Review on spray, combustion, and emission characteristics of recent developed direct-injection spark ignition (DISI) engine system with multi-hole type injector

Fuel 259 (2020) 116209

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Review article

Review on spray, combustion, and emission characteristics of recent developed direct-injection spark ignition (DISI) engine system with multihole type injector Ziyoung Leea, Taehoon Kimb, Sungwook Parkb,

⁎,1

, Suhan Parkc,

T

⁎,1

a

Department of Mechanical Convergence Engineering, Graduate School of Hanyang University, 222 Wangsimni-ro, Seongdong-gu, Seoul 04763, Republic of Korea School of Mechanical Engineering, Hanyang University, 222 Wangsimni-ro, Seongdong-gu, Seoul 04763, Republic of Korea c School of Mechanical Engineering, Chonnam National University, 77 Yongbong-ro, Buk-gu, Gwangju 61186, Republic of Korea b

A R T I C LE I N FO

A B S T R A C T

Keywords: Gasoline direct-injection (GDI) Charge stratification Spray and atomization Combustion Exhaust emission Alcohol fuels

Given diminishing fossil fuel resources and severe environmental pollution, governments are strengthening their regulations regarding the exhaust emissions and fuel consumption of transportation vehicles. To satisfy the new stringent requirements for emission and efficiency, researchers have attempted to combine the advantages of gasoline and diesel engines in a single engine. Studies on gasoline direct -injection (GDI) engines have been conducted since the early 1990s. In a GDI engine, the fuel is directly injected into the combustion chamber to form a stratified air/fuel mixture for ultra-lean combustion. To achieve the proper mixture in the combustion chamber of a GDI engine, various injection and airflow strategies have been implemented, such as the multipleinjection and spray-guided techniques. In addition, various emission after-treatment devices, such as a gasoline particulate filter and lean NOx trap have been used to reduce exhaust emissions. Major injector manufacturers recently embarked on the development of a piezo-outwardly type injector and a multi-hole type injector to improve engine performance and reduce exhaust emissions. In addition, researches are actively pursuing alternative fuels, such as bioethanol and biobutanol, eco-friendly alcohol fuels, for use in GDI engines. In this review article, we examine and discuss the GDI engine system. Our objective in the review is to examine the injection, spray, combustion, and exhaust emission characteristics of a GDI engine with a multi-hole injector. We also review numerical methods for modeling a GDI engine and describe the characteristics that make alcohol fuels useable in GDI engines.

1. Introduction 1.1. Overview As shown in Fig. 1, exhaust gas and fuel efficiency regulations for the transportation sector are being made more stringent [1,2]. The depletion of fossil fuel resources and increasing demand for ecofriendly vehicles require the development of power sources with high thermal efficiency and ultra-low emissions. Those requirements can be met by changing the operating mechanism of engines and through the study and development of clean alternative fuels. This trend is reflected in the ongoing change in emphasis in the gasoline-fueled vehicle market from “performance” to “low fuel consumption” and “low CO2 emission”. The existing port fuel injection (PFI) gasoline engine has the advantage of low emission of environmentally harmful gases such as carbon

monoxide (CO), unburned hydrocarbon (UHCs), and nitrogen oxides (NOx) through the use of a stoichiometric operation (λ = 1) and the application of a three-way catalyst (TWC). However, the fuel consumption in PFI engines is higher than that in diesel engines, which have a high thermal efficiency of 30%−40%. To solve the problem of the high fuel consumption of gasoline engines, many researchers have investigated various approaches, such as downsizing the engine, reducing the vehicle weight, and using turbo-charging, lean burn, compression ignition, and direct fuel injection [3–7]. Among many new approaches, the most interesting and promising is the use of a directinjection fuel system in a gasoline engine, GDI (gasoline direct injection) engines. At the early stage of GDI engine development in 1990s, research on GDI systems focused on stratified charge gasoline injection, which targets lean combustion to achieve higher combustion efficiency [8,9].

⁎

Corresponding authors. E-mail addresses: [email protected] (S. Park), [email protected] (S. Park). 1 These authors equally contributed to this research as corresponding author. https://doi.org/10.1016/j.fuel.2019.116209 Received 14 July 2019; Received in revised form 10 September 2019; Accepted 12 September 2019 0016-2361/ © 2019 Elsevier Ltd. All rights reserved.

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Nomenclature 1D 3D AMD ASOS ATDC BTDC CA CAB CCD CCLNT CO D10 DDB DMF E85 ECFM ECU EGR EOI EOS FSI GDI GEFlash GPF GTDI HC HPDI HRM IDE IMEP

KH Kelvin-Helmholtz LCA life cycle analysis LHV lower heating value LIF laser-induced fluorescence LNT lean NOx trap LPG liquefied petroleum gas MPI multi-port injection NOx nitrogen oxides P.M. particulate matter PDA phase Doppler anemometry PDF probability density function PDPA phase Doppler particle analyzer PFI port fuel injection PM particulate mass PN particulate number RT Rayleigh-Taylor SCGI stratified charge gasoline injection SCR selective catalytic reduction SI spark ignition SMD Sauter mean diameter SOI start of injection TAB Taylor analogy breakup TDC top dean center TWC three-way catalyst UHC unburned hydrocarbon VLIM variable length intake manifold VOF volume of fluid VTPR-EOS volume-translated Peng-Robinson equation of state VVL variable valve lift VVT variable valve timing WOT wide open throttle

one-dimensional three-dimensional arithmetic mean diameter after start of spark after top dead center before top dead center crank angle cascade atomization breakup charge-coupled device close-coupled LNT carbon monoxide droplet mean diameter droplet deformation and breakup dimethyl-furan 85% ethanol + 15% gasoline extended coherent flamelet model engine control unit exhaust gas recirculation end of injection equation of state fuel-stratified injection gasoline direct injection gasoline-ethanol flash gasoline particulate filter gasoline turbocharged direct injection hydrocarbon high-pressure direct injection homogeneous relaxation model injection direct essence indicated mean effective pressure

In addition to stable mixture formation at elevated ambient pressure, multi-hole type injectors offer numerous advantages over swirltype injectors, including enhanced air entrainment caused by the separated spray jets. In addition, the spray jet from each nozzle hole can develop towards the desired location with flexible fuel amounts, which improves the matching between nozzle configuration and combustion chamber shape. Thus, most GDI engines now use multi-hole type injectors. Spray characteristics and the following combustion and emission characteristics of multi-hole GDI injectors are significantly different from those of swirl type injectors. In addition, technologies for multihole type GDI injectors have developed rapidly during the past decade. Therefore, a comprehensive review of the spray, combustion, and emission characteristics of GDI engines equipped with a multi-hole type nozzle is needed. We’ve composed the present review article in 6 sections: Sections 1–6. In the rest of Section 1, we describe the general characteristics of GDI engines, including comparisons with conventional PFI engines. We review research papers on the spray characteristics of GDI injectors with multi-hole nozzles in the Section 2. Nozzle configuration and injection strategies play key roles in the combustion and emissions performance of GDI engines. Thus, the Section 3 deals with research to enhance combustion and emission performance. In Section 4, we describe the research trends for modeling air/fuel mixture formation, combustion, and emissions in GDI engines. Finally, we discuss the effects of alcohol blending on GDI engine performance, including combustion, emissions, and spray characteristics in the Section 5.

Based on abundant research on GDI systems, Japanese companies began mass production of stratified charge gasoline injection engine: Mitsubishi Motors [10] and Toyota Motors [11] in 1996 followed by Nissan Motor in 1997 [10]. To find the optimum method to provide fuel directly into the combustion chamber, various types of GDI injectors, such as swirl-type, slit-type, and multi-hole type. The first mass-produced GDI engines used swirl type GDI injectors, which produce a hollow cone with a wide spray cone angle of 70–80 degrees under atmospheric pressure [12–14]. The operating injection pressures of such injectors are in a range of 7 MPa–15 MPa with SMD (Sauter mean diameter) distribution 15 μm–20 μm [9]. The advantages of swirl-type injectors are that they are suitable for both combustion modes of GDI engines, homogeneous and stratified charge combustion. In homogeneous charge combustion mode near the WOT (wide open throttle) condition (i.e., larger fuel amount), the swirltype injector produces a wider spray angle because of low ambient pressure during the intake stroke, which suppresses smoke generation. On the contrary, the spray from a swirl type injector collapses and exhibits shorter spray tip penetration at high ambient pressure during the latter half of the compression stroke, which is compatible with stratified charge combustion mode. Generally, swirl-type injectors are considered first generation injectors. Zhao et al. [15] reviewed research on the mixture dynamics and combustion characteristics of GDI engines, mainly equipped with swirl-type injectors in 1999. However, in the early 2000s, multi-hole nozzle injectors began to be adopted in new GDI engines. As reported in previous studies, the spray from swirl-type injectors is sensitive to ambient conditions (e.g., easy collapse at high ambient pressure), as shown in Fig. 2 [16], which can sometimes cause unstable engine operation. On the contrary, multi-hole injectors can produce stable mixtures that are less sensitive to engine operating conditions, as described in several studies [17–20].

1.2. Major features and recent trends in GDI engines Depending on the research and technology focus, researchers have 2

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European Emission Regulation Emissions standard

Effective Date

NOx CO HC (g/km) (g/km) (g/km)

NMHC (g/km)

PM PN (mg/km) (#/km)

Euro 6b

Sep-14

0.06

1.0

0.1

0.068

4.5

6x10 12

Euro 6c

Sep-17

0.06

1.0

0.1

0.068

4.5

6x10 11

LEV III ULEV50 Emission category NMOG+NOx (g/mi)

CO (g/mi)

HCHO (mg/mi)

PM (mg/mi)

0.05

1.7

4.0

3.0

(a) Emission regulation of EU and US

(b) Future European Fleet CO2 emission targets Fig. 1. Exhaust gas regulations and future fuel economy target [1,2].

ignition (DISI) gasoline systems as GDI systems.

1.2.1. Combustion modes of GDI engines Generally, GDI engines are designed for two different combustion modes: stratified lean combustion and homogeneous charge stoichiometric combustion. In stratified lean combustion mode, the GDI system uses central or side fuel injection to directly inject high-pressure fuel into the combustion chamber, where a stratified air/fuel mixture forms around the spark plug. In middle of the 1990s, the first generation of GDI engines adopted wall- or air-guided type stratified charge lean combustion and exhibited 20%–30% increases in fuel economy because of improved volumetric efficiency and reduced pumping loss. Despite the superior fuel economy of a GDI engine operating in stratified charge lean combustion mode, the conversion efficiency of the conventional TWC available using that mode is low [15,21,22]. In addition, aftertreatments to reduce NOx (such as lean NOx trap and selective catalytic reduction) were uncommon and expensive in the middle of the 1990s. Thus, current GDI engines adopt the homogenous charge stoichiometric burn mode, in which fuel is injected during the intake stroke to generate a homogeneous mixture. Without the benefits of lean combustion, these GDI engines retain significant benefits in fuel economy, mainly because the cooling effect of the vaporizing fuel allows a higher compression ratio than in conventional PFI engines. Table 1 lists selected specifications of GDI engines developed by

Fig. 2. Spray development process of swirl-type GDI injector. Sprays from swirltype GDI injectors are collapsed at high ambient pressure [16].

used different names to describe GDI fuel systems, including fuel stratified injection (FSI), injection direct essence (IDE), high-pressure direct injection (HPDI), and stratified charged gasoline injection (SCGI), in addition to GDI. In this article, we refer to direct-injection spark3

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Table 1 Engine and injection system analysis of recently developed GDI vehicles [23–40]. Manufacturer

Displacement (L)

Compression ratio

Injector mount/pressure

Special characteristics

Emission control

Audi

1.984

9.3

1.499

11.0

Chrysler Daimler

3.239 1.991

10.7 8.6

Ford

1.499

10.0

Dual high and low pressure fuel injection Audi Valve lift System Piezo injectors Valvetronic variable valve lift system Variable intake and exhaust valve timing Piezo injectors Variable intake and exhaust valve timing Water-cooled intercooler (IVLC), Direct injection,

3-Way catalyst converters

BMW

Side mount Up to 15 MPa Central mount Up to 20 MPa Side mount Central mount Up to 20 MPa Side mount Up to 15 MPa

3-Way catalyst converters 3-Way catalyst converters & Lean NOx trap 3-Way catalyst converters

GM

2.457

11.3

idling stop Intake valve lift control (IVLC)

3-way catalyst converters

Honda

1.993

13.0

Hyundai

1998

9.5

Toyota

2.949

13.0

Central Up to 20 MPa Side mount Up to 20 MPa

Atkinson cycle Cool EGR VTEC Variable intake and exhaust valve timing Cooled EGR Atkinson cycle Cool EGR Thermal Efficiency 38.5%

Side mount Up to 15 MPa Side mount Up to 20 MPa

3-Way catalyst converters

3-Way catalyst converters

3-Way catalyst converters 3-Way catalyst converters

thermal dissociation improves the thermal efficiency. The thermal efficiency of a GDI engine can also be increased by a variable valve timing (VVT), variable valve lift (VVL), and variable length intake manifold (VLIM). Precise control of the injection timing and air-fuel ratio using an engine control unit (ECU) can also increase the thermal efficiency by promoting ultra-lean combustion. The GDI combustion system improves the transient behavior of the engine and reduces HC formation during cold start by reducing liquid film formation, which enables more precise control of the air-fuel ratio [42,43]. A GDI engine has the advantage of low fuel consumption enabled by lean combustion, although it is more difficult to reduce its exhaust emissions using a TWC compared with a PFI engine because a GDI engine mainly operates under lean mixture combustions at low engine speed and low load conditions. In addition, the lean combustion conditions, as in a diesel engine, results in the formation of a large amount of NOx and particulate matter (P.M.) The formation of CO depends on the air-fuel ratio, increasing with the richness of the mixture. Because a GDI engine operates under lean combustion conditions, CO emissions are very low. However, HC emissions might increase because the direct fuel injection trends to produce a liquid film on the surface of the piston head. HC formation is a function of the temperature, and it could increase under cold start conditions; it depends significantly on the incylinder temperature distribution, in-cylinder air flow, fuel evaporation, and fuel-air mixture. CO and HC formation is actually not a major problem in GDI engines. However, NOx cannot be entirely converted into clean gas [44] under lean mixture operating conditions. A GDI

major manufacturers [23–40]. Current GDI engines use central or side mount injection with a maximum injection pressure from 15 MPa to 20 MPa. The advantages and disadvantages of different injector mount type (i.e., side- or central-mount) will be discussed in Section 2. All engines in listed in Table 1 are equipped with 3-way catalytic converters to control emissions of HC, CO, and NOx. In particular, the GDI engine from Daimler has a lean NOx trap along with a 3-way catalytic converter because the engine operates in lean combustion mode for low load conditions [34]. Thus, interest in the lean combustion mode for GDI engines is increasing even though most current GDI engines operates in stoichiometric combustion mode. Recent research on the combustion modes of GDI engines will be reviewed in Section 3. 1.2.2. Benefits of GDI engine over conventional PFI engine Compared with PFI gasoline engines, injecting the fuel directly into the combustion chamber in a GDI engine reduces the average in-cylinder temperature during the evaporation of the atomized fuel droplets. If the fuel is injected during the intake stroke, the volumetric efficiency is improved to 10% [41]. The reduction of the in-cylinder temperature reduces the possibility of knocking, and affords a higher compression ratio [41]. In addition, operating GDI engines in stratified combustion mode enables ultra-lean combustion without misfire because the fuel injected at the end of the compression stroke is combusted in the cavity of the combustion chamber. In stratified combustion mode, reducing the combustion that occurs on the cylinder wall raises the specific heat ratio, and the accompanying reduction of

Fig. 3. Schematic descriptions of three possible stratified-charge combustion concepts [41,44,46]. 4

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layout of a gasoline combustion system including the injector should be optimized to reduce PN by means such as adjusted spray targeting, laser drilling of the spray holes, and flexible hole design. Research on pollutant emissions from GDI engines will be reviewed in Section 3.

engine thus requires exhaust gas after-treatment measures such as the use of a lean NOx trap (LNT) and selective catalytic reduction (SCR). In addition, the exhaust of a GDI engine contains more P.M. than that of a PFI engine [44,45] because of the rich mixture that forms around the spark plug. The major difference between a GDI engine and a diesel engine is that the ignition of the charge in the latter occurs automatically by compression, whereas a GDI engine requires a spark plug for ignition. In a GDI engine, it is therefore necessary to concentrate the ignitable mixture in the vicinity of the spark plug in the combustion chamber, and that requires special airflow such as swirl or tumble to transport the injected fuel to the spark plug. The swirl or tumble flow can be achieved by any of three methods, wall-guided, air-guided, or spray-guided systems, as shown in Fig. 3 [41,44,46]. We present a detailed discussion of these methods including their spray and combustion characteristics, in Section 2.

