diesel blends

Energy Conversion and Management 100 (2015) 300–309

Contents lists available at ScienceDirect

Energy Conversion and Management journal homepage: www.elsevier.com/locate/enconman

Experimental study on fuel economies and emissions of direct-injection premixed combustion engine fueled with gasoline/diesel blends Jiakun Du, Wanchen Sun, Liang Guo ⇑, Senlin Xiao, Manzhi Tan, Guoliang Li, Luyan Fan State Key Laboratory of Automotive Simulation and Control, Jilin University, Changchun 130025, China

a r t i c l e

i n f o

Article history: Received 6 February 2015 Accepted 24 April 2015

Keywords: Premixed combustion Gasoline/diesel blends Emission Combustion phasing Exhaust Gas Recirculation

a b s t r a c t The effects of gasoline/diesel blended fuel composed of diesel fuel with gasoline as additives in volume basis, on combustion, fuel economies and exhaust emissions were experimentally investigated. Tests were carried out based on a turbocharged Common-rail Direct Injection engine at a constant engine speed of 1800 r/min and different loads of 3.2 bar, 5.1 bar Indicated Mean Effective Pressure. Additionally, the effect of combustion phasing and Exhaust Gas Recirculation were evaluated experimentally for various fuels. The results indicated that with the fraction of gasoline increasing in blends, the ignition delay was prolonged and the combustion phasing was retarded with the common injection timing. This led to a significant increase of premixed burning phase, which was in favor of smoke reduction; although, too much gasoline might be adverse to fuel consumption. An optimum combustion phasing was identified, leading to a higher thermal efficiency and better premixed combustion with blended fuels. A combined application of Exhaust Gas Recirculation and blended fuel with a high gasoline fraction was confirmed effective in reducing the oxides of nitrogen and smoke emissions simultaneously at the optimum combustion phasing without giving significant penalty of fuel consumption. A compound combustion mode with its emission lower than the conventional Compression Ignition engines, and efficiency higher than the typical Spark Ignition engines, could be achieved with a cooperative control of Exhaust Gas Recirculation and combustion phasing of the gasoline/diesel blended fuels. Ó 2015 Elsevier Ltd. All rights reserved.

1. Introduction The diesel engines are widely used in various sectors such as agriculture, transportation and industry in regards to the popularity, durability and fuel economy when compared with other types of internal combustion engines. However, in conventional diesel engines, fuel spray burns while mixing with the intake air and, hence, a rich mixture and a high temperature burning area come into being in the cylinder. Consequently, NOx (Oxides of Nitrogen) and PM (Particulate Matter) are generated unavoidably in conventional burning process due to the non-uniformity of the mixture and temperature distribution [1]. Since they are verified linked to a number of health issues for human being, regulations on these emissions from diesel engines are becoming increasingly stringent all over the world [2]. The major objective in researches relating to the diesel engine is to overcome the emission ‘‘trade-off’’: to reduce the NOx and PM emissions simultaneously. To solve this NOx-PM trade-off problem, researchers have made a lot of effort in developing advanced technological systems and ⇑ Corresponding author. E-mail address: [email protected] (L. Guo). http://dx.doi.org/10.1016/j.enconman.2015.04.076 0196-8904/Ó 2015 Elsevier Ltd. All rights reserved.

high efficiency exhaust after-treatments to reduce the pollutant emissions from diesel engines. DPF (Diesel Particulate Filters) have been proven to be highly effective in reduction of PM [3]. Thanks to the catalyst technologies, NOx emissions can be controlled through the use of SCR (Selective Catalytic Reduction) [4]. However, the after-treatment system, which is seemingly effective in settling the problems, are normally very expensive, complicated and will bring down the fuel economy advantages of diesel engine [5]. Therefore, there is a great deal of interest for researchers and engineers to develop some alternative combustion modes of diesel engines to produce low emissions without efficiency penalties. It is widely believed that the future engine will be operated on an alternatively compound combustion mode involving an integration of diesel engine and gasoline engine, taking into account both high efficiency and low emissions [6]. One potential way to reduce simultaneously the NOx and particulates emissions with high engine efficiency is to convert the diffusive combustion to premixed combustion in the CI (compression ignition) engines. Many researchers have shown that HCCI (Homogeneous Charge Compression Ignition) concept is a promising technique for simultaneous NOx and soot reduction with the advantage of thermo-dynamical attraction [7]. However, several technical

J. Du et al. / Energy Conversion and Management 100 (2015) 300–309 Table 1 Key specifications of this engine. Category

Properties

Engine type

Common rail turbo charged with intercooler

Injection

Maximum injection pressure (bar) Fuel injection timing (BTDC) Injection duration (ls)

1600 0–20 380–600

Nozzle

Number of holes Spray angle (°) Orifice diameter (mm)

6 150 0.13

Valve timing

Intake valve open (BTDC) Intake valve close (ABDC) Exhaust valve open (BBDC) Exhaust valve close (ABDC)

24 55 54 26

Bore (mm) Stroke (mm) Displacement (L) Compression Ratio Max torque (N m) @ speed (rpm)

