combustion and exhaust emissions of a di diesel engine operated ...

Seoul 2000 FISITA World Automotive Congress June 12-15, 2000, Seoul, Korea

F2000A030

COMBUSTION AND EXHAUST EMISSIONS OF A DI DIESEL ENGINE OPERATED WITH DUAL FUEL Masahiro ISHIDA1)*, Jung Jea CHO2) and Toshihiro YASUNAGA2) 1) 2)

Professor, Nagasaki University, Nagasaki 852-8521, Japan Graduate student, ditto

In order to improve the trade-off between smoke and NOx in a DI diesel engine, the effect of pre-mixed natural gas on combustion and exhaust emissions has been investigated experimentally. The effects of the CNG charging rate and the EGR rate on ignition and combustion rate are mainly examined, in addition, the effects of CNG and EGR on NOx reduction was analyzed from the viewpoint of water content in the EGR gas. As results, it is found that, (1) the burning rate of natural gas pre-mixture is larger under the higher temperature condition, thus, the larger burning rate results in shortening the combustion duration and leads to lower fuel consumption, (2) ignition and the burning rate of natural gas are suppressed by EGR, (3) significant NOx reduction is obtained by charging natural gas under the low load condition, and (4) the NOx reduction rate due to EGR is about twice of the one due to the port water injection. Key Words: Diesel Engine, Combustion, Natural Gas, EGR, Exhaust Emissions

INTRODUCTION

of EGR and preheating on exhaust emissions such as NOx, smoke and the unburned hydrocarbon were also measured. Especially, the effect of EGR on NOx reduction was compared with the effect of port water injection on NOx reduction taking into consideration of the extended Zel’dovich mechanism. To obtain better improvements in the trade-off between smoke and NOx and the trade-off between fuel consumption and NOx, it is found that an appropriate combination between the natural gas charging rate, the EGR rate and the preheating is required for each engine-load condition.

To meet increasingly stringent emission standards in automotive engine field, extensive research has been carried out to explore various ways to reduce NOx and particulate emissions from diesel engines. In various methods for reducing both exhaust NOx and smoke, utilizations of natural gas as one of the alternative fuels and the oxygenated fuels have been examined (Kusaka, et al.[1], Miyamoto, et al.[2]), and marked reductions in smoke and/or NOx were obtained. Methanol, one of the oxygenated fuels having the low evaporation temperature, is also expected to be another alternative fuel for low emission vehicles. In order to reduce smoke and NOx simultaneously in DI diesel engines, it is a key to successfully burn fuels having a low evaporation temperature and a low cetane index, such as methanol, gasoline as well as natural gas, igniting by means of a small amount of ordinary gas oil having the good ignitability but the high evaporation temperature (Ishida, et al.[3]). The target of this study is to reduce NOx and smoke simultaneously and also to improve the trade-off relationship between smoke and NOx markedly by means of an appropriate combination of the pre-mixture of natural gas, the exhaust gas recirculation “EGR” and preheating of the intake premixture. In the present experiment, combustion test of the premixed natural gas was conducted in a four cylindered naturally aspirated DI diesel engine under both low and high engineload conditions by changing the rate of natural gas charged homogeneously into the suction air, simultaneously changing the amount of gas oil injected directly into the cylinder. Secondarily, the respective effect and the combination effect of preheating of the intake pre-mixture and the exhaust gas recirculation “EGR” rate on ignition and burning rate of the pre-mixed natural gas were investigated, in which the effects * [email protected]