2. Spray and atomization characteristics Compared to a PFI engine, a GDI engine in which the fuel is injected directly into the cylinder has many advantages and exhibits higher efficiency and performance. One primary aspect of a GDI engine is the effectiveness of the mixture formation and precise control of the fuel delivery process including the injection quantity and timing [15,52]. The achievement of high fuel economy and low pollutant emission depends largely on optimizing the mixture formation before ignition. For a direct injection system, including both diesel and GDI engines, mixture formation is an important influence on the combustion and emissions characteristics. Delphi, a major manufacturer, presented four requirements for a GDI fuel supply system [53]:

1.2.3. Recent GDI injectors The injector is one of the most important component of a GDI engine system. Multi-hole injectors were recently developed from single-hole swirl-type injectors. As mentioned above, GDI injector manufacturers are working to improve the combustion efficiency and reduce exhaust emissions through the use of multi-hole injectors, high injection pressure, and precise targeting and timing. Specifications of the injectors most widely used in GDI engines are listed in Table 2 [47–49]. All of the listed injectors have multi-hole type nozzles with 4–12 holes. The maximum operating pressures are from 200 bar to 300 bar, and the maximum cone angles are around 100 degrees. We discuss development trends for GDI injectors in detail in Section 2.

1) Precise metering – The GDI injector should deliver a precise fuel mass to the engine. With the increase of injection pressure in a GDI injector, the pulse width becomes too short to supply the required amount of fuel. When the pulse width is shortened, the part-to-part flow variation increases. 2) Precise targeting – The GDI injector should exhibit consistent spray characteristics; i.e., variations in the spray plume angle and plume targeting should be minimized to increase the precision of the spray penetration and spray cone angle. The reduced deposit build-up that results improves the performance of the injector over its lifetime. 3) Fine atomization – The injection pressure of a GDI injector should be high to ensure the production of finely atomized droplets. From a thermodynamic viewpoint, high injection pressure increases combustion efficiency through improved air utilization. From the viewpoint of emissions, high injection pressure reduces HC emission during cold start by improving the mixing and reducing wall wetting. In addition, the mass and number of P.M. can be reduced by increasing the injection pressure. 4) Accurate timing – Precise timing control is required when using multiple injections to reduce particulate emissions by improving the overall homogeneity of the mixture in the chamber and the dilution tolerance through local enrichment and higher turbulence at the spark plug, which also enhances flame initiation and early propagation and penetration control to minimize wall wetting and particulate emission.

1.2.4. Emissions regulation of gasoline engines Automobile engines should meet emission regulations. Many researchers have suggested new technologies to meet emission regulation for GDI engines. Thus, we briefly introduce the history, current, and predicted future emissions regulations for GDI engines in this section. As mentioned above, exhaust gas regulations have become steadily more stringent since 1992, when exhaust emissions legislation was first introduced in Europe, Japan, and the US. The Euro-6 legislation stipulates a maximum of 1.0 g/km and 0.1 g/km of CO and HC emissions, respectively. The regulation of NOx was made more stringent by reducing the maximum emission from 0.15 g/km (Euro-3) to 0.06 g/km (Euro-6b). The P.M. rule is under the Euro-5a (2009), and Euro-6b (September 2014) regulations stipulates a maximum emission of 4.5 mg/km. Euro-6b particularly controls the particulate number (PN) and particulate mass (PM). The Euro-6b stipulation of a PN of 6 × 1012 #/km may be applied to Euro-6 GDI vehicles upon the request of the manufacturer until three years after the date for type approval. The PN regulation will be made more stringent by reduction to 6 × 1011 #/km, which will be the standard emission allowance for diesel engine vehicles in Euro-6c (September 2017). Conversely, the US has two emissions regulations, the Tier standard of the US Environmental Protection Agency (US-EPA) and the low-emission vehicle (LEV) standard of the state of California. Currently, the US government controls PM (particle mass) based only on the Tier 2 standard. From 2017, the US will control PM based on the Tier 3 standard, which has a more stringent limit of 0.006 g/mile (70% reduction compared to the Tier 2 limit). In the LEV standard, LEV 3, which will take effect in 2017, stipulates a PN (particle number) limit of 3.0 × 1012 #/mile under the FTP-75 mode. P.M., which includes soot, is mainly formed by insufficient mixing of air and fuel. It is formed in diesel and GDI combustion systems, which use direct injection, because of the short mixing time available to form a homogenous mixture. To satisfy the stringent emission regulations in Europe and the US, GDI-engine vehicles could use various emission after-treatment technologies such as gasoline particle filter (GPF) [46], a combination of TWC and SCR [50], or close-coupled LNT (CCLNT) [51]. In addition, the hardware and

In GDI engine systems, the time available for mixture formation through droplet atomization and evaporation is shorter than in PFI engine system. In addition, the direct injection of fuel can form a liquid film or wall wetting on the piston head and cylinder wall. Wall wetting from the fuel spray of GDI injectors has been minimized by significantly Table 2 Comparison of GDI injectors recently produced by major manufacturers [47–49].

Product Type Number of hole Maximum operating pressure Working flow range @ 100 bar SMD Cone angle Min. dynamic flow

5

Bosch

Delphi

Continental

HDEV 5.2 Multi-hole 4–7 200 bar

Multec® 12.1 Multi-hole 12 300 bar

DI XL3.1 Multi-hole Max. 9 holes 250 bar

Up to 17.1 g/s N/A 110 deg. A/A

Up to 20 g/s

Up to 20 g/s

6 μm @300 bar 40–90 deg. 3.5 mg/pulse @ 100 bar

15 μm 35–90 deg. ~2.0 mg/pulse @ 200 bar

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test region. The light refracted by the fuel is refocused on the opposite side of the parabolic mirror and enters the sensor of the camera. The difference between Schlieren and shadowgraph techniques is that knifeedge is used in the Schlieren technique. When refocused by the second parabolic mirror, the knife-edge at the focal point cuts off the part of the light and determines the sensitivity of the system. Kook et al. [60] found that a light cutoff of approximately 60% was optimal for imaging the internal pattern of sprays. Li et al. [61] found that a better noiseremoval image and spray boundary could be obtained when the threshold was set as 2%. Fig. 5(a) and (b) illustrates vapor phase spray acquired with the Schlieren technique and liquid phase spray acquired with the mie scattering technique. [63] Comparison of spray images and penetration of different phase is used to analyze the fuel evaporation characteristics.

optimizing combustion chamber airflow and injector spray patterns. The use of multi-hole injectors with pulsed sprays has also significantly reduced in wall wetting [54]. To get good combustion performance in GDI engines, a stable and ignitable mixture should form around the spark plug. A good design for the fuel delivery system, including a high pressure injector, is essential to establishing evaporation and atomization in a relatively short mixture preparation time. As stated above, recent technology trends for direct injection gasoline engines are moving toward higher injection pressure (Fig. 4 [53]) and more sophisticated injection strategies, including double- and split-injections. Therefore, the combustion chamber geometry and a high pressure injector is important, and the overall spray characteristics such as spray pattern, shape, spray tip penetration, spray cone angle, droplet size, and velocity should be comprehensively understood. In this chapter, we review the spray and atomization characteristics of GDI injector, especially, multi-hole injectors. The purpose of this chapter is to provide comprehensive information about the spray and atomization characteristics of multi-hole GDI injectors, and state-of-the-art technologies and research trends for multi-hole GDI injectors. In addition, this chapter includes a brief comparison between swirl type injectors and multi-hole injectors.

2.2. Swirl type GDI injector First generation GDI injectors have a swirl type, and satisfy the requirement to reduce specific fuel consumption and improve pollutant emissions using techniques such as fuel spray with small droplet size, wide-spreading spray, and low-penetration at high-load conditions, while maintaining a well-atomized compact spray and stratified charge in the low-load condition [70]. The typical characteristics of a swirl type GDI injector are the vortex-like structure at the spray edge, the hollow cone spray shape, and the initial spray slug. After injection, a rotating vortex-like structure appears over time at the spray tip region. The vortex structure appears because of the air motion caused by the local pressure difference and the increase of spray resistance to the ambient air with the increase of spray width [70]. The hollow cone spray shape that emerges from a swirl type GDI injector is caused by the centrifugal force of the angular momentum at the nozzle exit [16,71], and it forces most droplets to concentrate on the outer edge. In addition, the swirl motion of the injected spray makes an air core because the air is sucked into the middle of the liquid [70], which then forms the hollow cone spray shape. The swirl type GDI injector produces an initial spray slug because of the dead sac volume between the tangential slot and the nozzle needle in the injector. This initial spray slug consists of large droplets with high axial momentum but without angular momentum. Therefore, it is main source of unburned hydrocarbon. Many researchers have examined the effects of the injection and ambient pressures on the spray characteristics [15,16,52,70–72], and found that the injection pressure significantly influences the spray structure. Increased injection pressure enhances the rotating air motion and the vortex-like structure by the increase of the injection velocity, as shown in Fig. 6 [70]. The effect of injection pressure on spray tip penetration and spray cone angle is insignificant because the increase in

2.1. Vapor phase analysis Mie scattering [55,56] is primarily applied to record images of liquid-phase sprays. Since gasoline fuels have a high evaporation rate, analysis of the vapor phase fuel is also an important factor in analyzing the spray and mixture formation. Images of vapor-phase sprays are usually recorded via the technique using reactivity with specific wavelength (PLIF, LAS) [55–58] and the technique using density difference (Schlieren, shadowgraphy, BOS) [59–67]. Planar laser-induced fluorescence (PLIF) technique measure the fluorescence signal from the tracer mixed with single-component fuel. Heechan et al. [55] used 3-pentanone as fuel tracer in isooctane that is a surrogate fuel of gasoline. Nd-YAG laser provided the incident light laser beam that react with fuel tracer to produce fluorescence signal. Fuel tracers are present in both the vapor-phase and the liquid-phase, so fluorescence is captured in the entire spray area and it is hard to clarify the evaporation process. Laser absorption-scattering (LAS) [57,58,68] is a technique of inserting beams of two different wavelengths and taking images after absorption and scattering by vapor-phase and liquid-phase fuel to obtain the distribution of liquid and vapor phase. Li et al. [68] measured the extinction of ultraviolet light with a wavelength of 266 nm and a visible light with a wavelength of 532 nm by the absorption of the vapor-phase and the scattering and absorption of the liquid-phase. LAS is a technique that can simultaneously capture the liquid and vapor behavior in the same spray. However, for the LAS optical technique, the fuels must be replaced with test fuels that can absorb lasers of different wavelengths, because neither gasoline nor ethanol can absorb lasers strongly. Additionally, the LAS technique requires complex experimental equipment. Background oriented schlieren (BOS) technique enables gas-phase gasoline fuel photography without using surrogate fuel [64,65]. Lee et al. [65] placed patterned paper between the spray and light source and captured refracted image by fuel injection. The displacement vector can be derived from subtle shift from background and contour plot of vector means density gradient. The Schlieren/shadowgraph visualization technique is the most effective method for recording the vapor-phase spray owing to its simple experimental setup and testing procedure [67,69]. Kook et al. [60] and Li et al. [61] presented details of the shadowgraph/Schlieren system setup and image-processing method. There are several types of Schlieren/shadowgraph system and Z-type two-mirror system is most commonly used in spray study. The light from the light source is reflected on the parabolic mirror and propagates in parallel through the

Fig. 4. The increasing trend of injection pressure in gasoline fuel injector [53]. 6

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(a) Vapor plume evolution (Schlieren)

t=0.0ms

t=0.2ms

t=0.4ms

t=0.6ms

t=0.8ms

t=1.0ms

t=1.2ms

t=1.4ms

(b) Liquid plume evolution (Mie-scattering)

t=0.0ms

t=0.2ms

t=0.4ms

t=0.6ms

t=0.8ms

t=1.0ms

t=1.2ms

t=1.4ms

Fig. 5. Vapor and liquid plume evolution (Injection pressure: 9 MPa, Injection duration: 4 ms, multi-hole injector) [63].

strategy with various hole arrangements. In a GDI engine, the development, shape, and patterns of the spray, including the tip penetration and cone angle, are important factors that should be optimized and adapted to the engine geometry for better delivery of fuel into the cylinder. In a GDI engine with a multi-hole injector, the optimization of the spray geometry and its characteristics is especially important because of the flexibility of the plume targeting to prevent fuel impingement on the piston surface, which is a main source of hydrocarbon and smoke emissions [74]. Therefore, various studies have examined the spray development process and spray geometry for multi-hole injectors under various injection conditions. One widely used method for investigating spray geometry involves the use of the Mie-scattering visualization technique and a high-speed charge-coupled device (CCD) camera. Hoffmann et al. [75] used a highspeed camera to investigate the effect of injection pressures as high as 40 MPa on the spray structure and development processes. To compare the results for injection pressures of 5 MPa, 10 MPa, 15 MPa, and 20 MPa, they fixed the injection quantity at 11 mg/pulse and adjusted

pressure energy induces an increase in the rotating momentum and enhances atomization (Fig. 7). An increase in ambient pressure causes the spray width to decrease because of the increase of ambient density and the drop-drag force, which make the hollow-cone spray into a solidcone spray shape. In addition, the increase of ambient pressure decreases the pressure difference at the nozzle exit, which thus decreases the injection velocity and the swirl motion. 2.3. Multi-hole GDI injector 2.3.1. Spray shape and development process Although the swirl type GDI injector has good atomization performance, it is very sensitive to engine operating and thermodynamic conditions. Therefore, manufacturers have recently developed and investigated a multi-hole GDI injector, which has a relatively high spray stability at high chamber pressure, and enforced air entrainment characteristics because of its enlarged spray surface area [72,73]. In addition, multi-hole GDI injectors permit adjustment the injection 7

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detail for an injection pressure of 20 MPa under ambient conditions. Using the analogy of a diesel engine [78], they divided the gasoline spray development behavior into two parts as shown in Fig. 10. The first part is the primary area characterized by low standard deviations and linear penetration progression. In this area, the spray seemed to be very stable and was thus not significantly affected by the in-cylinder air motion caused by increased velocity and momentum. The second part is the secondary spray area, which is an approximate function of the square root of the time. In this area, the spray is significantly affected by the increased air interaction, resulting in a larger standard deviation of the macroscopic spray parameter. Based on the results of an experiment in which they used phase Doppler anemometry (PDA) to measure the droplet velocity, they extensively discussed the macroscopic recirculation phenomenon, which is usually detected at the edge of the leading spray. They concluded that the overtaking phenomenon of the droplet velocity between the front spray and the main spray could lead to the formation of shear zones that could affect the macroscopic recirculation area. Oh et al. [79] used the modified KIVA-3V code to predict the effects of the ambient and injection pressures on the process of spray development generated by a conventional six-hole GDI injector, and validated their finding using the Mie scattering visualization technique. They reported that an increase in the ambient pressure significantly reduced the spray shape and formed a locally concentrated fuel region. They reported that increasing the injection pressure to 20 MPa led to the wide dispersion of the fuel droplets and increased the vortex intensity. Parrish evaluated the effect of ambient conditions on the formation of liquid and vapor envelopes under engine-like conditions [80]. Using a combination of high-speed Schlieren and Mie scattering imaging, he captured the liquid and vapor phases of the spray produced by a multi-hole injector. The measurements showed that a high ambient temperature reduced liquid penetration and shortened the residence time by enhancing vaporization. It was found that the elevated ambient pressure reduced the liquid penetration but did not significantly reduce the residence time. Marchitto et al. [81] used X-ray tomography visualization and a CCD camera to investigate the internal structure of the dense sprays injected by a six-hole injector to obtain a better understanding of the spray geometry and quantitatively determine the fuel mass density in the near nozzle region. They obtained three dimensional computed tomography (3D CT) images of the GDI sprays in a region very close to the nozzle exit, and reconstructed the six-jet

Pinj Time after injection

2.0MPa

3.6MPa

4.6MPa

5.0MPa

0.8ms

1.6ms

2.4ms Fig. 6. Spray images at various injection pressures under room temperature (Injector: Swirl-type GDI injector, Injection pressure: 2.0 MPa, 3.6 MPa, 4.6 MPa, and 5.0 MPa, Ambient temperature: 300 K, Ambient pressure: 0.1 MPa) [70].