93 102 2.771 17.2:1 260 @ 1800

hurdles need to be overcome before the widespread use of this strategy. HCCI is typically characterized by relatively lean mixtures and undergo auto-ignition dominated by the chemical kinetics of fuel–air mixtures. Thus, HCCI engine lacks a direct control on the auto-ignition timing and combustion rate, which restricts this combustion strategy at limited loads. Therefore, it is necessary to couple the fuel injection event and the combustion event [8]. PCCI (Premixed Charge Compression Ignition) provides the potential of flexible control on ignition and heat release process [9]. In the PCCI combustion, fuels are introduced into the combustion system by early direct injection. This avoids the formation of regions with unfavorable fuel–air ratios that lead to particulates formation. Higher EGR (Exhaust Gas Recirculation) ratios could make the specific conditions in PCCI possible, due to the long ignition delay and low combustion temperature [10]. However, too high EGR ratios can reduce the NOx but will cause increases in soot emission, and meanwhile, the high EGR ratios are very hard to be obtained from the practical engines due to the coupling between exhaust gas status and intake charge pressure [11]. Moreover, this combustion strategy generally suffers from high levels of CO (carbon monoxide) and UHC (unburnt hydrocarbon) emissions, which may related to the fuel spray sticking to the cylinder wall with early injection timing [12]. Similar result has been found in [13]

that the significance of CO and HC is also associated with the higher EGR ratios. It seems that late injection timing has practical advantages in preventing fuel from adhering to the cylinder walls. However, compared to gasoline, diesel fuel has a high reactive (i.e., high cetane number) and a relatively higher boiling point, which means that the mixture can auto-ignite readily, and thus an early injection timing is desired for premixing process. Hence, the diesel fuel might not be an ideal candidate for premixed combustion operation due to its high cetane number (CN) [14]. Suzuki et al. [15] also mentioned that the trade-off relation between NOx and PM emissions depends significantly on average boiling point. Montajir et al. [16] have studied the emission behavior and auto-ignition behavior of mixed normal paraffin fuels. It is believed that the fuel property is an important influential factor for combustion and emissions. Kitano et al. [17] studied systematically the effect of distillation characteristics and cetane number on PCCI combustion. They pointed out that fuel with high volatility was effective in producing lean mixture during the ignition delay. Moreover, the lower cetane number was more powerful in suppressing the ignitability and reducing NOx emissions. Tsujimura and Goto [18] investigated the effects of various fuel properties, such as auto-ignitability, volatility and aromatic hydrocarbon components on combustion performance. They suggested that the poorer auto-ignitability and superior volatility is effective for producing lower soot emission and wider the operation range with ultra low emissions. Further, Bessonette et al. [19] studied the effect of mixed fuels, which developed in the gasoline and diesel boiling range covering a broad range of ignition quality, fuel chemistry, and volatility, on HCCI engine operating range and emissions. In this study it was found that the fuels with medium ignitability between the gasoline and diesel could be the best fuel for HCCI-type operation. Gasoline and diesel mixtures fueled combustion in a multi-cylinder light duty diesel engine was reported in [20]. In this study, it was found that an increased proportion of gasoline fuel reduced the smoke emissions at higher operating loads, as a result of the increased ignition delay and fuel volatility. Experimental results in [21] show that better combustion stability could be obtained with gasoline/diesel blended fuel when comparing with gasoline-fueled HCCI engine. At the University of Birmingham, the investigations that based on the idea of fuel designing through blending gasoline with diesel have demonstrated that the ‘‘dieseline’’ is more suitable for PPCI combustion compared to neat diesel [22]. Similarly, Han et al. [23] have investigated blends of diesel and gasoline up to 40% with cetane number down to 31. They have highlighted that

Injecon Controller

Ambient Air

PC

Diluon System EEPS 3090

Turbocharger

AVL 439

PC

Gas Line Data Line Heat Exchanger

301

EGR Valve Fig. 1. Schematic of test engine configuration.

HORIBA 7100DEGR

302

J. Du et al. / Energy Conversion and Management 100 (2015) 300–309

Table 2 Specifications of the measurement devices. Measurement system

Equipment

Model

Manufacturer

Uncertainties

Dynamic parameters

Dynamometer

CW160

CAMA

Fuel-flow meter

DF-2420

ONO-SOKKI

Torque: ±2 N m Speed: ±2 rpm ±0.2%

Combustion parameters

Pressure transducer Charge-amplifier Crank angle encoder

6052C 5018A 2614B

KISTLER KISTLER KISTLER

±1% ±0.6% ±0.02°CA

Emissions parameters

Exhaust gas analyzer Smoke meter

MEXA-7100DEGR 439

HORIBA AVL

±0.5% FS ±0.2%

Table 3 Main properties of basic fuels used. Fuel

Unit

Method

Diesel

Gasoline

Density (15 °C) Cetane number (CN) Octane number Distillation 50 vol.% Low heating value Elements C H O S

