As can be observed from Figure 20, CO emissions decrease sluggishly for early injections. This may be attributed to the partial burning cycles. For late injections, CO emissions decrease substantially due to the significant portion of misfire cycles.

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NO (ppm vol) / T

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Figure 18: Volume fraction of NOx in exhaust gas for various injection timings. The average temperatures in the cylinders for various cases are also plotted (black crosses).

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Figure 19: Volume fraction of hydrocarbons in exhaust gas for various injection timings.

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Figure 20: Volume fraction of carbon monoxide in exhaust gas for various injection timings.

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5. Conclusions

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In present study, the effects of different fuel injection timings on the combustion characteristics and stability inside a wall-guided GDI engine are investigated experimentally. Numerical simulation was conducted to determine the air/fuel mixing characteristics such as droplet size distributions, SMD, and in particular equivalence ratio distributions under various injection timings. The effects of various injection timings on the combustion stability was evaluated using the coefficient of variation in indicated mean effective pressure, COVimep . By coupling the net heat release rate analysis with the geometric data of equivalence ratio distributions obtained from numerical simulation, a few important qualitative conclusions can be drawn:

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1. The geometric characteristics of equivalence ratio distributions depend highly on the injection timings. Late injections lead to poor mixture homogeneity with steep gradients, which are undesirable for PM suppression. In addition, the equivalence ratio under spark plug at SOS is usually too high to be ignitable and misfires occur predominately. Early injections result in relatively homogeneous but lean mixtures whose burning rates are usually slow. Consequently, partial burning cycles or even misfires were observed. 2. Even without knowing the exact cause of cycle-to-cycle variation, the effects of cycle-to-cycle variation on the combustion stability can be inferred from the equivalence ratio distribution near spark plug at the time of spark discharge. The equivalence ratio distribution for late injection such as SOI=48◦ indicates that the spark plug is surrounded by a small but finite high concentration region with steep gradients at the time of spark ignition. Consequently, one can expect that misfires occur predominately and some successful ignitions may result from the cycle-to-cycle variation. The equivalence ratio distribution for early injection such as SOI=78◦ reveals that the spark plug is located at the boundary of a relatively homogeneous but lean mixture and fresh air. Due to the slow burning of lean mixture, one can expect the existence of partial burning cycles or even misfire cycles may occur due to the cycle-to-cycle variation. 3. With the same amount of fuel injected and the same timing of ignition and even without any optimization in current experimental setup, the stratified-charge combustion produces 10% larger imep than that 36

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produced by stoichiometric combustion at the expense of some insignificant combustion stability if proper injection timing is adopted. 4. With the geometric data of equivalence ratio distributions obtained from numerical simulation, the double peak combustion characteristics of GDI engines can be well explained.

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Acknowledgments

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The authors would like to thank Mr. Hua-Yi Wang for producing the graphics of simulation results. This work is partly supported by the Ministry of Science and Technology (MOST) under the grant number MOST 104-3113E-027-002-CC2 and Hua-Chuang Automobile Information Technical Center (HAITEC). References

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[1] F. Zhao, M.C. Lai, D.L. Harrington, Automotive spark-ignited directinjection gasoline engines, Progress in Energy and Combustion Science, 25, 437–562, 1999.

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[2] Kume T., Iwamoto Y., Iida K., Murakami M., Akishino K., and Ando H., “Combustion Control Technologies for Direct Injection SI Engine”, SAE Technical Paper, No. 960600, 1996.

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[3] Fraidl G.K., Piock W.F., and Wirth M., “Gasoline direct injection: actual trends and future strategies for injection and combustion systems,” SAE Technical Paper, No. 960465, 1996.

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[4] Jackson N. S., Stokes J., Whitaker P. A., and Lake T. H., “Stratified and homogeneous charge operation for the direct injection gasoline engine high power with low fuel consumption and emissions,” SAE Technical Paper, No. 970543, 1997.

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engine,” Journal of Engineering for gas turbines and power, 139, 1028011, 2017.

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Highlights

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CFD simulation of fuel spray in a wall-guided GDI engine. Correlation of CFD distribution of equivalence ratio with combustion stability disturbed by cycleto-cycle variation. Proper injection timing under part load conditions.

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 