Fig. 7. Effects of injection pressure on Sauter mean diameter (Injector: Swirltype GDI injector, Injection pressure: 5 MPa and 7 MPa, Ambient pressure: 0.1 MPa, Injection duration: 1.0 ms) [16].

the pulse duration to deliver the same quantity. Based on their measurements, the final plume penetration length, which can also be called as the fully developed stage of the spray, was almost independent of the injection pressure, remaining at almost the same value as shown in Fig. 8. They also reported that the integral spray momentum, which enhances air entrainment in the spray, was proportional to the square root of the fuel pressure. To compare the spray areas and cone angles for injection pressure of 5 MPa and 10 MPa, Mittal and Schock [76] used the Mie-scattering visualization method to obtain spray development images for a multihole injector. They found that an increase in the injection pressure significantly increased the spray area but not the spray cone angle. Mitroglou et al. [72] reported similar results. Using Mie scattering spray visualization and a CCD camera, they quantified the effects of injection pressures of up to 20 MPa and ambient pressure of up to 1.2 MPa on spray development in a constant-volume high-pressure chamber. They found that the spray cone angle from a multi-hole injector was independent of the injection and ambient pressures, and they attributed that to the effect of the configuration of the holes. They also demonstrated the process of air entrainment in individual jets, where pairs of counter-rotating vortices were attached to the individual spray jets in the fully developed spray structure, which they called a “fishbone structure” (Fig. 9). In a detailed analysis of the characteristics of the spray produced by a multi-hole injector, such as spray propagation behavior, Kramer and Kull [77] used the shadow-graph imaging technique to perform an optical investigation. They discussed spray propagation behavior in

Fig. 8. Mie spray images for fuel system pressures of 5, 10, 20, and 40 MPa (quantity = 11 mg/pulse, pulse width adjusted in accordance with fuel system pressure) [75]. 8

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2.3.2. Spray atomization The formation of a good mixture largely depends on the spray atomization, the droplet distribution, and the mixing of the air and fuel because of the shorter time available for mixture preparation in a GDI engine than in a conventional PFI engine. A comprehensive understanding of the spray atomization characteristics is a key requirement for improving combustion efficiency. Laser-based measurement techniques such as laser diffraction and PDA are commonly used to characterize spray atomization because they are non-intrusive and do not require direct contact between the working fluid and the measurement element [83]. In addition, because of the very small volume required for the measurement, those methods can achieve high resolutions of the spatial and temporal droplet distributions [83]. To characterize the structure of the unsteady spray produced by one hole of a multi-hole injector in terms of the temporal and spatial distributions of the droplets, Ahmed et al. [84] performed high-density LDA/PDA experiments under an injection pressure of 7 MPa and atmospheric conditions. To obtain substantial information about the characteristics of the spray atomization in the nozzle region, they spaced the measurement points by 0.5 mm in the radial direction and 2 mm in the axial direction as shown in Fig. 11. They reported that atomization near the nozzle occurred predominantly at the center of the spray. They also observed a high droplet velocity at the center of the spray region. By size-classified analysis, they found that droplets with diameters of less than 15 μm exhibited the ability to follow the airflow, whereas those with diameters greater than 20 μm experienced drag, resulting in momentum decay. They also concluded that droplets with diameters between 15 μm and 20 μm determined the follow/penetration near the nozzle. Heldmann et al. [85] performed PDA experiments to examine the effects of the ambient temperature on the spray atomization processes. Their results revealed that an increase in the ambient temperature reduced the kinematic energy of the fuel, resulting in a reduction in the droplet size. They also characterized the effect of the physical properties of the fuel on the spray atomization using three different fuels, gasoline RON 95, n-hexane, and n-decane. They found that the high viscosities of typical biofuels such as ethanol and butane negatively affected spray atomization, and that the higher latent heat of biofuels also impaired the evaporation process because of increased evaporative cooling. Many researchers have tried to enhance the atomization performance of multi-hole GDI injectors by methods such as increasing the injection pressure and modifying the nozzle shape and hole configuration. To verify whether the injection pressure significantly reduced the droplet diameter, Hoffmann et al. [75] compared the droplet size parameters Dv10, D32 (Sauter mean diameter, SMD), and Dv90 for injection pressures ranging between 5 MPa and 40 MPa. Fig. 12 shows the significant effect of fuel pressure on spray atomization. They found

Fig. 9. Typical air entrainment in an injector spray produced by 5 + 1 central holes [76].

Fig. 10. Area specifications for gasoline spray propagation [77].

2mm

structure. This tomographic reconstruction technique enables very clear determination of the spray shape as well as the propagation direction. Using X-ray absorption analysis and a pseudo-color scale ranging from blue for low absorption intensity to red for maximum values, they also represented the fuel mass densities in the direction of the spray propagation and across the spray cross section. Based on the various measurements, they reported that the fuel supplies through the different holes propagated independently and that the fuel density peaked at the nozzle exit and sharply decreased in the direction of the spray propagation. In addition, based on the observation of slight asymmetries in the longitudinal direction, they concluded that interaction phenomena between the fuel and ambient-air were activated in the downstream region. Moon et al. [82] investigated the effect of hole length and number on the initial spray formation using the X-ray phase contrast imaging technique. They used four prototype multi-hole GDI injectors with different hole lengths and hole numbers to investigate jet breakup, spray dynamics, and droplet formation using single- and double-expose X-ray phase contrast imaging. They reported that a decrease in the hole length caused an increase in the axial- and radialvelocity along with an increased void fraction in the orifice, and improved jet breakup performance. In addition, they revealed that the jet flows at the nozzle exit are highly turbulent and perturbed. They observed that increasing the hole number decreased the axial- and radialflow velocity by slow breakup.

Injector Radial, r, mm Axial, z, mm

0.5mm

Fig. 11. Locations of measurement points [84]. 9

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air entrainment, cross-flow velocity, and turbulence. To do that, they used simulations (CFD) and experiments, which included the use of laser diagnostics such as laser-induced fluorescence (LIF), Mie-scattering, and PDA. Their results revealed that a symmetrical nozzle hole configuration tended to produce isolated clouds of fuel vapor, whereas an asymmetrical configuration produced coherency in the vapor cloud. However, they suggested that the distance between the spray plumes was more important than whether the nozzle holes were symmetrical or asymmetrical because a shorter spray plume distance easily increased the interaction between the plumes, thereby affecting the fuel vapor cloud coherency, cross-flow velocities, turbulence, and fuel/air ratio.

2.4. Comparison of swirl-type and multi-hole-type GDI injector As reviewed in Sections 2.1 and 2.2, GDI injectors can be swirl-type or multi-hole type. Fig. 15 shows a comparison of the spray patterns produced by the multi-hole (left) and swirl type (right) GDI injectors. Each nozzle of a multi-hole injector produces six different spray plumes. Six-hole injectors are usually preferred [89–94], but eight or more holes can also be used [95–98]. The spray pattern of a multi-hole injector can be conveniently changed by simply changing the arrangement of the holes and without changing the position of the injector mount [91,99,100]. Conversely, the swirl type injectors form a hollow cone shaped spray that is separated into many thin strings [101]. These strings break up into small droplets with further propagation [102]. Preussner et al. [12] evaluated the performance of each injector type. They found that the multi-hole type could produce a great variety of spray shapes, but the atomization quality was unsatisfactory for an injection pressure of 10 MPa. In contrast, the inward-opening type exhibited the best performance with regard to atomization quality, flexibility of the spray pattern, and other parameters. However, more recently, the perspective on multi-hole GDI injector has changed because of their improved injection pressure (up to 35 MPa), precise injection control, and the importance of spray targeting for the stratified combustion strategy. From the viewpoint of fuel film, a multi-hole injector exhibits much better characteristics than a swirl type injector. Drake et al. [91] compared the fuel film on the piston heads of multihole and swirl injectors in similar operating conditions. The endoscope images of the fuel spray and pool fire, and the thickness of the wall film on the piston for each injector are shown in Fig. 16. The upper and lower rows respectively show the results for the swirl and multi-hole injectors. Although the authors insisted that the multihole spray was clearly more dispersed than the swirl spray, the figure shows no distinct differences in the spray patterns from both injectors. Both sprays are directed toward the edge of the piston bowl, and as the

Fig. 12. Effect of fuel pressure on the spray droplet-size distribution parameters Dv10, D32, and Dv90 [53].

that the effect of injection pressure on spray atomization decreased at fuel pressures higher than 20 MPa, but they did not attain an asymptotic condition until 40 MPa. Mitroglou et al. [72] reported that the effect of injection pressure on droplet diameter was smaller than that of the ambient pressure. They quantified the effects of injection and chamber pressures of up to 20 MPa and 1.2 MPa, respectively, on droplet velocity and size as determined by PDA. They found that the droplet arithmetic mean diameter (AMD) in the main spray was on the order of 15.5 μm and 13.0 μm for injection pressure of 12 MPa and 20 MPa, respectively, under atmospheric conditions. They reported that the effect of the injection pressure on the droplet size was rather small compared to that of an increase in the ambient pressure from atmospheric pressure to 1.2 MPa, which resulted in a much lower droplet velocity and up to a 40% increase in the droplet size. Lee and Park [86] reported similar results. To investigate the effect of injection pressure on the atomization process of GDI sprays, they performed a 2D phase Doppler particle analyzer (PDPA) experiment with a focus of a particular one of the six jets. The obtained droplet information such as the droplet velocity and size distribution, 50 mm from the nozzle tip and reported a limit to enhancing spray atomization by increasing the injection pressure. An increase of the injection pressure from 5 MPa to 10 MPa and then to 20 MPa decreased the SMD linearly by approximately 10 μm, but further increase of the injection pressure beyond 20 MPa produced no significant decrease in the SMD, as shown in Figs. 13 and 14. Another method to improve the atomization characteristics is to innovate the hole shape and hole configuration of a multi-hole injector. To investigate the effect of nozzle configuration on spray atomization, Lee and Park [87] performed PDPA experiments using an adjustable GDI injector that could be adapted to different types of nozzles. One of the nozzles they used was a single-hole nozzle with design features representative of current gasoline multi-hole injectors. The other was a group-hole nozzle with three orifices equally spaced around a center hole at radial distance of 0.260 mm. The area of each orifice was fixed to ensure that equal quantities of fuel were injected through each. Their results revealed that the average size of the droplets produced by the group-hole nozzle was less than that of the droplets produced by the single-hole nozzle by approximately 2 μm. Based on a comparison of the spatial distribution of the droplet size within a region of the spray propagation with that in the surrounding area, it was also reported that the group-hole nozzle had a stronger effect on the rapid dispersion of the droplets owing to active air entrainment. Dahlander and Lindgren [88] investigated the effects of the configuration of the nozzle holes in a multi-hole injector on spray formation and spray-induced air movement. Using four nozzles with different hole configurations and umbrella angles, they conducted a detailed investigation of the effects of the configuration on the fuel distribution,

Fig. 13. Comparison of droplet diameter profiles for injection pressures of 5, 10, 20 and 30 MPa [86]. 10

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piston moves up, the fuel spray gathers near the spark plug. However, the amount the fuel film generated and the consequent combustion procedure do differ between the injectors. A dramatic difference can be observed in the figure between the pool fire luminosities and the images of the piston wall films from the two injectors. As the premixed flame develops, the fuel film on the piston burns out and radiates a yellow light in the pool fire. Although the multi-hole spray also forms a fuel film that burns in the pool fire, it is smaller and burns out within a shorter time. In contrast, the pool fire of the swirl injector burns for a long period until the fuel film completely vaporizes. As shown in the second images, the swirl injector produces a thicker and bigger fuel film. Because the fuel film is a major cause of soot formation, switching from a swirl injector to a multi-hole injector improves the combustion and reduces emissions; the fuel film mass is reduced by an order of magnitude and the soot emission becomes nearly zero through reduction of the pool fire. The emission of unburned hydrocarbons also decreases significantly when incomplete combustion decreases [91,103]. Fig. 14. Comparison of droplet distributions for injection pressures of 5, 10, 20, and 30 MPa [86].

2.5. GDI injection system for in-cylinder flow The major difference between a GDI engine and a diesel engine is that the ignition of the charge in the latter occurs automatically by compression, whereas a GDI engine requires a spark plug for ignition, which requires special airflow to transport the injected fuel to the spark plug. The injection system for the formation of swirl or tumble flows in a GDI engine system can be categorized into three groups, spray-guided, wall-guided, and air-guided systems, depending on the mechanism of the mixture formation, as shown in Fig. 3 [12,41,44]. However, they can be more roughly categorized into two groups depending on the position of the injector mounting, side mount (wall- and air-guided) and center mount (spray-guided) [104]. The wall-guided system [12,41,44] uses the geometry of a special piston head to direct the injected fuel to the spark plug, which is installed at the center of the combustion cylinder. The interaction between the fuel spray and the surface of the proper cavity on the top of the piston initiates the stratification process. This system is easier to produce than air-guided and spray-guided systems. However, in a wallguided system, it is difficult to sufficiently evaporate the fuel before ignition, and some of the injected fuel thus produces a liquid film on the piston surface. This method therefore has some disadvantages, such as

Fig. 15. Spray patterns produced by multi-hole injector (left) [90] and outwardly opening (swirl) injector (right) [104].