kg/m3

ASTM 4052 ASTM D613

823 52

766 –

°C

ASTM D86

– 272

97 97

MJ/kg

ASTM D4809

42.85

43.90

84.69 13.55 0 0.0030

84.90 15.10 1.8 0.0003

Mass Mass Mass Mass

% % % %

ASTM D5291 ASTM D5453

increasing the gasoline in the blend up to 40% reduces the soot emissions peak particularly and the effect of high injection pressure on soot emissions can be limited. A comparative investigation on Homogeneous Charge Induced Ignition (HCII) and Gasoline/Diesel Blend Fuel (GDBF) by Yu et al. [24] has found that both approaches have the potential for improving the thermal efficiency of gasoline engines and the GDBF mode gives less HC emissions than HCII. On this basis, Yang et al. [25] demonstrated that wide distillation fuels are a promising approach to reduce soot and NOx emissions with providing a high thermal efficiency in low temperature combustion engines. Liu et al. [26] studied the sensitivity of gasoline octane number on combustion and emission characteristics in premixed LTC mode. The results show that the wide distillation fuels can significantly reduce soot emissions and gasoline octane number has little impact on both CO and HC emissions. Many researchers have focused on the effects of gasoline or diesel and injection strategies for the control of ignition under PCCI conditions. A common conclusion could be summarized that high volatility gasoline promotes the mixing process and retards the combustion. But most researches concentrated on the engine combustion and emission performance in fixed injection timing or PCCI operating conditions and few of them have paid close attention to the combustion phasing. It was reported in [27] that the combustion phasing has significant influence on thermal efficiency as well as fuel properties. An improvement in the indicated thermal efficiency could be observed as the combustion phasing approaching the top dead center [28]. Hence it seems that combustion phasing

Table 4 Main properties of basic fuels. Fuels

Unit

Method

G20

G30

G40

G50

G60

Percentage of gasoline Estimated CN Density (15 °C)

v/v %

–

20

30

40

50

60

– kg/ m3

– ASTM 4052

40 810

35 804

31 801

27 798

24 793

is a noteworthy feature for premixed combustion. Furthermore, an appropriate gasoline fraction (or fuel properties) as well as suitable controlling methods for the premixed combustion are not clear yet. Generally, if the mixing process can be accelerated, the fuel can have longer time to mix with the air before ignition taking place, although the time between injection and ignition is still extremely short. Thus, a combined module of PCCI and conventional Direct Injection (DI) combustion is proposed in details in this study. The effect of gasoline/diesel fuel blends, composed of diesel, and gasoline in various volume fractions as additive, on fuel economies and exhaust emissions is experimentally investigated using a diesel engine at a constant engine speed and with different loads. A reasonable gasoline fraction which is able to produce high efficiency premixed combustion in a diesel engine with moderate injection timings was also proposed. In order to reduce the NOx together with the smoke emissions, and at the same time keeping the high fuel efficiency, the combustion phasing and EGR are considered as a whole to optimize the combustion process achieved using an appropriate gasoline blend ratio. 2. Experimental section 2.1. Experimental engine and test equipment The main objective of this research is to investigate how the fuel component, combustion phasing, and EGR fraction affect the auto-ignition processes and emissions of a diesel engine with premixed combustion near the top dead center. A four-stroke, four-cylinder, direct-injected turbocharged diesel engine with a displacement of 2.771 L was used for the investigation. Table 1 shows the key specifications of this engine. The engine was fitted with an adjustable high-pressure loop EGR system, including a cooler that allows the EGR gas temperature to be controlled. In order to achieve the correct amount of EGR, the exhaust gas flow through the EGR system to the intake system was controlled with an electronic EGR valve. The engine was also equipped with a Common Rail injection system managed by a fully opened electronic control unit, the ECU (Electronic Control Unit), in order to command the fuel injection strategies precisely. The ECU was allowed to control the injections number, the injection pressure and the energizing time. The injection timing was determined by actuating on the SOE (Start of Energizing) command and fixed constant for different fuels. The actual SOI (Start of Injection) occurred at about 3 °CA after SOE under the above conditions. Fig. 1 shows a schematic diagram of experiment. A Kistler 2614B crank angle encoder was connected to the crank-shaft to collect crank angle signals. Crank angle-resolved cylinder pressure was obtained with a Kistler 6052C piezoelectric pressure transducer which located in the glow plug adapter of the engine head. A Kistler 5018A charge-amplifier and a DS-9100 combustion analyzer were adopted for analyzing the combustion characteristics. For each operating condition investigated, the cylinder pressure was recorded at 0.25 CAD (Crank Angle Degree) increments and

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J. Du et al. / Energy Conversion and Management 100 (2015) 300–309

8

1800 r/min

G0

6XKYY[XKA36GC

7 IMEP = 3.2 bar 6 SOE = 11eCA BTDC

G20 G30

5

G40

4

G50

3

G60

2 1 0 -25

-15

-5

5

15

25

6XKYY[XKA36GC

8

5 4 3 2 1 0 -25

35

G0 G20 G30 G40 G50 G60

1800 r/min

7 IMEP = 5.1 bar 6 SOE = 11eCA BTDC

-15

-5

)XGTQ'TMRKAJKMC

5

15

25

35

)XGTQ'TMRKAJKM ]

Fig. 2. In-cylinder pressure history, 3.2 bar and 5.1 bar IMEP, 1800 r/min.

92

1800 r/min IMEP = 3.2 bar SOE = 11eCA BTDC

G20 G30

72

G40

52 32 12 -8 -25

G50 G60

-5

152

G0

5

8U.8A0q)'C

8U.8A0q)'C

112

1800 r/min

G0

132 IMEP = 5.1 bar 112 SOE = 11eCA BTDC

G20 G30

92

G40

72

G50

52

G60

32

-5

5

12 -15

-5

5

15

25

-8

-25

-15

)XGTQ'TMRKAJKMC

-5

5

15

25

)XGTQ'TMRKAJKMC

1200

20 19 18 17 16 15 14 13 12 11 10

NOx [ppm]

3.2bar 5.1bar

1000

3.2 bar

800

5.1 bar

600 400

1.6

8 6 4 2 48

0

2 1.2 0.8 0.4 0

350 300 250 200 150 100 50 0

2400 2000

46 42 40

1600

38 G0

G20

G30

G40

G50

G60

1200

ITE [%]

44

36

800

CO [ppm]

10

0

BD [deg]

16 14 12 10 8 6 4 2 0

Smoke Opacity [%]

200

HC [ppm]

CA50 [deg ATDC]

ID [deg]

Fig. 3. Rate of heat release, 3.2 bar and 5.1 bar IMEP, 1800 r/min.

400 G0

G20

G30

G40

G50

G60

0

FUEL

FUEL Fig. 4. Combustion characteristic metrics with respect to the gasoline blend ratio, 3.2 bar and 5.1 bar IMEP, 1800 r/min.

ensemble-averaged over 100 consecutive combustion cycles. The emissions analysis system was comprised of an AVL 439 opacimeter and a HORIBA MEXA 7100DEGR exhaust gas analyzer. Particle

Fig. 5. Exhaust emission characteristics with respect to the gasoline blend ratio, 3.2 bar and 5.1 bar IMEP, 1800 r/min.

size distribution in the exhaust gas was measured by TSI EEPS 3090. In addition, an ONO-SOKI DF-2420 fuel flow meter was used for measuring the fuel consumption rate. The main Specifications of the measurement devices used in tests are presented in

304

J. Du et al. / Energy Conversion and Management 100 (2015) 300–309

Table 2. In this paper, figures show error bars representing 95% confidence.

2.2. Test fuel As stated before, the aim of this study is to characterize the combustion and emission performance of the mixture made from gasoline and diesel fuel with specific volumetric fraction. Compared to diesel, gasoline has some advantages that may be of particular interest for mixing preparation within the combustion bowl, such as its higher volatility, lower cetane number and higher heating value. In this research, an ultra-low sulfur (<10 ppm) diesel with a cetane number of 51 was chosen as reference fuel, while a high-quality gasoline with a octane number of 97 was selected as the test fuel. The main properties of the basic fuels used for the test are listed in Table 3. In total, six blends with different volume fraction of gasoline were used which is defined in a format of G-XX according to the gasoline volumetric percentage of each fuel. The XX represents the volumetric percentage of gasoline in the blends. Table 4 shows a summary of the main properties of each fuel. Further, it is worth pointing out that the CN of blended fuels were estimated value which was computed using Kay’s mixing rule [29]. Although, the estimation of CN using the above method might give some errors against the true value, especially for the low CN fuels (or high gasoline fraction fuels); the empirical assessing method used in this research was still thought to be suitable for presenting the fuel auto-ignitability within a wide range between gasoline and diesel.

2.3. Test conditions

RoHR [J / deg CAD]

In this study, early injections at a moderate level near the TDC (Top Dead Center) were applied. The tests aimed to understand the effect of combustion phasing (CA50) and EGR on engine out emissions and combustion performance of the gasoline/diesel blends at different loads. The injection pressure remained constant for all tests at 800 bar. The investigation was carried out at 3.2 bar and 5.1 bar IMEP (Indicated Mean Effective Pressure) with constant engine speed of 1800 r/min. The engine coolant and lubricant temperature were kept in constant at 85 °C to minimize the discrepancies of the testing results. Before a new fuel being applied, the residual fuels from the last test left in the fuel supply line and fuel flow meter were completely drained by operating the engine under a heavy load for at least 30 min to prevent the undesired mixing of different fuels.

194 174 0 CA ATDC 154 134 114 94 74 54 34 14 -6 -10 0

G0

In this paper section, the results concerning of combustion and emission will be presented below for the baseline diesel fuel (G0) and the gasoline blended fuels including, G20, G30, G40, G50 and G60. Each of the parameters assessed will be compared to the pure diesel. Firstly, the effect of the gasoline volumetric fraction on combustion and engine out emissions were examined and presented. Secondly, results of fuel economy and emissions for the six fuel blends were illustrated for the CA50 sweep. Finally, the cooperative controlling strategies for the premixed combustion process optimization with suitable gasoline/diesel blended fuel were presented. 3.1. Effect of injection pressure on combustion characteristics and specific emissions In this section, the combustion characteristics and the results of engine exhaust emission obtained with a single injection of the test blends were compared to each other for analysis. The effects of the fuel properties of the diesel/gasoline blends on ignition and combustion under constant injection timing were also investigated. Figs. 2 and 3 show the indicated pressure and RoHR (Rate of Heat Release) at 1800 rpm for all the fuels at different loads. The pressure curves were the averaged value of 100 individual pressure traces. These figures highlight that a quasi-premixed-combustion can be produced with a higher gasoline fraction. A two-stage heat release pattern was observed with the gasoline fraction increased. The first stage of heat release was associated with the low-temperature-reactions, and the second stage was with the high-temperature-reactions. The higher the gasoline fraction, the more distinct the first stage peak, due to the lower global fuel reactivity. In the meantime, the difference of combustion performance of various gasoline/diesel blends are found, as that, the gasoline is slower than the diesel to reach the peak pressure. The combustion phasing is retarded with the increase of gasoline volumetric fraction for a given SOI under the same operating condition. The conclusion is actually in consistence with [24]. Results of Fig. 3 also emphasize that there was a significant difference in RoHR for the blends investigated, that is, the peak of RoHR decreased with the increasing volume fraction of the gasoline at 3.2 bar; but it increased at 5.1 bar. This difference should be related to the distinction of engine operating conditions or other combustion parameters for these works. Well premixed combustion status could be achieved at 3.2 bar with diesel due to the high air– fuel-ratio under the light load. However, the mixture may be over-lean if too much gasoline was blended into the fuel, and it will be hard to burn completely under a light load condition, due to the