Fig. 16. Endoscopic images of fuel spray at BTDC 61° CA (left); fuel film on piston head (center); and pool fire due to fuel film. Upper row: Swirl injector. Lower row: Multi-hole injector [91]. 11

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interaction between the fuel spray and the piston surface, thereby reducing fuel film formation [38,107,108]. Drake et al. [91,98] studied the fuel film deposited on the piston head by a wall-guided system in 2003 and that produced by a spray-guided system in 2005. In those works, the authors noted that less pool fire was observed in the sprayguided engine than in the wall-guided engine. Thus, the spray-guided engine performed better regarding both combustion efficiency and emissions.

higher HC and CO emissions. In addition, the possibility of reducing the fuel consumption is low. For those reasons, pure wall-guided systems are generally unimportant. The wall-guided system is mainly used in combination with the air-guided system. In an air-guided system [12,41,44], the fuel is injected into the airflow to transport the compact spray plume to the spark plug in the combustion chamber. Wall wetting does not occur in a pure air-guided system. Stable airflow facilitates the maintenance of a compact spray plume and its transportation to the spark plug. A homogenous air/fuel mixture forms inside the cloud, with in-cylinder airflow produced by the special shape of the inlet port. The interaction between the fuel spray and the airflow induces the stratification. In an air-guided combustion system, the injector is located far from the spark plug. The velocity of the air is controlled by air baffles, specially shaped inlet ports in the intake manifold. To best use the advantages of both wall- and air-guided systems, a combination system, the air-wall-guided system [44], can be used in GDI engines. In the combined system, the piston has two bowls, a fuel bowl on the intake side and an air bowl on the exhaust side. A specially shaped intake port produces tumble flow. The fuel is thus simultaneously guided by the air and fuel bowls to the spark plug. We expect that stratified charge gasoline injection and air-wall-guided combustion will be the dominant systems of the future. In the spray-guided system [12,41,44], the spray and its dynamic characteristics affect the stratification process. The fuel is injected around the spark plug, where the fuel spray evaporates. The advantages of this system are a larger stratification region, reduced wall wetting, lower sensitivity to the in-cylinder air-flow and variations between cylinders, and reduced HC emissions. The stratified-charged gasoline injection system with piezo-injection is a type of spray-guided combustion system. It operates with high excess air, which results in excellent fuel efficiency. Multiple injections extend the lean-burn operation mode to higher engine speeds and larger loads. During each compression stroke, a series of injections spaced just a fraction of a second apart enables better mixing and combustion and lower fuel consumption. The spray-guided system can therefore be used to achieve higher fuel efficiency than other systems. Although various studies have been conducted to determine the optimal injection system, which one is best remains debatable. In 1998, Preussner et al. [12] emphasized the limitations of the spray-guided-type: limited mixture formation time because of the close arrangement of the injector and spark plug. In such systems, the fuel is injected directly toward the spark plug, and the mixture formation procedure depends on the spray characteristics. That leads to fuel impingement on the spark plug, with a resultant shortening of the lifetime of the electrode materials. In contrast, in the wall-guided-type, the injector and spark plug are installed farther apart, and the fuel spray is driven toward the spark plug by interaction with a well-defined piston. That system extends the mixture formation time and allows combustion to be less sensitive to the spray. On the other hand, in 2005, Drake et al. [98] reported that the wallguided system was almost obsolete and that the next generation of GDI engines would focus on the spray-guided system. They argued that firstgeneration GDI engines with the wall-guided design had poor injector performance and that the spray-guided design was vastly superior to the wall-guided design, offering improved fuel efficiency, substantially lower emissions, a wider range of stratified combustion, and less fuel film [98,104–106]. The thermodynamic efficiencies and HC emissions of the wall- and spray-guided systems are compared in Fig. 17. In an ideal cycle, wall- and spray-guided systems exhibit similar thermodynamic efficiencies. However, losses caused by incomplete combustion in a wall-guided system are significantly greater than those in a spray-guided system. This trend can be identified in Fig. 17(b), which shows the HC emission graphs. The wall-guided system produces about three times as many HCs, which implies severe incomplete combustion. The spray-guided system also minimizes undesired

3. Combustion and emissions characteristics GDI engines generally have about 15% improvement of fuel efficiency than that of PFI engines [109,110]. The improved fuel efficiency results from increased volumetric efficiency and compression ratio. Fuel evaporation increases the volumetric efficiency by absorbing heat energy from the intake air [109,111]. Also, the higher compression ratio increases the power output of the engine. Directly injected fuel supplies the fuel with turbulence energy and affect combustion structure. With late injection around spark timing, turbulence due to fuel injection remain and accelerate the kernel growth and flame propagation [112]. However, shorten time for fuel evaporation and fuel/air mixture formation results in worse homogeneous mixture and increase of HC and particulate matter emission [112,113]. Thus, optimizing engine output performance parameters, such as low fuel consumption, high combustion performance, and low exhaust emissions, requires different engine operating modes, as shown in Fig. 18. A GDI engine should operate in the stratified mode at low engine load conditions and in the homogeneous mode at high load conditions. Also, a homogeneous-lean (or homogeneous-stratified) combustion mode exists between the homogeneous- and stratified-

Fig. 17. Comparison of (a) thermodynamic efficiencies, and (b) HC emissions (Engine speed: 2000 rpm, Mean effective work: 0.2 kJ/l) [104]. 12

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(ηv) and indicated thermal efficiency (ηi) are improved, and hence the IMEP as well. Another parameter affected by injection timing is the amount of fuel impingement. Fig. 20(a) shows the amount of fuel impingement at various injection timings. The left wall impingement increases with increasingly early injection, but the latest injection produces the highest right wall impingement. It is interesting that the differences between the left and right wall impingements are larger for the early injection cases, and no difference is discernible for SOI at BTDC 180°CA. The amount of fuel impingement can be reduced by a split injection strategy [106,107,115–117]. As shown in Fig. 20(b), the overall reduction achieved by split injection is about 50% compared to a single injection because the split injection shortened the penetration. Li et al. [108] demonstrated the effect of split injection. They compared the ambient airflows at the peripheries of the sprays produced by single and double injections. In the single injection, the droplets at the spray tip were continuously replaced by newer and larger ones that could overcome a strong drag force. Thus, only the large droplets tended to accumulate at the spray tip, and the high density of the droplet clusters at the spray tip formed a fuel-rich region [41]. The spray tail contained a very lean mixture because of the excessive air supply from air entrainment at the tail. In the split injection, the density at the spray tip was lower, and air entrainment, which leaned the spray at the tail, facilitated the evaporation of the next spray. The overall spray structure thus became more homogeneous and combustible. Consequently, the split injection achieved levels of charge homogeneity similar to those in a PFI engine [115]. These observation are believed to

Fig. 18. GDI engine operating modes depending on load and speed [202].

modes. Injection strategies, such as injection during the intake stroke or compression stroke and double- and split-injections, are one way to change the engine operating modes. 3.1. Effects of injection strategies A GDI engine system can take advantage of various injection strategies because the fuel is directly injected into the cylinder. The injection timing is one adjustable parameter highly related to the mixture quality and combustion characteristics. Injection too late could cause inhomogeneous mixing because of the slow ambient airflow and insufficient time for the evaporation of droplets. The low-quality mixture would then result in incomplete combustion and the production of significant amounts of CO and CO2. Conversely, injection too early could cause fuel impingement. We explained the fuel film phenomena caused by fuel impingement and its effects on combustion in Section 2. Costa et al. [93] and Hao et al. [114] compared the effects of early injection during the intake stroke and late injection during the initial stage of the compression stroke. The spray and flow field interaction was captured by three-dimensional visualization [114] and it was revealed that the spray plumes interacted with the intake air flow in the early injection timing. Whereas sprays at late injection (i.e., stratifiedcharge spray) were less influenced by in-cylinder flow, compared to sprays for the early injection. The injection timing also affects to the volumetric efficiency. The cooling effect of fuel evaporation increases the density of the intake air, leading to an increase in the amount of intake air. Retarding the injection timing in a GDI engine thus increases the volumetric efficiency. As shown in Fig. 19, the initial volumetric efficiency at the start of injection (SOI) at BTDC 360°CA is almost the same as that of a PFI engine because the injected fuel absorbs heat energy from the piston head and not from the intake air. However, as the injection is increasingly delayed, the volumetric efficiency increases to about 2.5% higher than that of a PFI engine with SOI around BTDC 270°CA. However, further delay of the injection decreases the volumetric efficiency because the fuel spray does not have enough time to cool the ambient air. The improvement in the volumetric efficiency increases the indicated mean effective pressure (IMEP), with the maximum IMEP eventually attained around BTDC 270°CA [92,111]. This can be explained by the following expression of IMEP [111]:

IMEP = ηv ·ηi·ηc ·

1 Pa · ·QLHV AF RTa

(1)

The air-fuel ratio (AF), combustion efficiency (ηc), lower heating value of the fuel (QLHV), and intake air characteristics (Pa/RTa) are unaffected by the injection timing. However, the volumetric efficiency

Fig. 19. Comparison of (a) volumetric efficiencies and (b) IMEPs of GDI and PFI engines [111]. 13

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Fig. 20. Fuel impingements for (a) changed injection timing (Injection pressure: 5 MPa) and (b) single and double injections (Start of injection: BTDC 210 deg, Injection pressure: 5 MPa) [116].

and the even lower power output. However, other researchers have insisted that the split injection strategy increased the torque [92,111]. They noted that split injection could simultaneously increase the volumetric efficiency and reduce the possibility of knocking, although the requirements for high volumetric efficiency and low knock tendency generally conflict. Further study is thus required to establish the effects of split injection and enable its proper application.

reduce soot and HC emissions [117]. Nonetheless, researchers have applied split injection not to achieve higher engine performance or higher torque, but to promote quality mixture formation or generate better emissions characteristics. Serras-Pereira et al. [115] used triple injection to reduce fuel impingement and improve the mixture quality. Fig. 21 shows the flame propagation procedures for single and triple injections. As shown in the figure, the gasoline single injection cases produced the brightest flame, which developed toward the lower side of the images. In contrast, the flames produced by triple injections were rather dark and developed symmetrically. These trends reflect the difference between the mixture quality achieved by single and triple injections. The brightness of the flame is related to the mixture homogeneity, and the low-luminosity flames produced by triple injection are associated with the reduced soot formation that results from homogeneous mixing [98]. In addition, the symmetrical shape of the flames produced by triple injection indicates the formation of a homogenous mixture because the flames cannot propagate toward the lean fuel region, as in the case of the flames produced from single injection. However, the use of triple injection did not achieve higher engine power in this study. Triple injection produced slower combustion than single injection and consequently the lowest peak in-cylinder pressure. Fig. 22 shows the in-cylinder pressure produced by single injection,

3.2. Cycle-to-cycle variation of spray and combustion It is evident that flame formation and propagation in spark ignition engine is remarkably effected by the turbulence varied by in-cylinder flow and has significant cycle-to-cycle variation (CCV) than compression ignition engine. Pattern of directly injected fuel in GDI engine also is altered by in-cylinder flow and impact on combustion. Thus, cycle variation from in-cylinder flow and fuel injection hinder to combustion stability and lead to more fuel consumption and excessive engine emissions. The study on cycle-to-cycle variation was conducted based on evaluating cylinder pressure and optical images. IMEP, heat release, maximum rate of pressure rise and its location, maximum pressure and its location and combustion phasing were calculated to demonstrate cycle-to-cycle variation in various injection strategies [118]. Variations of cylinder pressure and calculated values 14

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were strongly connected with the number of injection and injection timing, whereas the combustion duration of 10–90 percent of fuel burned were relatively less sensitive to fuel injection strategies than crank angle at 10 percent and 50 percent of fuel burned. Hao et al. [119] evaluated the fluctuation of in-cylinder air motion velocity and spray pattern by proper orthogonal decomposition (POD) analysis in and optical SIDI engine. POD analysis decomposed the most dominant pattern (mode 1) of set of spray image or velocity field and secondary dominant patterns (mode 2) that represented fluctuation of each images. quantified cyclic variation calculated from the POD identified the variations of the spray pattern effected by the in-cylinder air flow on a cyclic basis. The variation of early flame development for various swirl ratio was analyzed using centroid and location of flame image by David et al. [120]. Fluctuation of flame formation and centroid revealed the cycleto cycle variation of flame. High swirl ratio produces lower variation of flame structure (i.e., centroid and location), leading to improving for the COV of the IMEP and combustion stability. At four degrees earlier than flame formation with low swirl level, flame kernel was formed with high swirl level. Velocity field was derived from two sequential flame images in initial stage of combustion and used to correlate the cycle-to-cycle variation and flame speed by POD analysis [121]. Average and fluctuating parts calculated by coefficient and mode of POD provided quantitative correlation with crank angle at 5% of total heat release (CA05) and peak pressure. High fluctuating energy resulted in an earlier CA 05 and higher peak pressure. Andrew et al. [122] used neural network to predict the IMEP and cycle variation based on the COV of IMEP. 15 neurons in one hidden layer with 109 experimental inputs was trained by experiment data and correlation between prediction of COV and experiment results is significant (R2 = 0.845). There were still limitations that neural network deduced over predicted COV of IMEP for low COVs while under predicted COVs for higher variation. 3.3. Cold and warm start conditions It is well known that a cold start produces excessive amounts of soot [94]. When an engine is cold, the fuel atomization efficiency is noticeably low, and the combustion is consequently poor. The spray shapes for coolant temperatures of 20 °C and 90 °C and the corresponding flame propagation procedures are compared in Fig. 23. As shown in the figure, the spray plumes were more defined in the cold engine, and the six-spray plumes were not well dispersed. Conversely,

Fig. 21. Flame growths for single and triple injections (Injection pressure: 150 bar, Ignition timing: 325 deg CA ATDC) [115].

Fig. 22. Mean in-cylinder pressures for single and triple injections (Injection pressure: 150 bar, Ignition timing: 325 deg CA ATDC) [115]. Fig. 23. (a) Spray and (b) flame development for injection at BTDC 280° CA (injection duration = 0.8 ms, Coolant temperature: 20 °C, 90 °C) and spark at BTDC 35° CA [90]. 15

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as warm coolant water heated the injector and in-cylinder air, the atomization and evaporation processes were enhanced, and the sixspray plumes became barely distinguishable. The poor atomization characteristics in the cold atmosphere also contributed to reducing the flame speed and lowering the in-cylinder pressure. As shown in Fig. 23(b), the flame growth for 90 °C was much faster than that for 20 °C. According to Aleiferis et al. [89], the highest measured flame speeds for 90 °C were 5%–10% higher than those for 20 °C. The flame speed also affected the in-cylinder pressure. Fig. 24 compares the incylinder pressure histories for 20 °C and 90 °C coolant temperatures. For gasoline fuel, the peak in-cylinder pressure for 90 °C was about 1.12 MPa, which was about 25% higher than that for 20 °C. Shinagawa et al. [123] applied the multiple injection strategy to improve the cold-start combustion performance. This injection strategy has two injection events, in the intake stroke and the compression stroke. The first injection in the intake stroke is for the formation of a lean mixture in the combustion chamber, and the second injection is for the formation of a combustible gas mixture around the spark plug. They reported a reduction of engine start time at extremely cold start conditions using this multiple injection strategy, as shown in Fig. 25. According to their report, splitting the injection restricted the spray length in the intake stroke and minimized the wall wetting on the piston bore and head.

Fig. 25. Cold start injection strategy (Multiple injection) and cold start time [123].

plug than well-burning cycles. Because the injection and spark timings are closely related in stratified combustion, the mixture preparation procedure is sensitive to the injection strategy, which makes it particularly difficult but also important to achieve optimal mixing near the spark plug. The injection timing significantly affects the emission characteristics. As shown in Fig. 26, NO emissions decreased with an increasing injection delay [95,97]. For gasoline, the amount of NO emissions decreased from 225 ppm for SOI at BTDC 31°CA to 50 ppm for SOI at BTDC 23°CA. The NO emission reduction resulted from a decrease in the peak in-cylinder temperature. The late injection and spark delayed the combustion phase, reducing the peak in-cylinder temperature, as shown in the second graph. However, further study is required to clarify the mechanism of NO emissions reduction with increasing injection delay. Conversely, for gasoline, soot emissions and combustion instability also increased with increasing injection delay. This trade-off between NO and soot emissions is consistent with the findings of another study [126] in which researchers varied not only the injection timing, but also the spark timing to achieve the lowest IMEP variables. According to Drake et al. [98], an earlier spark produced much faster flame development. Fig. 27 shows the effect of the spark timing on the flame propagation speed for a fixed end of injection (EOI) at BTDC 36°CA; the flame front was detected by OH chemiluminescence. The flame speed can be clearly compared with that for an after start of spark (ASOS) 24°CA (the image for ATDC 2°CA is shown in the upper figure, and that for BTDC 10°CA is shown in the lower figure). These images show that a shorter delay between the injection and the spark increased

3.4. Stratified combustion strategy GDI engines are believed to enhance fuel efficiency by means of stratified combustion at low/medium engine speeds and low engine loads. A fundamental feature of stratified combustion is injection during the combustion stroke to form the combustible mixture near the spark plug while keeping the equivalence ratio near the cylinder wall low. Many researchers have implemented stratified combustion and found that the combustion characteristics were improved by modifying the engine technique and optimizing the mounting position of the injector [12]. However, combustion instability and unacceptable emissions characteristics remain major issues of stratified combustion. The combustion instability is strongly related to the injection and spark timings and local in-cylinder flow because the equivalence ratio distribution and mixture motion near the spark plug affect the flame initiation and development [124]. Velocity field and equivalence ratio distribution for the stable combustion cycles and misfire cycles were compared to identify the cause for the instability of stratified combustion using [125]. Misfire cycles have locally weaker gas velocity and leaner regions at the end of fuel injection in the vicinity of the spark

Fig. 24. In-cylinder pressure histories for coolant temperatures of 20 and 90 °C (Equivalence ratio = 0.83) [89]. 16

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the flame speed because of the higher kinetic energy induced by the spray motion. The spray motion produced by a high injection pressure increased the turbulence, resulting in faster flame propagation. Castagné et al. [127,128] observed a similar trend and noted that an injection-spark interval shorter than the optimal value could cause some misfiring; otherwise, it would increase the combustion speed. Conversely, a longer interval would result in unstable combustion. The swirl level is another important factor of the mixture preparation and combustion processes. Fig. 28 shows the effect of the swirl intensity on the flame propagation. Blue indicates flame and yellow or green indicates soot emissions. In this study, the swirl index ranged between 0.2 and 3.1, and the injection and ignition timings were fixed. As shown in the figure, a higher swirl produced faster flame propagation, resulting in higher in-cylinder pressure. However, a higher incylinder swirl level caused locally rich combustion because the swirl swept more fuel onto the cylinder wall. The authors noted that an increase in the swirl caused undesirable mixture preparation and combustion characteristics.