G20

SOE / eBTDC G0 = 15.2 G20 = 15.7 G40 = 16.9 G60 = 18.7

10

3. Results and discussion

SOE / eBTDC G0 = 10.7 G20 = 11.3 G40 = 12.0 G60 = 14.4

5 CA ATDC

-5

G40

5

15

10 CA ATDC

0

10

G60

SOE / eBTDC G0 = 6.0 G20 = 6.8 G40 = 7.7 G60 = 10.4

20

15 CA ATDC

5

SOE / eBTDC G0 = 2.0 G20 = 3.0 G40 = 4.7 G60 = 7.8

15

Crank Angle [deg CAD] Fig. 6. Rate of heat release (RoHR) history with various gasoline–diesel blends, Pinj = 80 MPa, EGR = 0%, CA50 = 0, 5, 10, 15 CA ATDC.

25

305

J. Du et al. / Energy Conversion and Management 100 (2015) 300–309

50

G0 G20

48

G40

46

G60

ITE [%]

LP-RoHR [deg ATDC]

18 16 14 12 10 8 6 4 2 0 -2 -4

G0

G20

G40

G60

44 42 40

IMEP = 5.1 bar EGR = 0% -5

0

5

10

15

38 20

36 -5

IMEP = 5.1 bar EGR = 0 % 0

5

10

15

20

CA50 [deg ATDC]

CA50 [deg ATDC] Fig. 7. Location of peak RoHR versus CA50, 5.1 bar, 1800 r/min.

Fig. 9. Indicated thermal efficiency (ITE) for a CA50 sweep, 5.1 bar, 1800 r/min.

low combustion temperature and the high volatility of the gasoline fuel. Fig. 4 shows the comparison of the combustion characteristics (Ignition Delay, Burning Duration, CA50 and Indicated Thermal Efficiency) of the fuels under different operating conditions with IMEP = 3.2 bar and 5.1 bar, respectively. In this paper, the Start of Combustion (SOC) was defined as the 10% crank angle and the End of Combustion (EOC) as 90% of the total released heat; the Ignition Delay (ID) as the crank angle difference between the start of the injection and SOC; the Burning Duration as the crank angle difference between SOC and EOC. The Indicated Thermal Efficiency (ITE) is defined as the ratio of the indicated work per cycle to the amount of the fuel energy supplied per cycle. During the test, the fuel energy is calculated using the mass and low heating value of basic fuels according to the volumetric percentage of gasoline in the blends. The higher gasoline fraction produces, as expected, a significant extension of ID and a reduction of Burning Duration (BD), for any operating condition. In particular, the decrement of BD varies between the two loads, and the greater effect of gasoline on BD is found under the higher load. The low cetane number of gasoline prolongs the ID and retards the combustion phasing (CA50), giving more time for mixing and making the local rich regions within the combustion chamber thinner. In addition, the higher volatility of the gasoline promotes the dispersion of the fuel vapor within the combustion chamber, further enhancing the mixture preparation. However, the desired premixed combustion will be accompanied with a negative effect, as illustrated in Fig. 4, that there is a decrease in combustion efficiency which might cause an increase of fuel consumption. In particular, the blended fuel with a gasoline fraction over 40% results in a low efficiency combustion pattern. This may be related to the too lean local equivalence ratio (over-mixing), which leads to an incomplete combustion on relevant time scales.

Fig. 5 shows the effects of gasoline blends on the exhaust emission characteristics, such as NOx, smoke, HC, and CO emissions. As shown in the figures below, with the fixed injection timing, the NOx emissions changes slightly with the increasing gasoline volume fraction. However, the smoke level decreases as the gasoline fraction increased, due to the higher-volatility-promoted mixing process. This suggests that the premixed combustion could be achieved by using gasoline/diesel blended fuels to enhance the mixing process, despite the injection timing is relatively late. In other words, this may be useful to reduce the requirement for higher EGR ratio under early injection timings. Both HC and CO emissions exhibit similar trends for a sweep of the gasoline fraction, which is a monotonously increase with the fraction of the gasoline. Similar trend was observed in [30] and the reason could be related to the extended ignition delay, which produced more over-mixing regions [31]. The above two emissions also increase as the gasoline fraction greater than 40% and the results might correspond to the former combustion characteristics showed in Fig. 4.