3.5. Emission characteristics 3.5.1. NO Exhaust gas recirculation (EGR) is one of the most effective methods for reducing NO emissions [12,129,130], and this is achieved by decreasing the in-cylinder temperature [129]. However, a higher EGR rate can cause unstable combustion. Fig. 29 shows the trade-off relationship between NO emissions and combustion instability. The mole fraction of oxygen represented on the X-axis indicates the dilution effect of the EGR; a higher EGR rate was achieved by additional N2 flow into the intake air. The graph shows that the best combustion stability, with less than 1% standard deviation of the IMEP, occurred without EGR (i.e., 21% O2), but the corresponding NO emissions were the worst. The NO emissions were significantly reduced from 67 ppm to 5 ppm by increasing the EGR rate. However, the standard deviation increased to 5%. Fig. 30 gives more details about the combustion instability. The figure shows the IMEP for 500 consecutive combustion cycles. As shown in the figure, combustion without EGR produced stable IMEP

Fig. 26. Effects of injection timing on NO emission, peak combustion temperature, soot emission, and standard deviation of IMEP [97].

Fig. 27. Effect of ignition timing on flame propagation. Comparison of experimental ensemble-averaged spectrally resolved combustion luminosities of OH* (blue) and soot (red and green), shown in the top row of each two-row set; heat-release rate obtained by CFD (blue); and rich combustion products (yellow), shown in the bottom row of each two-row set [98]. (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)

(a) Spark timing BTDC 22° CA

Flame Images ATDC

CFD results

(b) Spark timing BTDC 34° CA

Flame Images

CFD results

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(a) Swirl index 0.2

Flame Images

CFD results

(b) Swirl index 1.4

Flame Images

CFD results

(c) Swirl index 3.1

Flame Images

CFD results

Fig. 28. Effect of swirl on flame propagation with fixed spark timing at BTDC 26° CA (1st row: Experimental ensemble-averaged spectrally resolved combustion luminosities of OH* (blue) and soot (red and green), 2nd row : heat-release rate obtained by CFD (blue); and rich combustion products (yellow)) [98]. (For interpretation of the references to color in this figure legend, the reader is referred to the web version of this article.)

Fig. 29. Effects of intake dilution on NO emission and IMEP instability [97].

with a mean value of about 270 kPa. Conversely, combustion with 19% O2 produced irregular IMEP, and the difference between the maximum and minimum IMEP values was about 60 kPa, which is actually larger than the 20% mean IMEP. Nevertheless, although combustion instability increased, the decrease in the mean IMEP was insignificant at about 3%. Thus low NO emissions can be achieved while maintaining a reasonably high engine power, provided that the combustion remains

Fig. 30. Illustration of IMEP instability for operations with and without intake dilution [97].

stable. The possible causes of combustion instability are irregularity of the spark plasma, internal mixture distribution, and local in-cylinder flow [95,97,115]. Because the development of the spark plasma is also

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emissions exceeded that originating from the fuel film, which implied the existence of other sources of HC emissions. Other researchers have insisted that fuel film is not a major source of HC emissions. Frank and Heywood [132] investigated the effect of piston temperature on HC emissions and found only a minor relationship between the two. The low dependency of HC emissions on piston temperature suggests that fuel film is a minor source of the emissions. Drake et al. [91] compared the amounts of fuel film and HC emissions for swirl and multi-hole injectors. Their results also indicated no correlation between the fuel film and HC emissions, and they found no noticeable difference between the HC emissions for swirl and multi-hole injectors. They demonstrated that the amount of fuel film compared to the total HC emissions was only 15% for the swirl injector and 2% for the multi-hole injector. Furthermore, considering that a portion of the fuel film was converted into soot emissions, the actual contribution of the fuel film to HC emissions was even lower.

affected by the internal flow in the vicinity of the spark plug [131], further investigation of the internal flow and mixture preparation are required to reduce the cyclic variations. Sjöberg and Reuss [97] investigated another interesting idea. They showed that late injection near the TDC and early spark might increase the turbulence intensity during combustion, thereby increasing the mixing rate of the combustion products and the relatively cool surrounding unused air. Enhancement of the mixing rate could also be used to reduce NO emissions by shortening the duration of the temperature at which NO was produced. The effect of changing the injection timing is shown in Fig. 26. As could be expected, NO emissions were reduced by delayed injection. 3.5.2. Soot Unlike PFI engines, GDI engines produce substantial amounts of particulate matter. Although the technique of injecting fuel directly into the cylinder improves volumetric efficiency and engine power, it also has the problems of mixture inhomogeneity and rich combustion. Therefore, several studies have paid particular attention to soot emissions. Soot emissions are caused by rich combustion conditions. Thus, the soot is initially produced just after the spark in a partially fuel-rich region. However, the soot produced in the fuel-rich region is soon oxidized, and the amount emitted from the engine is negligible. The major source of emitted soot is the pool fire; i.e., the combustion of liquid-phase fuel film on the piston head [91,96,107,112]. During the pool fire, the integrated soot volume sharply increases, as shown in Fig. 31. This figure compares the soot formation and oxidation processes for the homogenous (II) and stratified (I, III) combustion modes. In the absence of a locally rich fuel region in homogenous combustion, soot formation is initiated by the pool fire at ATDC 45°CA. In stratified combustion, however, soot formation is initiated earlier than in homogeneous combustion because of the presence of a locally rich region. After the production of the maximum soot volume, part of it is oxidized, but a considerable portion remains. Fig. 32 provides more evidence that pool fire is the major cause of soot emission; the fuel film mass is proportional to the soot emissions. In this study, the authors noted that for a swirl injector, roughly 10% of the fuel film mass was converted into soot emitted from the engine [91]. They also observed that the multi-hole injector produced negligible amounts of fuel film and smoke. The spray penetration length is therefore crucial to soot emissions [101], and in that regard a multi-hole injector exhibits characteristics superior to those of a swirl injector. Meanwhile, Chen et al. [94] investigated the effect of cold conditions on soot emissions by adjusting the coolant water temperature to 20 °C and 80 °C. For a thorough discussion of the cold start, see Section 3.2. Fig. 33 compares the PM and PN emissions under cold and warm conditions. The emissions of a cold engine always contains a higher PM concentration because of poor atomization.

4. Numerical analysis In recent years, engine modeling has been used to optimize the design of GDI engines and verify the results of experimental analyses. Compared to a PFI engine, the mixture preparation process of a directinjection engine is complex. Because of the complex fuel–air mixing mechanism, most analyses of the mixture formation process are done by simulation. Numerical analyses of a multi-hole-type injector for a GDI engine have been conducted for the purposes of optimization, developing new analysis methods, and analyzing experimental results. Several attempts have also been made to conduct meaningful analyses for investigations that could not be undertaken experimentally. Hence, in this section, we merely introduce various studies on multi-hole GDI injectors rather than categorizing them along conventional lines. Here we introduce the application of modeling to GDI engines in three parts devoted to spray modeling, combustion modeling, and optimization.

3.5.3. HC and CO Many studies have revealed that delayed injection increases HC and CO emissions. Generally, HC and CO emissions depend on the mixture homogeneity and combustion efficiency because they are formed from unburned fuel or incomplete combustion. It is thus necessary to apply optimized injection strategies and ensure homogenous mixing to reduce these emissions [15,108,110,117]. The conditions for the formation of CO and HC have few differences. CO emissions are generally formed in a locally fuel-rich region because the oxygen concentration governs the oxidation reaction of the CO. However, HC emissions can be formed in both fuel-rich and fuel-lean regions; in a fuel-rich region, the incomplete combustion produces the HC, whereas unburned fuel in an extinguished flame is the source of the emission in a fuel-lean region. Although the authors focused on the fuel film as a source of HC emissions and proposed a creative idea for eliminating the effect of other sources, they observed that the amount of engine-out HC

Fig. 31. Spatially integrated soot volume as function of time, and experimental conditions (engine speed = 1500 rpm, injection pressure = 13 MPa) [96]. 19

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with multi-hole injector. [58,133] A few one-dimensional (1D) methods can determine the characteristics of flash boiling. Negro et al. [134] used the homogeneous relaxation model (HRM) to predict evaporation during internal flashing and the standard thermodynamic analysis of the CJ-point to predict the evaporation outside the nozzle. They determined the thermodynamic properties using the volume-translated Peng-Robinson equation of state (VTPR-EOS). Using that approach, they evaluated the mass flow rate and non-dimensional extinction length, an indirect measure of the evaporation rate of the jet, and got results in good agreement with their experimental results. Neroorkar et al. [135] investigated the vapor-liquid equilibrium of blended gasoline-ethanol fuel under flash boiling. The non-ideal nature of blended fuels makes the modeling of the vapor-liquid equilibrium difficult. They used the gasoline-ethanol flash (GEFlash) model and Aspen Plus to determine the flash boiling spray properties of the blended gasolineethanol fuel and demonstrated the feasibility of predicting the flash boiling of blended fuels. Researchers can use those approaches to obtain the initial and boundary conditions for a three-dimensional (3D) simulation. However, efforts are yet to be made at an active connection with 3D simulation. Coupling between 1D flash boiling modeling and 3D spray modeling is necessary to further investigate the mixture formation process under part-load conditions because flash boiling distorts the spray structure from within the nozzle, as shown in Fig. 34 [135]. Negro et al. [136] investigated combined 1D and 3D numerical modeling of flash boiling sprays for ethanol, methanol, and alcohol/gasoline blends. Their study covered the entire process of flash boiling spray modeling for a singlehole injection spray. They also used HRM to model the thermodynamic instability in the nozzle. They obtained initial droplet diameter distributions by 1D simulation and used them as input for a 3D Lagrangian spray simulation. They predicted the spray characteristics such as the evaporation rate, spray tip penetration, and SMD (D32 ) for flash boiling spray. However, they did not compare their simulation results with experimental ones. To use a 1D flash boiling model to simulate combustion, proper validation is essential. Hence, further study is required to validate the model for the simulation of flash boiling in multi-hole injection.

Fig. 32. Relationships between soot emission and maximum fuel film mass for swirl injector (filled circles) and multi-hole injector (open squares) [91].

4.1. Modeling of multi-hole-type GDI sprays The characteristics of a multi-hole-type GDI injector are similar to those of a diesel injector in terms of the nozzle concept. Numerical modeling of multi-hole GDI and diesel sprays can be done based on the blob concept, which assumes that a liquid droplet is inserted into the computational domain. However, a few differences between the two models are important. Diesel fuel is injected into ambient pressure above atmospheric pressure, whereas gasoline fuel is injected into ambient pressure below atmospheric pressure through the use of throttle valves. Hence, flash boiling occurs in a GDI engine. In addition, the droplets break up under different regimes. In diesel injection, the droplets disintegrate under a catastrophic breakup regime because the Weber number is high. However, in GDI injection, the droplets disintegrate under a bag breakup or stripping breakup regime because the Weber number is lower than that for diesel injection. Hence, different approaches are required to model the two spray breakups. In addition, the spray–wall interaction is the main subject in GDI engine modeling because the spray tip penetration exceeds the cylinder size, resulting in higher PM emissions. 4.1.1. Flash boiling In a homogeneous charge-type GDI engine, fuel is injected during the early stage of the intake stroke. Particularly in the case of a naturally aspirating GDI engine, the intake pressure is sub-atmospheric. Because the fuel is superheated in the sub-atmospheric ambient pressure, flash boiling should be considered in the analysis of a GDI engine under part-load conditions. Also, effect of flash boiling on spray shape and penetration is significant and it results in strong collapse of spray

4.1.2. Spray breakup As stated above, a GDI spray with a relatively low Weber number requires a different spray breakup model than the type used to analyze a diesel spray. Over the past few decades, the focus of spray breakup modeling has been on diesel spray, which is characterized by a high Weber number. However, models originally derived based on highpressure diesel injection have been used to analyze GDI engines [137].

Fig. 33. Total PM and PN emissions for different gasoline/ethanol blends from cold (20 °C) and warm (80 °C) engines [94]. 20

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Fig. 34. Schematic of flash boiling inside nozzle [135].

using various inclination angles for the hole axis. They used the Schmidt particle interaction model of droplet collision and coalescence, which enables proper modeling of a colliding spray. The Schmidt particle interaction model exhibits improved grid dependency characteristics by using separate collision grids, as distinct from a gaseous-phase grid. In addition, they used a presorting algorithm to reduce the computational time. With regard to the spray breakup process, they assumed that the droplet disintegration process was completed immediately after the injection instead of applying a conventional spray breakup model. Because the spray breakup and air entrainment processes occur near the nozzle, they adopted a normal-distribution-like droplet size distribution at the nozzle. The validation was done by adjusting the initial droplet size. Using that approach, they determined the vapor distribution near the nozzle. Fig. 37 compares the vapor fuel distributions determined by experiment and simulation. In addition to methods for analyzing the macroscopic spray characteristics, other LES methods can capture the primary breakup process near the nozzle. Befrui et al. [146] used the VOF-LES method to analyze the primary breakup process from within the nozzle to 4 mm outside it. Because they were modeling an axisymmetric three-hole injector, they analyzed only one nozzle. They compared the spray cone angle determined for the nozzle by the VOF-LES simulation with that determined by an X-ray experiment. As shown in Fig. 38, the spray cone