G0

G20

G40

G60

CE [%]

100 95 90 85 80

IMEP = 5.1 bar EGR = 0 %

75 70

-5

0

5

10

15

CA50 [deg ATDC] Fig. 8. Combustion efficiency for a CA50 sweep, 5.1 bar, 1800 r/min.

20

3.2. Fuel effects at matched combustion phasing This section illustrates the fuel economy performance and the engine out emission against CA50 produced by each tested fuel sprayed with a single injection strategy. Generally, the crank angle, CA50, is usually used as the standard indication of combustion phasing, the relative position of the combustion within one cycle. It also has been proven as an effective parameter for comparing different fuels [32]. To eliminate the effects of combustion phasing between fuels at the baseline operating condition, the CA50 was maintained constant, within ±0.25 CAD, by adjusting injection timing to establish the same baseline point with each fuel. In this paper, the combustion phasing, or CA50, was maintained constant for all the tested fuels at 0, 5, 10, 15° ATDC, respectively. The above results show that the combustion and emission parameters change monotonously with the fraction of the gasoline. In order to highlight the effect of combustion phasing on combustion and emission characteristics distinctly, four typical fuels were chosen for further study. Fig. 6 shows the RoHR history for the CA50s with four typical tested fuels. For all the CA50 cases tested in this research, although the combustion phasing is maintained constant for each fuel, the mixing time plays a major role in how much over-leaning occurring in the cylinder prior to combustion. In other words, the high volatility of the gasoline plays a role in promoting premixed combustion. In addition, a quasi-premixed combustion occurs at the retarded CA50 and the advanced ones are dominated by the premixed combustion followed by a mixing-controlled combustion for diesel, which is in consistence with Tsujimura and Goto [18]. This is attributed to the relatively longer premixing time caused by the milder combustion process

J. Du et al. / Energy Conversion and Management 100 (2015) 300–309

210 G0

205

G20

G40

G60

ISFC [g/kw.h]

200 195 190 185 180 175

IMEP = 5.1 bar EGR = 0 %

170 165 -5

0

5

10

15

20

CA50 [deg ATDC] Fig. 10. Indicated specific fuel consumption (ISFC) for a CA50 sweep, 5.1 bar, 1800 r/min.

as a result of the retarded injection timing. However, the combustion processes are dominated by the premixed combustion of the gasoline–diesel blends at any CA50 case, especially G40 and G60. Fig. 7 illustrates the location of the peak RoHR (LP-RoHR) as a function of CA50 for the typical fuels used in this study. The lack of mixing controlled combustion makes the CA50 linearly related to the LP-RoHR. This could be due to the premixed combustion event gives a single, sharp heat release mode. In this case, the RoHR curve is similar to isosceles triangle. Thus, as seen in this figure, a well premixed combustion could be achieved with G40 and G60, respectively. Fig. 8 shows the combustion efficiency (CE) for a CA50 sweep at the 5.1 bar IMEP. All the fuels exhibit similar trends for various combustion phasing and displaying a maximum value of ATDC CA50 at around 5–10°. The disadvantage of advancing the combustion phasing, in order to achieve a well premixed mixture, is that there is a decrease of combustion efficiency that causes an increase of the fuel consumption, as illustrated in Figs. 8–10. In addition, G20 does not give significant influence to the CE; and a dramatic decrease occurs when the gasoline fraction goes up to 60%. The effect of

gasoline–diesel blends on the indicated specific fuel consumption (ISFC) and the ITE against the combustion phasing are presented in Figs. 9 and 10, respectively. It is observed that there is a significant difference in ITE and ISFC at the CA50 condition produced by each fuel. An optimal combustion phasing for diesel of which the maximum ITE is 46% could be achieved within 5° ± 0.25 ATDC, whereas the corresponding value for G20, G40, G60 are found to be 45.6%, 44.7% and 43.4%, respectively. Additionally, it is observed that the gasoline–diesel blends increases the ISFC in different degrees. For all the typical blends, the ITE decreases by 0.4, 1.3 and 2.6 percentage points for G20, G40 and G60 for 5° ATDC CA50 condition relative to the baseline diesel. The ISFC increases by 0.9%, 3.6% and 5.9% respectively at the same condition, for the above fuels. Generally, the increment in ISFC becomes larger as the gasoline fraction increased. In other words, the engine consumes more fuel with gasoline–diesel blends than with baseline diesel fuel to generate the same work. This could be related to the incomplete combustion through the expansion stroke [33].

20

12 10

0

5

10

15

20

25

G60

0.8 0.6 0.4

IMEP = 5.1 bar EGR = 0 %

-5

0

5

10

15

G0

1.4

1

G40

G60

1.2 1 0.8 0.6 0.4

IMEP = 5.1 bar EGR = 0 %

0.2 0

20

G20

-5

0

5

10

CA50 [deg ATDC]

CA50 [deg ATDC]

(a) NOx

(b) Smoke

15

20

3 G0

G20

G40

G60

G0

2.5

IMEP = 5.1 bar EGR = 0 %

G20

G40

G60

IMEP = 5.1 bar EGR = 0 %

2 1.5 1 0.5

-5

0

5

10

15

20

35

Fig. 12. Ignition delay (ID) for each fuel at varying EGR ratios, 5.1 bar IMEP, 1800 r/ min.

Smoke Opacity [%]

G40

30

EGR Rao / %

CO [h 103 ppm]

NOX [h 103 ppm]

G20

1.2

0.2

HC [h 103 ppm]

16

1.6 G0

1.4

1 0.9 0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1 0

18

14

1.6

0

G0 G20 G40 G60

22

ID [deg]

306

0

-5

0

5

10

CA50 [deg ATDC]

CA50 [deg ATDC]

(c) HC

(d) CO

Fig. 11. Emission characteristics for a CA50 sweep, 5.1 bar IMEP, 1800 r/min.

15

20

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20 18

SOE [deg BTDC]

and combustion incompleteness. The larger amount of the products of the incomplete combustion for gasoline–diesel blends could be explained by the fact that a longer ignition delay for a specified CA50 will lead to more significant over-leaning in the cylinder.

G0 G20 G40 G60

19 17 16 15

3.3. Effect of EGR on combustion and emission

14 13 12 11 10

0

5

10

15

20

25

30

35

EGR Rao / % Fig. 13. SOE for each fuel at varying EGR ratios, 5.1 bar IMEP, 1800 r/min.

The above explanation can be partially verified by Fig. 11 in which the four major emission pollutants are shown over the various CA50 cases. Obviously, the NOx decreases monotonically with the retarded CA50, since the combustion temperatures go down with the retarded combustion phasing, whereas the smoke increases to a certain level according to the fuel used. However, slight changes in smoke opacity are found when the gasoline fraction goes up to more than 40%. It therefore could be concluded that the higher volatility and the longer mixing time offer a potential approach to strike a balance between NOx and smoke emissions. Differences did appear among the tested fuels when comparing HC and CO versus CA50. An inflection point occurs at CA50 in advance of more than 5° ATDC and has a sharp increase after 10° BTDC as given in Fig. 11 (a). This could be explained by the fact that under the earlier CA50 condition, the fuel sprays will miss the piston bowl and impinge on the piston surface, which leads to incomplete combustion. However, when the CA50 is retarded to more than 10° BTDC, the combustion will be pushed further back into the expansion stroke and more unburned fuel produces. At these conditions, unlike the HC, the CO increases monotonically with the retarded CA50. And as expected, the higher gasoline fraction, the more escape of the unburned fuel due to the over-leaning

In the diesel compression ignition mode, EGR is a very effective technique to reduce the NOx emissions due to its effects in reducing the temperature and oxygen atom concentration. Unfortunately, further increases in EGR tend to increase the soot formation; this is because that the lack of oxygen would lead to the formation of the soot in rich reaction area. However, the soot formation would be inhibited by creating a better premixed mixture or increasing the in-cylinder temperature in the late cycle. The effects of EGR on engine out emission and fuel economy performance with the four typical fuels will be explored in detail in this section. As mentioned in the previous section, the CA50 has been fixed in order to ensure a controllable higher efficiency operation condition. In order to eliminate the impact of fuel density and low heating value on combustion heating release, the total heat obtained from the whole burning process has been fixed for all the tested fuels by adjusting the quality of fuel injected. Fig. 12 illustrates the effects of EGR on ignition delay at 5° ATDC CA50. Due to the different auto-ignition quality of tested fuels, CA50 is fixed by adjusting the SOE timing under various EGR ratios. Fig. 13 shows the corresponding SOE values for these four typical fuels. Both gasoline fraction and EGR have a notable impact on ignition delay. The ID increases as the EGR levels are increased; and the increment became larger when more EGR is introduced. For any of these blends, an apparent increase of ID could be observed with EGR ratio higher than 23%. This could be related to the fact that at a higher EGR ratio, the auto-ignition reactions are slowed down by the reduced oxygen level and lower in-cylinder temperature. And thus, the ID would become more sensitive at a larger EGR level, especially with the fuel of a higher gasoline fraction.

0.8

2.5

Smoke Opacity [%]

NOx [h 103 ppm]

G20

0.6

G40

0.5

G60

0.4 0.3 0.2 0.1 0

0

5

10

15

20

25

30

G40

1.5

G60

1 0.5 0

5

10

15

20

EGR Rao [%]

EGR Rao [%]

(a) NOx

(b) Smoke 3 G0

0.7 0.6

G40

0.5

G60

0.4 0.3 0.2

25

30

35

25

30

35

G0

2.5

G20

CO [h 103 ppm]

HC [h 103 ppm]

G20

2

0

35

0.8

G20 G40

2

G60

1.5 1 0.5

0.1 0

G0

G0

0.7

0

5

10

15

20

EGR Rao [%]

(c) HC

25

30

35

0

0

5

10

15

20

EGR Rao [%]

(d) CO

Fig. 14. Emission characteristics for each fuel at varying EGR ratios, 5.1 bar IMEP, 1800 r/min.

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50

210

G0

48

G20

ISFC [g/kw h]

ITE [%]

46 44 42

Thermal Efficiency of Typical SI engines

40

G0 G20

38

G40

36 34

G60

0

5

10

15

20

25

30

35

200

G40 G60

190 180 170

0

5

10

EGR Rao [%]

(a) Indicated thermal efficiency (ITE)

15

20

25

30

35

EGR Rao [%]

(b) Indicated specific fuel consumption (ISFC)

Fig. 15. Fuel economy characteristics for each fuels at varying EGR, 5.1 bar IMEP, 1800 r/min.