Another approach involves the development of a spray breakup model that properly describes a breakup characterized by a low Weber number. Such models include the Reitz-Diwakar breakup model, which is based on bag breakup and stripping breakup regimes [138]. The spray breakup models generally used for multi-hole GDI injectors include the Reitz-Diwakar model, Huh-Gosman model, and KH-RT model [138–140]. Other models, such as the Taylor analogy breakup (TAB) model, droplet deformation and breakup (DDB) model, wide-area virtualization engine (WAVE) model, cascade atomization breakup (CAB) model, and a hybrid of those models, can also be used to simulate a multi-hole GDI spray [141,142]. Those spray breakup models have been used in many studies investigating the spray characteristics of multi-hole GDI injectors. Malaguti et al. [143] analyzed spray characteristics in a test vessel under non-evaporative ambient conditions. They compared the Reitz-Diwakar and Huh-Gosman atomization models against their experimental results in terms of the spray tip penetration and droplet mean diameter (D10 ). Malaguti et al. [140] used multi-component modeling to investigate the characteristics of a gasoline-ethanol blend GDI multi-hole spray. Multi-component modeling is a main topic in engine analysis because the fuels generally used in commercial engines are composed of various types of hydrocarbons. There are two approaches to analyses that use multi-component modeling. The first is continuous multi-component modeling, wherein the fuel is considered to comprise continuous components, and a probability density function (PDF) is used to describe the fuel properties. The second is discrete multi-component modeling, wherein the fuel is considered to comprise discrete components. Although fuels are composed of hundreds of hydrocarbons, most fuels can be accurately modeled using only a few components [144,145]. The computational time increases dramatically when a multi-component fuel is used in an engine simulation. The choice of the key hydrocarbon elements for discrete multi-component analysis is thus essential. In this study, the authors used seven hydrocarbons and a compressed multi-component model to reduce the computational requirements. They reported that the evaporation rate was adequately predicted using a seven-component model of gasoline. Fig. 35 shows that the fuel evaporates rapidly in a multi-component approach. These characteristic results for reduced spray tip penetration compared to those obtained by single-component modeling indicate higher injection pressures. Fig. 36 shows the reduction in the spray tip penetration obtained by the multi-component approach. Because the SMD decreased with increasing injection pressure, the penetration reduction rates were higher for high injection pressures. Nishida et al. [139] performed spray simulations for two-hole nozzles

Fig. 35. Comparison of multi-component and mono-component modeling of vaporized fuel (Composition of multi-component modeling (% by mass): isopentane (0.1661), cyclo-hexane (0.2405), iso-octane (0.1920), toluene (0.1765), ethyl-benzene (0.1293), n-decane (0.0847), and naphthalene (0.0109), Injection pressure: 5 MPa, 10 MPa, and 7 MPa, Ambient pressure: 0.1 MPa, Injection duration: 1 ms) [140]. 21

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Kelvin-Helmholtz waves were detected from entrance of nozzle when hydraulic flip occurs. Fig. 39 shows stationary nozzle flow with hydraulic flip and Kelvin-Helmholtz waves from entrance of nozzle. From volume fraction contour, spray angle by primary breakup can be measured. Also, breakup length was measured by comparing with nozzle diameter. They concluded that counter bore increases vortex, convergent nozzle increases discharge coefficient and suppress hydraulic flip. They applied this approach to skew-angled nozzles and compared macroscale parameters like plume trajectory and cone angle [150]. Macroscale parameters were obtained from experimental images and well-captured by LES analysis. Fig. 38 and 39 shows the quantitative comparison between experimental results and LES approach. Plume trajectory angle which is given by beta and cone angle near nozzle were well-captured. Fig. 36. Comparison of multi-component and mono-component modeling of spray tip penetration (Composition of multi-component modeling (% by mass): iso-pentane (0.1661), cyclo-hexane (0.2405), iso-octane (0.1920), toluene (0.1765), ethyl-benzene (0.1293), n-decane (0.0847), and naphthalene (0.0109), Injection pressure: 5 MPa, 10 MPa, and 7 MPa, Ambient pressure: 0.1 MPa, Injection duration: 1 ms) [140].

4.1.3. Spray–wall interaction One of the main issues for a GDI engine is P.M. emissions. Over the past few years, multi-port injection (MPI) engines have injected fuel into the cylinder at an intake port immediately before the intake valves. Because the fuel has sufficient time to evaporate in the cylinder, the P.M. emission is negligible even though fuel film forms on the backside of the intake valves. Nonetheless, significant P.M. emissions remain a problem for GDI engines, primarily caused by fuel impingement on the piston head and cylinder wall. When fuel impinges on the head and wall, some of it forms a film on the surfaces that remains until the start of combustion and causes P.M. and HC emissions, as well as reducing the fuel economy. A multi-hole injector produces a long spray tip penetration through the momentum it generates, which increases the possibilities for P.M. formation. In a homogeneous-charge concept engine, the relatively low ambient temperature prevents the evaporation of the fuel film. In a stratified concept engine, the late injection shortens the time available for the fuel film to evaporate. Hence, regardless of the charge concept, multi-hole-type GDI modeling should consider the spray–wall interaction. Two main temperatures divide the droplet impingement regime: the Nakayama temperature and the Leidenfrost temperature. When the wall temperature increases above the boiling temperature of the fuel, the evaporation rate increases and reaches its maximum value at the Nakayama temperature, which is above the boiling temperature of the fuel. At the Leidenfrost temperature, the evaporated fuel forms a thin layer between the droplets and the wall, which minimizes evaporation. Two main physical phenomena are important in this regard: wall spray development and wall film evolution [41]. Depending on the Weber number, the wall temperature, and the dryness of the wall, some droplets stick to the wall, some rebound, some break up, and some splash. Different authors describe the criteria for these regimes and their classification differently. Because the modeling of the spray–wall interaction is important to reducing P.M. emissions and because the mechanisms of the interaction are yet to be clearly established, many studies have been conducted to model the impingement [151–153]. Bai and Gosman developed a wall impingement model widely used in the simulation of multi-hole GDI injectors [151–153]. Montanaro et al. [151] analyzed the spray–wall interaction and compared their simulation results with those from experiments conducted in a constant-volume vessel with optical access. They used the Kelvin-Helmholtz–Rayleigh-Taylor (KH-RT) model for the spray breakup and the detailed Bai and Gosman model for the spray–wall interaction. The injection rates determined in the experiments were applied to the spray modeling. Before they modeled the spray-wall interaction, they modeled the spray under atmospheric conditions and validated the spray morphology and spray tip penetration with their experimental results. Furthermore, they analyzed the impingements on the horizontal and inclined walls. To quantify the spray impingement on a wall, they evaluated the radial and height-wise penetrations after the impingement. The geometrical variables of the wall plate determined by the simulation agreed well with the experimental results,

(a) Measured vapor phase mass distribution (kg/m2)

(b) Computed equivalence ratio distribution (-)

Fig. 37. Comparison of vapor phase distributions obtained by experiment and computation (Injection pressure: 15 MPa, Injection quantity: 1.88 mg (singlehole nozzle), 3.44 mg (two-hole nozzle), Injection duration: 1.3 ms, Ambient pressure: 1.0 MPa, Ambient Temperature: 500 K) [139].

angle caused by the primary breakup was well predicted by the VOFLES method. Furthermore, they investigated the droplet size distribution, stream-wise velocity distribution, and spray cone angle and successfully determined the initial conditions for Lagrangian spray modeling. Befrui et al. [147] also used the VOF-LES method to predict the primary breakup. However, they coupled the VOF-LES nozzle flow simulation data with the spray simulation. The initial blob size distribution was obtained by analyzing the nozzle flow. Shost et al. [148] used the VOF-LES technique to investigate the effect of the length-todiameter ratio of the nozzle hole. They focused on the spray cone angle and concluded that reducing the length-to-diameter ratio of the nozzle hole increased the spray cone angle and caused a deviation of the spray trajectory from the geometric axis of the nozzle. Befrui et al. [149] conducted Volume-of-Fluid Large-Eddy-Simulation (VOF-LES) approach focusing on primary breakup from GDI singlehole flow. They investigated the effect of counter bore, nozzle geometry and nozzle l/d ratio with neglecting cavitation inside nozzle. Using LES, 22

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Fig. 38. Comparison of cone angles determined by shadow graphic imaging and VOF-LES prediction (Test fuel: n-heptane, Injection pressure: 10 MPa, Ambient pressure: 0.1 MPa, Ambient Temperature: 293 K) [146,150].

Gosman model for a flat plate. Montanaro et al. [152] performed simulations to investigate the spray impingement on a flat plate in a test vessel and compared those results with experimental ones. They used the Kelvin-Helmholtz instability model for the droplet disintegration and the Huh-Gosman theory to define the droplet velocity components at the nozzle exits. The nozzle flow was also modeled based on the assumption of potential flow. The droplet collision and coalescence were modeled using the Nordin approach. Their model assumes that the parcels collide when their trajectories intersect and that they reach the intersection point at the same time, which means that it exhibits little grid-dependency, which is important for a simulation that uses unstructured or deforming

but the radial variables determined by the simulation indicated longer penetration than that found in the experiment. They reported that the secondary droplets in the simulation had a larger tangential momentum than that determined by experiment. Because the Weber number is a critical non-dimensional parameter of the spray–wall interaction, they investigated the mean Weber numbers of the spray before and after the wall impingement. They concluded that the impingement occurred in the splash regime in the Bai and Gosman wall impingement model. They also investigated the spray morphology. Higher injection pressures produced higher height-wise penetrations after impingement, and the downward penetration of the spray on the inclined wall was long. In this study, the spray-wall interaction was well predicted by the Bai and

Fig. 39. Hydraulic flip and Kelvin-Helmholtz waves from entrance of nozzle (VOF = 0.5 iso-surface plots, Test fuel: n-heptane, Injection pressure: 20 MPa, Fuel temperature: 293 K, l/d: 1.10) [149]. 23

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walls with various geometries is required for proper modeling of a multi-hole-type GDI injector.

grids [154]. They modeled wall impingement using the approach proposed by Stanton et al. [155], which assumes four regimes of droplet impingement on the wall: sticking, spreading, rebounding, and splashing [156]. The regimes are classified based on the Weber number. Shallow water formulations were used as the flow equations to conservatively determine the mean thickness, momentum, and energy for the wall film. They analyzed the spray produced by a single-hole injector to evaluate the atomization model. They then considered the multi-hole spray to evaluate the collision model. Finally, they investigated the impingement of the spray on a flat plate to assess the modeling of the spray–wall interaction. All the modeling was done using iso-octane as the fuel. They verified the breakup model by examining the spray tip penetrations produced by different injection pressures and used them to model the spray–wall interaction. They successfully modeled the spray impingement. The morphology of the impingement was well estimated by the spray–wall interaction model, and they detected the development of the fuel film in three directions on a flat plate. Abart et al. [153] used various spray–wall interaction models to develop an appropriate impingement model. They then used their model to evaluate the piston bowl geometry, calling their procedure the “multi-modeling” approach. They evaluated the spray impingement model for an edge plate with a geometry similar to that of a piston bowl and for the exit angle. Using that geometry, the different impingement models produced different spray morphologies. Thus, their model is appropriate depending on specific conditions. The geometry of a multi-hole-type GDI engine piston is too complex to form a stratified mixture by the wall-guided method. In a wallguided-type GDI engine, the spray flows over the surface of the piston bowl. The proper geometric design of the piston bowl is therefore essential. Thus, an in-depth study of the impingement of the spray on

4.1.4. Fuel/air mixture formation The fundamental difference between GDI and MPI engines is the mixture formation process. When fuel is injected into the cylinder, it evaporates, and the vapor mixes with air. Proper mixing guarantees proper combustion characteristics. Thus, the mixture formation process requires in-depth investigation. Experimental investigation of the mixture formation process is difficult because of the difficulty of capturing the vaporized fuel. However, important mixture characteristics, such as the equivalence ratio distribution, turbulence intensity, and in-cylinder flow, can be easily investigated using a numerical approach, with several methods available for analyzing the results. The modeling methods discussed above are used to model the mixture formation. Careful evaluation of each model is therefore necessary for reliable mixture formation analysis. During the past few years, studies on multi-hole type GDI engines have focused more on the formation of a homogeneous mixture than a stratified mixture. Presently, stratified combustion, which is the main advantage of GDI engines, requires more attention to improve fuel efficiency. The formation of a rich mixture near the spark plug while maintaining a globally lean condition is the key feature of the lean-burn concept of GDI engines. The fuel is injected during the late stage of the compression stroke to allow less diffusion time. Because the fuel is injected at the end of the compression, the spray develops in the piston bowl. The condition inside the cylinder is also vital to achieving stable combustion. Thus, proper analysis of an engine that uses the stratified concept is essential. In a homogeneous-type GDI engine, the focus is on the homogeneity and turbulence intensity. The higher the mixture homogeneity, the

Fig. 40. Instantaneous breakup droplet numbers for different breakup mechanisms (Injection quantity: 11.55 mg, Injection duration: 1.5 ms, Fuel temperature: 290 K, Ambient temperature: 291 K) [159]. 24

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injection was timed to coincide with the compression stroke. Double injection with increased injection pressure was used to study its effect. By investigating the equivalence ratio at the spark plug for different injection timings, they concluded that two-stage injection tended to improve the fuel economy and led to the formation of a relatively lean mixture near the spark plug. A single-injection strategy is more appropriate for obtaining a rich mixture near the spark plug than a twostage injection. Millo et al. [160] compared the swirl injector and multi-hole injector under part-load and full-load conditions in a turbo-charged GDI engine. First, they validated the spray model with respect to the morphology, axial velocity, spray tip penetration, and SMD. Then, they focused on the air-fuel ratio inside the cylinder and the liquid fuel on the liner and valve surfaces. Jajcevic et al. [161] used high-pressure injection to reduce the fuel concentration in the exhaust system of a two-stroke GDI engine. They applied the WAVE and TAB models to the spray breakup. Only one case was implemented using the TAB model because the rail pressure was very low, and the Weber number for the condition was too low for the adoption of the WAVE model. They used the wall-jet model rather than the wall-film model because they were analyzing the warm engine condition, and the high-pressure injection had a high Weber number. First, they determined the flow characteristics and suggested three spray patterns. They then investigated the equivalence ratio distributions for those patterns at different engine speeds. Grover, Jr. [162] used fuel tracer diagnostics to separately track the liquid and vapor fuel components of individual spray plumes from a multi-hole-type GDI injector. He also compared the mixture preparation processes of multi-hole and swirl-type injectors. Fig. 41 shows a sample fuel tracer diagnostic. The fuels injected by the different nozzles are colored differently. He used this procedure to determine where the vapor fuel near the spark plug came from and how much effect the nozzle had on the formation of the fuel film. He reported that the vapor mass fraction of the fuel injected by nozzle 3 (directed mostly toward the upside of the cylinder) constituted most of the fuel near the spark plug and that the vapor mass fractions of the fuel injected by nozzles 1 and 5 (directed toward the downside of the cylinder and the piston) constituted most of the fuel film. He also investigated the fuel vapor distribution of the fuel injected by each nozzle. Based on the obtained results, he redesigned the spray pattern and performed simulations using the new pattern. The new spray pattern produced improvements by reducing fuel film formation and generating a rich mixture near the spark gap. Kӧpple et al. [163] adopted a multi-component surrogate fuel in modeling the spray–wall interaction and investigated the effect of multi-component modeling on the mixture formation process in an engine. They reported that the application of a single-component fuel to spray–wall interactions created a problem. When n-heptane was used as a surrogate fuel, the calculated fuel film evaporation rate was high, resulting in more homogeneous mixing. To capture the relatively rich mixture near the piston wall produced by the wall film, the range of the boiling temperature of gasoline was mapped using a three-component surrogate fuel of n-hexane, iso-octane, and n-decane. They also used a calculated modified wall temperature rather than the fixed wall temperature conventionally used for engine simulations. When spray impacts a wall, the wall temperature decreases through the spray-cooling effect, which increases the fuel film on the in-cylinder surface, especially near the Leidenfrost temperature. Using this procedure, they evaluated the simulation results by comparing them with those from experiments. They used the WAVE breakup model and the Kuhnke spray–wall interaction model [164] as refined by Birkhold [165], the approach for evaluating wall film evaporation at medium and high Reynolds numbers proposed by Sill [166] and Himmelsbach [167], and that for evaluating the heat transfer between the wall and the wall film proposed by Birkhold [165]. The application of the multi-component fuel to the simulation increased the relatively fuel-rich region because

Fig. 41. Sample fuel tracer diagnostic for tracking liquid and vapor fuels injected by each nozzle [162].

lower the emissions because P.M., HC, and CO are generated in the relatively rich mixture region, and a relatively lean mixture tends to cause flame extinction, which increases HC emissions. The turbulence intensity determines the turbulent flame speed. The higher the flame speed, the higher the IMEP, which is directly related to engine performance. Knocking can also be suppressed by increasing the turbulence intensity [157,158]. James et al. [138] compared the equivalence ratios at the spark plug for swirl and multi-hole injectors with different injection timings (90°, 180°, and 270° ATDC). They shrouded the intake valves to increase the tumble motion in the cylinder and enable the implementation of the airguided concept. Iso-octane was the fuel. They used the Reitz-Diwakar model to predict the spray breakup and the Bai wall impingement model to predict the spray–wall interaction. They investigated the mixture quality based on the equivalence ratio at the spark plug and the volume fractions for equivalence ratios of 0.5–1.5. They concluded that a high-pressure multi-hole injector was suitable for a high-tumble GDI engine because it satisfied the full-load mixture requirements and provided good mixture quality under part-load conditions. Li et al. [159] used hybrid breakup models to analyze the effects of the background and injection pressures in a constant-volume vessel on the mixture formation process in a wall-guided GDI engine with a newly designed piston bowl. They used the Huh-Gosman breakup model or KH breakup model to model the primary breakup and the RT breakup model to model the secondary breakup. To determine whether to apply the KH model or the Huh-Gosman model to the primary breakup, they calculated the breakup time scales of the two models and chose the model with the shorter time scale. They used the secondary breakup model when the length between the nozzle and the parcel exceeded the breakup length. They compared the results of the Huh-Gosman–KH–RT model they developed with those of the Huh-Gosman–Reitz-Diwakar model and the KH-RT model. They reported that the Huh-Gosman–Reitz-Diwakar model predicted a linear increase of the penetration, which indicated a lower level of the secondary breakup. The KH-RT breakup model indicated that the droplet sizes hardly changed during the primary breakup, and it could also predict the initial spray development. Compared to those models, the Huh-Gosman–KH–RT breakup model well predicted the spray tip penetration, which meant that it well captured the breakup speed of both the primary and secondary breakups. They also investigated the instantaneous breakup droplet number for each breakup model. Fig. 40 shows the effect of each model on the spray breakup. The primary breakup was implemented by the Huh-Gosman model rather than by the KH model. They used this model to analyze the mixture formation process in an engine that used the stratified combustion concept. To maintain the stratified mixture, the 25

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Lean premixed Flame branch

of the high wall film mass and the late evaporation of the wall film. Yi et al. [168] investigated mixture formation in a GDI engine using iso-octane as a surrogate fuel for gasoline. They used the KH-RT model for the spray breakup, the O’Rourke collision model to predict the droplet collision and coalescence, and the wall-jet model to capture the spray–wall interaction. They used experimental data to evaluate the spray model in terms of the spray tip penetration and droplet size. After the validation, they adjusted the engine speed, injection timing, and injection pressure to analyze the mixture preparation process and investigated the equivalence ratio at the spark plug and the vapor fuel distribution. They also compared the fuel film formation processes of the wall-jet and wall-film models. Liu et al. [169] simulated the mixture formation process in a multihole-type GDI injector. They used the Huh-Gosman spray breakup model and assumed the droplet size to have a Rosin-Rammler distribution. The evaporation of the liquid fuel was modeled using the Dukowicz model, and the spray-wall interaction was modeled using the method developed by Naber and Reitz [170]. They used the spray tip penetration, spray morphology, and droplet size to evaluate the spray model. After the validation, they investigated the mixture formation process under full-load conditions using different engine speeds and analyzed the in-cylinder mixture flow, equivalence ratio distribution, vaporized fuel mass, global equivalence ratio, equivalence ratio at the spark plug, and tumble ratio. Li et al. [142] used 2,5-dimethylfuran (DMF), an alternative fuel for SI engines, to model the mixture formation process in a GDI engine. They adopted the primary breakup model developed by the Max Planck Institute to increase the performance of the blob concept in the primary breakup model, and they applied the TAB, CAB, and KH breakup models to the secondary breakup. They investigated the penetration, spray axial velocity, and AMD used for the spray modeling and concluded that the CAB model produced a better droplet size distribution than the KH and TAB models. Their simulation of the mixture formation process revealed that DMF had a lower evaporation rate, which increased the inhomogeneity of the mixture. They recommended the development of a new spray pattern for DMF fuel for application to a GDI engine to increase the mixture homogeneity. As summarized above, various studies on the mixture formation process have used simulations when experimental investigation was not possible, which has led to the improved reliability of numerical analysis techniques. Researchers have investigated the effects of applying various spray breakup models, spray–wall interaction models, injector types, injection strategies, and spray patterns to the mixture preparation process. Special methods that use only simulation, such as fuel tracer diagnostics, have also been studied. Those approaches have been used to analyze experimentally inaccessible data such as the equivalence ratio distribution, equivalence ratio at the spark plug, turbulence intensity, in-cylinder flow, and fuel film formation on the in-cylinder surface. Improved computational technology has broadened the scope of application and increased the number of available numerical approaches. Various models of GDI engine spray have been developed, and numerical analyses have been used to design engines and improve their performance. We expect the application of numerical analyses using experimentally inaccessible data obtained by simulation to increase.

Diffusion flame O,O2,CO2,H2O,etc

Stoichiometric Premixed flame Diffusion flame Rich premixed Flame branch

Droplets CO,H,OH,etc

Fig. 42. Structure of triple flame [172].

generated in the region behind it. After propagation of the premixed flame, hydrocarbons are left in the relatively fuel-rich region, and the oxidant is left in the relatively fuel-lean region. Because the temperature of the burned region is high, a reaction occurs between the hydrocarbons and the oxidant. This process is called a “triple flame.” A schematic of a triple flame is shown in Fig. 42. This phenomenon should be considered in modeling the combustion of a GDI engine. In this section, we introduce two combustion models for a triple flame: the extended coherent flamelet model (ECFM-3Z) and the G-equation model with CHEMKIN. Bonatesta et al. [171] applied the ECFM-3Z model to a wall-guided GDI engine with a multi-hole-type injector. The model assumes that the cell consists of a pure fuel zone, a zone containing pure air and residual gases, and a mixed zone, which means the model can be used regardless of the mixture regime. The diffusion flame and the premixed flame can thus be modeled using the same combustion model. Therefore, the ECFM-3Z can model not only the premixed combustion, but also the stratified combustion that contains the triple flame. Yang et al. [172] used the G-equation model with detailed chemical kinetics to model a gasoline turbocharged direct-injection (GTDI) engine and a PFI engine. The G-equation uses the flame speed to predict the premixed combustion and CHEMKIN to predict the chemical kinetics behind the flame and the knocking in front of the flame. Because CHEMKIN can predict the diffusion flame, it can also analyze a triple flame. Chiodi et al. [157] used simulation tools to investigate the design process of a multi-hole type GDI injector. First, they used a 1D-CFD tool to fix the main factors such as the valve timing, intake runners, and exhaust manifold design. They then used the Quick-Sim 3D simulation tool to evaluate the spray model and its morphology. They used that model to investigate the fuel concentration and turbulence. Finally, they investigated the combustion process under the wide open throttle (WOT) condition. They analyzed the flame development with respect to the mixture homogeneity and turbulence. In this engine, mixture homogeneity had a greater effect on flame development than did turbulence. Bonatesta et al. [171] used the CFD method to analyze a modern wall-guided GDI engine. They used the Reitz-Diwakar spray breakup model to evaluate the spray tip penetration and spray morphology and compared the results with those of experiments. They used the BaiONERA model to predict the spray–wall interaction during the mixture formation and applied ECFM-3Z to the non-homogeneous and SI turbulent combustions to investigate the mixture preparation and combustion processes. They further investigated the in-cylinder motion, fuel evaporation rate, fuel film on the wall, fuel-air mixture distribution, and combustion pressure and temperature distributions and reported that the mixture preparation continued after ignition. This phenomenon can be more effectively investigated by simulation than by experiment and should be considered in a combustion analysis. Costa et al. [173] investigated the benefits of multiple injections in the mixed mode to boost the operation of a GDI engine. They modeled the droplet breakup using the Huh-Gosman model and the droplet evaporation using the single-component Dukowitz model. They

4.2. Modeling the combustion process in GDI engine Numerical analysis of the mixture formation process in a GDI engine has attracted attention in recent years because the engine emissions can be predicted from the mixture quality. However, further analytical investigation of the combustion is also necessary. A few approaches to modeling the combustion of a GDI engine are available. The mixture produced in a GDI engine has a low level of homogeneity compared to that produced in an MPI engine. When the fuel is ignited in the stratified condition, the premixed flame propagates as the diffusion flame is 26

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the piston geometry, are the essential design parameters of a GDI engine with a multi-hole injector. Those parameters proactively affect the fuel efficiency and emissions characteristics, so their appropriate determination is essential to improving engine performance. Xu et al. [178] optimized the combustion chamber geometry by multi-component fuel modeling. Because the distillation behavior of real fuel is different from that of a single-component fuel, they used a multi-component fuel of n-pentane, iso-octane, and n-decane to approximate the distillation curve of gasoline for the simulation. Because of its heavy components, the multi-component fuel exhibited slower evaporation than a single-component fuel. They then used multi-component modeling to analyze the engine for different cold-start temperatures and found that the single-component fuel evaporated faster at high temperatures and more slowly at low temperatures than the multicomponent fuel. They analyzed the mixture preparation for different piston designs using the double-injection strategy for the cold start condition. They modified the bowl depth, back wall height, shapes of the bowl and back wall, and location of the bowl to form a rich mixture at the spark plug while maintaining turbulence intensity at the TDC. They investigated the air-fuel ratio distribution, air-fuel ratio at the spark plug, piston wetting, combustion stability, and smoke to determine the optimal piston design. They concluded that robust fuel-air mixture preparation is important for stable combustion under cold-start conditions. They chose a wide bowl design for improved mixture preparation and increased the turbulence intensity by using a smoother dome. Pontoppidan et al. [141] used a steady state analysis followed by a transient state analysis to optimize the port design to increase the tumble motion. They validated the spray simulation for a constantvolume chamber under atmospheric conditions, evaluating the spray width and spray tip penetration. They then modified several variables, such as the number of injector holes and the injection timing. Finally, they investigated the correlation of the burn duration and concluded that the error increased with decreasing engine speed. Given the complex geometry of a GDI engine, it is difficult to perform automatic optimization using an algorithm, such as a genetic algorithm. As noted above, geometric optimization is difficult. Moreover, geometric optimization using an experimental method is costly and time-consuming. Optimization by numerical analysis is therefore useful. The direction of the optimization to improve engine performance is determined by the user. Proper understanding of the mixture formation and combustion characteristics is thus essential to the optimization process.

assumed the initial droplet size to have a probabilistic log-normal distribution. The spray tip penetration was well captured by the spray model. They investigated the mass of the evaporated fuel to model the mixture formation and the combustion pressure to model the combustion. They concluded that double-injection improved the quality of the stratified mixture under moderate speed and moderate load conditions but not under high speed and high load conditions. For some studies, LES (Large Eddy Simulation) approach has been adopted for modeling combustion process of GDI engine. LES approach is usually used to predict cycle-to-cycle variation (CCV) of combustion pressure and investigate primary breakup of fuel from nozzle hole. CCV characteristics are hard to investigate using RANS which utilize average velocity. Therefore, there are efforts to predict CCV using LES approach. Also, primary breakup can be analyzed using LES. Because large eddies are directly solved, droplet disintegration form liquid core can be analyzed. For this reason, nozzle parameters which affect primary breakup can be investigated. Fontanesi et al. [174] utilized LES to investigate cycle-to-cycle variation for a turbocharged GDI engine including spray and combustion process. Reitz model [175] was adopted for spray breakup and Bai model [176] was used for droplet-wall interaction. Combustion was modeled by ECFM-3Z. Both LES and RANS with cycle-specific boundary condition were used to predict CCV. Cycle-specific boundary condition was obtained from experiment and 1D simulation. They concluded that LES is more appropriate for predicting CCV characteristics than RANS. Fig. 43 shows that results from LES approach exist between minimum and maximum combustion pressure history of experiment though combustion pressure histories from LES approach are biased to maximum combustion pressure history of experiment. They also conducted correlation coefficient analysis. Most significant factor was turbulent characteristics. Teodorczyk et al. [177] investigate the cycle-to-cycle combustion pressure variability. WAVE model was used for spray breakup. Gradient Combustion Model (GCM) was used for combustion analysis and chemical reaction was described by one chemical reaction equation. Pressure data from LES ranged between experimental minimum and maximum data. They concluded that CCV characteristics can be investigated using multi-cycle LES method. As stated above, numerical analysis can predict emissions based on the fuel-air mixture distribution and determine the engine output. Studies on exhaust emissions such as NOx, HC, and CO have been rare, whereas PM emissions, which are the main issue for GDI engines, have been investigated by combustion simulation [171]. In contrast, the flame development direction and speed have been thoroughly investigated. A former MPI gasoline engine used a TWC, which significantly reduced NOx, HC, and CO emissions by eliminating oxygen from NOx and providing oxygen to HC and CO. This exhaust emission treatment enabled the gasoline engine to exhibit remarkable exhaust performance. However, that performance was limited to the stoichiometric condition. Hence, an MPI engine requires a stoichiometric amount of fuel to use a TWC. A GDI engine uses stratified combustion. A relatively small amount of fuel is injected and guided to the spark plug to reduce pumping loss. However, the lean condition of stratified combustion prevents the use of a TWC. These combustion strategies enable emissions reductions, but a thorough combustion analysis is required to achieve clean exhaust. 4.3. Optimization of combustion chamber geometry and engine operating conditions One of the most interesting and beneficial outcomes of numerical analysis is optimization, which can be done with respect to various variables. Injection timing, spark timing, injection strategy, combustion chamber geometry, intake and exhaust port geometry, and spray pattern are some of the factors that can be optimized. The injection strategy, spray pattern, and combustion chamber geometry, including

Fig. 43. Comparison of combustion pressure between experiment (maximum, minimum) and LES analysis (cycle to cycle variation analysis) [174]. 27

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4.4. Modeling challenge

Table 3 Physical and chemical fuel properties of gasoline and bioethanol* [186].

Numerical analysis of multi-hole-type GDI engines is facing challenges to predict PM and PN which are the major emissions from GDI engines. Widely used prediction model for PM and PN is not present while the soot model is developed and used widely. For this reason, future researches would focus on modeling PM and PN formed in GDI engines. To predict PM and PN, accurate fuel film prediction is essential. Therefore, improved fuel film prediction model in addition to the PM and PN emission prediction model should be developed. Fuel film prediction can be improved by considering specific spray-wall interaction criterion, conjugate heat transfer, and multi-component fuel modeling. Fuel film model is used when the wall temperature is low and wall-jet model is used when the wall temperature is high. In future works, spray-wall interaction model for transition regime considering the wall temperature is necessary. After droplets collide to the wall and form film, wall temperature decreases due to heat of vaporization. This affects the vaporization rate and should be considered in future study. Also, the heavy component fuel in gasoline evaporate slowly and can be a cause of PM emission. For this reason, multi-component modeling for fuel film evaporation should be studied. In future studies, optimization technique would be actively used for design process. Especially, in multi-hole-type GDI engines, nozzle direction is the major design parameter because of its high penetration and momentum. To reduce fuel film and improve mixture homogeneity, nozzle direction should be properly designed. Therefore, numerical analysis would be utilized for spray pattern optimization to reduce emissions. Many spray pattern optimization studies would be performed on various types of multi-hole-type GDI engines. (e.g. high tumble engine, short stroke engine, high-speed focused racing engine, etc.)

Characteristics

Bioethanol (E100)

Gasoline (G100)

Chemical formula Main constituents (% by weight)

C2H5OH C: 52 H: 13 O: 35 789.67 1.57 0.027 98–100 26.8 15.8 3–19 9 Bioethanol 100%

C4–C12 chains C: 85–88 H: 12–15 O: 0 718.33 0.84 0.024 86–94 41.9–44.2 55.2–103.4 1–8 14.7 Gasoline 100%

Fuel density (kg/m3 @20 °C) Kinematic viscosity (mm2/s @20 °C) Surface tension (N/m @20 °C) Octane Lower heating value (MJ/kg) Reid vapor pressure (kPa) Ignition point – Fuel in air (%) Stoichiometric air/fuel ratio (by weight) Blending ratio (%)

* Handbook for handling, storing, and dispensing E85, U.S. Department of Energy, Energy Efficiency and Renewable Energy, July 2006, DOE/GO-1020062343.