Fig. 14 shows the effects of EGR on four major emission pollutants. As expected, EGR is effective in reducing NOx emissions which can be found decreasing linearly in the tested EGR range. For all the tested fuels, the NOx emissions have been decreased by about 80% or more, which is not in consistence with [34]. With regard to the smoke emissions, it is significantly suppressed in spite of the high EGR rates when the gasoline fraction is higher than 40% due to the fact that the well premixed mixtures could be prepared before the commencement of combustion. However, obvious increases in smoke have been found for the diesel, especially under the high EGR working conditions. This may relate to the high auto-ignition quality and poor volatility of the diesel. As mentioned before, since a CA50 is fixed by adjusting the injection timing, the diesel should be injected later compared to other fuels due to the higher auto-ignition quality and, hence, the mixture uniformity degraded. Furthermore, the lower smoke opacity at the high EGR level is a result of the higher volatility of the gasoline compared to the diesel, and this further leads to much quicker evaporation of the blended fuel and a higher fraction of premixed mixture even in a low-oxygen atmosphere resulting in less rich zones. In addition, the longer ignition delay with blended fuels could further increase the premixed mixture prior to ignition and leads to better utilization of the alongside oxygen. All these promote more premixed combustion which can decrease the smoke opacity. The CO emission following the smoke also increases with the EGR rate. This is due to the partial oxidation of the fuel as a result of the over-lean mixture and low temperatures caused by the longer ID and higher volatility of blendes. Furthermore, only limited oxygen is available for oxidizing carbon monoxide to carbon dioxide under such a high EGR level. The HC, unlike the CO, is relatively low and fairly constant in the tested range but increases slightly for G60 at the high EGR level, which is not in consistence with [35]. Fig. 15 shows the effect of EGR on the ITE and ISFC of the four typical fuels. As shown, the EGR ratio has only slight impact on the thermal efficiency with the fixed CA50 through the full range of the experimental variables. As discussed in previous sections, when compared to the diesel, the gasoline/diesel blended fuel has a relatively lower thermal efficiency and hence a higher fuel consumption rate. It is reported that the thermal efficiency of typical natural gas SI (Spark Ignition) engine is less than 34% [36] and the thermal efficiency of gasoline SI engine is no more than 33% [37]. Hence, the cooperative control of EGR and combustion phasing makes the efficiency as a whole higher than the typical SI engines.

4. Conclusion The effects of gasoline/diesel blended fuels, combustion phasing and EGR on combustion performance and exhaust emission of a

four-stroke, four-cylinder, direct-injected turbocharged diesel engine were studied. The results are summarized as follows: The impacts of the gasoline/diesel blend-ratio on combustion and emissions were investigated using a single injection near the TDC. Comparing to the diesel, the gasoline/diesel blended fuel can extend the ID and retard the combustion phasing (CA50), giving a longer mixing time that reduced the local rich-regions across the combustion chamber. This also resulted in a significant increase of the amount of the fuel burnt during the premixed burning phase, which could be beneficial for the smoke reduction. However, when a gasoline fraction rate of more than 40% was used, the fuel economy turned poorer due to unstable/incomplete combustion. The effects of CA50 and blended ratio on fuel economy and engine out emissions were also investigated in this research. Under late and early combustion phasing, the higher volatility and longer mixing time had shown a potential approach for striking a balance between NOx and smoke emissions. For all the fuels used, an optimal combustion phasing with a maximum indicated thermal efficiency could be achieved after TDC. However, the indicated thermal efficiency decreased for G20, G40 and G60 with the optimal CA50 condition relative to baseline diesel. Correspondingly, the incomplete combustion emissions, such as the HC and CO increased notably when the fuels with a gasoline fraction of more than 40% are used. The ignition delay increased as the EGR levels being increased. The increment became larger with a higher EGR ratio. For the gasoline/diesel blends, the ignition delay was found to be more sensitive when a higher gasoline fraction or a higher EGR ratio was applied. A combined use of EGR and high gasoline fraction blends led to a simultaneous reduction in smoke and NOx emissions at the fixed CA50 without giving any penalty in fuel economy. A compound combustion concept with emission lower than conventional DI engines, and efficiency higher than typical SI engines, could be achieved by the cooperative control of EGR and combustion phasing of the gasoline/diesel fuels. Acknowledgements The authors gratefully acknowledge financial support from the National Natural Science Foundation of China (Project code: 51176064) and Graduate Innovation Fund of Jilin University (Project code: 2014091). References [1] Donaldson K, Tran L, Jimenez LA, Duffin R, Newby DE, Mills N, et al. Combustion-derived nanoparticles: a review of their toxicology following inhalation exposure. Part Fibre Toxicol 2005;2:10. http://dx.doi.org/10.1186/ 1743-8977-2-10. [2] Kagawa J. Health effects of diesel exhaust emissions—a mixture of air pollutants of worldwide concern. Toxicology 2002;181:349–53. http:// dx.doi.org/10.1016/S0300-483X(02)00461-4.

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