Bioethanol has some properties that differentiate it from gasoline. Its heating value is about one-third that of gasoline; a greater quantity of bioethanol than gasoline is therefore required for the production of a given amount of energy. Bioethanol also has a higher latent heat of evaporation, which improves its volumetric efficiency through the absorption of a significant amount of heat in the combustion chamber. The 34.7% oxygen content of bioethanol improves combustion efficiency and increases the combustion temperature. In addition, bioethanol has a lower density than gasoline and contains no mono- or polyaromatic hydrocarbons. The low C–H ratio of bioethanol is the reason for its low adiabatic flame temperature. Finally, bioethanol has a higher laminar flame propagation speed than gasoline. This causes an early completion of the combustion process and improves the combustion efficiency. A detailed comparison of the fuel properties of gasoline and bioethanol is presented in Table 3 [186].

5. Alcohol-gasoline blends for GDI engines Due to the depletion of fossil fuel sources and the tightening of emissions regulations, many researchers are actively studying the use of alternative fuels such as bioethanol and bio-butanol as well as gas fuels such as liquefied petroleum gas (LPG) as substitutes for gasoline in GDI engines. Among them, bioethanol is the most representative alternative fuel for SI engines.

5.2. Effect of alcohol blending on spray and atomization characteristics Some detailed experiments were recently performed on the use of alcohol fuels such as butanol and DMF to blend with or replace gasoline. Due to their anti-knock properties and high oxygen content, which reduces CO and HC emissions [187,188], as well as other thermodynamic similarities to gasoline, much attention has been given to the possible use of such fuels in SI engines. Butanol is considered as an alternative to gasoline because of its higher energy content, lower latent heat, and lower corrosiveness. Therefore, Marchitto et al. [187] used a PDA system to compare the spray atomization processes of commercial gasoline and pure n-butanol. They determined the droplet diameter and velocity along the axis of the spray and analyzed the effect of using butanol on the spray produced by the GDI nozzle. They reported that the droplets produced by the nbutanol spray were larger than those of gasoline under all operating conditions, with the diameters ranging between 11 and 15 μm. In addition, the velocity of the n-butanol droplets 5 mm from the nozzle tip was higher than that of gasoline droplets by approximately 13 m/s. These results can be attributed to the higher viscosity and surface tension of n-butanol, which negatively affects the atomization of bigger droplets travelling at higher velocities and possessing larger momentum. The biomass derived from DMF has been regarded as a better alternative to gasoline than ethanol [189] because DMF has a higher boiling point and energy density than ethanol, so it affords significantly better fuel economy and combustion. Tu et al. [190] used PDPA and high-speed imaging to investigate the spray characteristics of pure fuels (DMF and iso-octane) under different ambient pressures (1, 3, and 7 bar) and injection pressures (50, 100, and 150 bar). Based on analyses

5.1. Production and properties of bioethanol Bioethanol is produced in the same way as ethanol, and both fuels have the same molecular structure (C2H5OH). The major difference is in the raw materials. Bioethanol can be produced from various feedstocks such as sugar cane, sugar beet, sorghum, grain, switch grass, and sweet potatoes [179–181]. The effect of CO2 reduction depends on the type of feedstock used to produce the bioethanol. Life cycle analysis (LCA) has revealed that, compared to gasoline, the reduction effect of cellulosicbiomass-derived bioethanol is 86%, whereas that of corn-derived bioethanol is 19%–52% [182]. Bioethanol is more reactive than hydrocarbons such as gasoline. It contains a hydroxyl radical as the polar part and a carbon chain as the non-polar part [183]. Bioethanol can therefore be easily dissolved in non-polar (e.g., gasoline) and polar (e.g., water) substances. In addition, bioethanol is widely used because of its regenerative and biodegradable characteristics. Bioethanol is mainly used in a blend with gasoline because the application of 100% bioethanol to an SI engine would require modification of the engine components and systems. Bioethanol can be classified as first or second generation depending on the feedstock used for its production. First-generation bioethanol is mainly produced from sucrose that contains biomass, including starchy biomass. Second-generation bioethanol is mainly produced from lignocellulosic biomass [184,185]. Generally, bioethanol is produced by the fermentation of glucose contained in sugars and starchy biomass. 28

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NO emissions of ethanol are lower because the stronger cooling effect that results from the higher heat of vaporization of the fuel decreases the combustion temperature. Its lower LHV and lower stoichiometric air/fuel ratio further decrease the in-cylinder temperature. To reduce soot emissions and combustion instability, Sjöberg and Reuss [95,97] injected E85 fuel instead of gasoline and used a spark even earlier than the injection. Consequently, the spark plasma developed before the spray reached the spark plug, as can be observed from Fig. 44, which shows the high-speed flame images of the spray–spark interaction and the combustion process. The abnormally early spark could reduce soot emissions with relatively small IMEP variations because of the high heat of evaporation of E85. E85 is more difficult to evaporate than gasoline because it requires more energy. These characteristics prevent the formation of an ultra-rich region even at the periphery of the liquidphase fuel. The strong cooling effect of the vaporization of ethanol also reduces the flame speed in the fuel-rich areas. de Francquevil et al. [117] used E85 to reduce HC and CO emissions. In Fig. 45, the upper row shows the emissions maps for various engine speeds and engine loads using gasoline fuel, and the lower row shows the different emissions for gasoline (i.e., E0) and E85 fuels. As shown in the figure, the HC emissions increase with increasing engine speed because of the reduced time available for post-oxidation. Meanwhile, the HC emissions were unaffected by the ethanol content of the fuels. The figure shows that the differences between the HC emissions of gasoline and E85 for all the engine operating conditions were less than 2 g/kWh. These results demonstrate that the HC emissions were more significantly determined by the injection strategy and the local equivalence ratio distribution than by the fuel properties. Conversely, a significant reduction in CO emissions was observed for E85, especially under high-load conditions. The oxygen content of the ethanol apparently enhanced the oxidation reaction during the combustion. In addition, the higher heat of evaporation of E85 lowered the LHV, which also contributed to CO oxidation. Chen et al. [94] investigated the P.M. characteristics and mixture formation of a GDI engine fueled by stoichiometric gasoline-ethanol blends. They tested the blended fuel using six different volumetric ratios. They reported that an increase in the ethanol content increased the PM and PN for both cold and warm operating conditions, as shown in

of the spray characteristics, such as the penetration, angle, and distributions of the droplet velocity and size under different injection conditions, they concluded that DMF has a larger SMD and penetration length than iso-octane. They attributed that result to the higher surface tension of DMF. They also concluded that the application of DMF to an automobile engine would require modification of the injector to produce a more spatially distributed spray than that achieved by existing injectors. Park et al. [186] investigated the spray characteristics and atomization performance of gasoline and bioethanol fuels. They found that 100% bioethanol produced a slightly wider spray and larger cone angle than gasoline. In addition, they reported that bioethanol produced larger droplets than gasoline because of its higher kinematic viscosity and surface tension. 5.3. Effect of alcohol blending on combustion and emissions characteristics It is known that ignition delay of ethanol is lower than gasoline or its surrogate fuel (i.e. iso-octane) in auto-ignition condition. [191,192] Therefore, ethanol can be used to promote reaction in Gasoline Compression Ignition (GCI) engines. Also, the laminar burning velocity of ethanol is higher than the gasoline. [193,194] Therefore, ethanol addition increases combustion speed for the conditions when the turbulent characteristics are not affected by fuel injection, such as, early injection strategy. Butanol shows similar characteristics with ethanol: shorter ignition delay time and faster combustion speed than gasoline [195,196]. However, the effect of reducing ignition delay and increasing combustion speed for addition of butanol is lower than that of ethanol [197]. The main effects of the addition of ethanol or bioethanol to gasoline are an increase in the laminar flame speed, a reduction of the adiabatic flame temperature, and an increase in the heat of vaporization. Several researchers have used bioethanol as an alternative fuel to reduce NO emissions [43,46,51,79]. Sjöberg and Reuss [97] compared NO emissions for gasoline and E85 (85% ethanol + 15% gasoline) and observed that the latter was about 43% lower. de Fracnquevill [117] also conducted similar experiments and reported that E85 produced about 35% less NO than gasoline. Ethanol has a lower LHV, higher heat of vaporization, and lower stoichiometric air/fuel ratio than gasoline. The

Fig. 44. Spray and flame images for injection at ATDC −6° CA, spark at ATDC −12° CA, and E85 fuel. (Crank angle in upper left, AHRR in lower left, cumulative AHRR in lower right, and LED illumination in upper right. Exposure duration = 40 μs) [95]. 29

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(a) Maps of HC & CO emissions for gasoline fuel

(b) Maps of differences between E85 and gasoline in HC & CO emissions

Fig. 45. HC and CO emissions for all investigated conditions [117].

resulted in considerable HC emissions. In addition, the ethanol blends reduced injector deposit formation by lowering the nozzle temperature and through the nature of the multi-component fuel. Turner et al. [201] studied the effects of the blending ratio of bioethanol and gasoline, spark timing, and injection strategy on a GDI engine under part-load and speed conditions. They found that an increase in the bioethanol content reduced the engine-out emissions and increased efficiency. They attributed those results to the change in the evaporation properties. An increase in the bioethanol content increased the vaporization pressure. The oxygen in the bioethanol increased the flame speed, resulting in enhanced combustion initiation and stable combustion and eventually improving engine efficiency. Splitter and Szybist [187] studied the effect of using an alcohol-gasoline blend on the combustion characteristics of an SI engine. They tested a single-cylinder direct-injection gasoline engine under full-load conditions (λ = 1) and found that biofuels with higher octane numbers were more suitable for higher compression ratios and improved the stoichiometric torque capability. E30 (30% ethanol) in particular doubled the stoichiometric torque capability for a compression ratio of 11.85:1.

Fig. 33, because of the difficulty of evaporation caused by ethanol’s high heat of evaporation. Fuel with a higher ethanol content formed a less homogeneous mixture, resulting in the emission of more soot. However, this observation contradicts the explanation presented in Section 3.3. It was noted in that section that the use of ethanol reduced soot emissions for the same reason. It can thus be concluded that ethanol is suitable for stratified combustion because it does not produce an ultra-rich mixture, but the emission produced by its homogenous combustion is worse because of poor atomization. Turner et al. [198] investigated the combustion characteristics of a sports car with a supercharger fueled by E85. They concluded that the use of the alcohol fuel reduced the CO2 emissions and that a low-carbon fuel (ethanol and methanol) was suitable for an SI engine charged with a high compression ratio. In addition, they reported that E85 with a high octane number could extend the knock limit of a high-performance engine, thereby improving the thermal efficiency. Cairns et al. [199] studied the effect of using gasoline-ethanol and gasoline-butanol on the combustion and emissions characteristics of a turbocharged direct-injection spark-ignition (DISI) engine. They characterized their experimental results into those for part-load operating conditions and those for boosted-high-load operating conditions. They found that the EGR tolerance could be improved by using high-ethanol-content blends under part-load conditions. In addition, they reported that the engine performance could be improved by using blended fuels and that the use of high-ethanol-content blends significantly reduced smoke emissions. Taniguchi et al. [200] studied the feasibility of using ethanol fuel in a direct-injection SI engine. Their results indicated that pure ethanol fuel improved the full-load engine performance of a high-compression-ratio engine by improving the volumetric efficiency and suppressing knock. However, poor combustion caused by insufficient air/fuel mixing

6. Summary In this review article, we presented an overview of the spray/atomization and combustion/emission characteristics of gasoline directinjection engines. In addition, we discussed the numerical analysis approach and the application of alternative alcohol fuels to GDI engines. The key features of the GDI engine system are summarized as follows: 1. The advantages of a multi-hole GDI injector include flexibility of 30

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2.

3.

4.

5.

6.

7.

number of nozzle holes, and injection timing can also be optimized. Optimization can improve the combustion stability, combustion speed, and exhaust emissions. Combination of engine optimization and artificial intelligence is future challenge in numerical optimization. 8. Alcohol fuels have been considered as alternative fuels for GDI engines. The representative alcohol fuels for GDI engines include bioethanol, biobutanol, and DMF. Their use increases the droplet size because of their high viscosity and surface tension compared to gasoline. The spray tip penetration of butanol or a blend of bioethanol and gasoline is slightly longer than or the same as that of pure gasoline. Furthermore, the addition of alcohol fuel to gasoline reduces NOx and CO emissions. HC emissions are mainly affected by the injection strategy and the engine operating conditions, including the local equivalence ratio. Research on the compensation for lower energy density of alcohol fuels than gasoline is essential.

plume targeting, precise control of the injection timing and quantity, and relatively higher spray velocity with a constant spray angle. Various valuable efforts have been made to investigate and quantify the spray characteristics of a multi-hole GDI injector under different injection conditions. Because of the configuration of the holes in a multi-hole nozzle, the spray cone angle is relatively constant under increasing injection and ambient pressures. Because of the direct fuel injection, the characteristics of the spray produced by a GDI injector are quite similar to those of a diesel spray with regard to the structure and the breakup and atomization processes. The effects of the design and configuration of the holes in a GDI injector on the spray atomization are more significant than the effects of the injection pressure with injection pressure under the 40 MPa. Interaction between each plumes from each holes with high injection pressure is crucial factor to spray behavior and should be sufficiently studied. Multi-hole injectors and spray-guided engines are preferred to swirl injectors and wall-guided engines because of their better emissions characteristics and higher efficiency. The use of a multi-hole injector reduces soot emissions because the more flexible spray pattern prevents impingement of the fuel spray on the piston head. A sprayguided engine exhibits higher effective efficiency because of reduced incomplete combustion losses and fuel film. The limitations of spray-guide that are spark plug wetting and lack of time for mixture formation should be transcended. Because DISI engines emit much more soot than PFI engines, several studies have been conducted on the soot production and oxidation processes. It is agreed that a major source of the soot is the pool fire, which burns the liquid-phase fuel film on the piston head. Much more soot is produced under cold start conditions than under hot start conditions because the fuel film does not easily evaporate from a cold piston surface. Study on spray pattern or injection strategies to resolve fuel film in cold start condition is future research challenges. Reduced NO emissions and higher fuel efficiency can be simultaneously achieved using a stratified combustion strategy. Close timing of the injection and spark increases the mixing rate of the combustion products and surrounding unburned air by increasing the spray momentum during the combustion. Enhancement of the mixing rate reduces NO emissions by decreasing the temperature. However, combustion instability and unacceptable emissions characteristics caused by incomplete combustion remain challenging issues. Spray modeling is used extensively in the numerical analysis of GDI engines because the mixture formation significantly affects the combustion characteristics. The spray breakup and spray–wall interaction have been considered in GDI spray modeling. The modeling of the mixture formation process has focused on the equivalence ratio in the spark plug gap, the in-cylinder spatial distribution of the equivalence ratio, and fuel film formation. Since most spray modeling is don under medium to high load conditions, flash boiling has been modeled and predicted when spray analysis. Future work on GDI spray analysis is modeling of fuel film for the transient condition, multi-component fuel and surface temperature variation. Combustion modeling has been carried out under the WOT condition. The combustion pressure and flame development process have been investigated with regard to the speed and direction of the flame propagation and the mixture homogeneity and turbulence intensity. However, no adequate study of exhaust emissions has been conducted. Compliance with future emissions regulations will require proper prediction of the emissions that result from GDI combustion. PM and PN emissions which are the main emission from GDI engine should be modeled in future. Numerical optimization of various parts of a GDI engine has been undertaken. Piston optimization effectively improved the mixture distribution and turbulence intensity. The port design, spray pattern,

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