EXPERIMENTAL APPARATUS The test engine was a four-cylinder high-speed naturally aspirated direct injection diesel engine for automobiles, which was the type 4JB1 manufactured by ISUZU Motors Ltd.; 93 mm bore, 102 mm stroke, compression ratio of 18.2 and a maximum output of 64.7 kW(88 PS)/3,600 rpm. The special VE-type fuel injection pump was used to adjust the fuel injection timing arbitrarily and easily. The test injector was the conventional multi-hole nozzle having four holes of 0.28 mm diameter, and the opening pressure of the needle valve was set at 18.5 MPa. Furthermore, the port water injection system was built up on the intake manifold using gasoline injectors with the injection pressure of about 0.5 MPa, and water was injected into each suction port of a four cylinder engine (Ishida, et al.[4]). In the dual-fuel engine system shown in Fig.1, a gas mixer was installed in the intake system for obtaining a complete mixing of natural gas and/or the EGR gas with the intake fresh air, and the pre-mixture was ignited by diesel sprays injected directly into the cylinder. In this system, natural gas was stored in the vessel as a compressed natural gas “CNG” at the high pressure of about 25 MPa. CO2 concentration was measured 1

Air Thermocouple Heater

Digital Thermometer

RESULTS AND DISCUSSION

05 40.0℃

EFFECT OF “CNG” RATE ON COMBUSTION

Valve Manometer

Figures 2(a) and (b) and Figs.3(a) and (b) show changes due to the CNG charging rate in time histories of the incylinder pressure “P (MPa)”, the apparent heat release rate “dQ/dθ (J/deg)” and the needle valve lift “Lift (mm)” which indicates the injection duration of gas oil. Figs.2(a) and (b) are the cases of two different temperature of intake premixture in the low load, and Figs.3(a) and (b) are those in the high load. In the cases of low load, as the CNG charging rate increases, ignition delay increases, however, the initial combustion heat release rate decreases because the amount of the injected gas oil decreases, and the natural gas ignited by injection of gas oil burns very slowly based on flame propagation, which results in a marked decrease in the maximum combustion pressure. The heat release rate in the combustion of natural gas is slightly lower in the low premixture temperature case than in the high temperature one. This seems to be due to low reaction rate resulting in slow flame propagation under the low temperature condition in the pre-mixed natural gas. On the other hand, in the cases of high load, ignition delay tends to decrease especially in the case of high premixture temperature shown in Fig.3(b). The initial combustion heat release rate is almost unchanged due to the CNG charging rate in the high load cases except for the case of xCNG=0.024 alone under the low intake temperature condition. As seen in Figs.3(a) and (b), the natural gas seems to be ignited after the peak of the initial combustion, and the natural gas burns in a short time based on the fast flame propagation in comparison with the low load cases, which results in a marked increase in the maximum combustion pressure. Obviously the burning rate, that is reaction rate, is larger under the higher temperature pre-mixture condition, thus, the large burning rate results in shortening the combustion duration markedly and leads to low fuel consumption. Why the natural gas burns so fast in the high loads ? The significant difference in combustion behavior between the low and high loads might be caused by two factors; one is the temperature of premixture in the combustion chamber as mentioned in the above, and another one is the number of flame kernels. The maximum equivalence ratio of natural gas in the present experiment was selected to be nearly equal in both low and high loads, that is, φCNG=0.39 for the low load and 0.40 for the high load, which is calculated by the relation of φCNG=16.86 xCNG. The higher load was, therefore, attained by injecting a larger amount of gas oil, that is, φGAS OIL=0.04 for the low load and 0.39 for the high load in the present experiment. The amount of injected gas oil is about one tenth that of the natural gas in the low load case and the ratio of the injected gas oil and the charged natural gas is nearly equal to unity in the high load case. The pre-mixture of natural gas in the cylinder is ignited by many flame kernels which are generated by ignition of the injected gas oil sprays, that is, the multi-point ignition occurs in the pre-mixture because each gas oil droplet might play a role of ignition point. Comparing the two cases between

Flow Meter Press. Regulator Valve

Valve Mixing Chamber "TIN" CNG Press. Vessel

Valve Test Engine

Throttle CO2 Meter

Exhaust Gas

Fig.1 Natural gas supply system and EGR system Table 1 Composition of tested natural gas (13A) CH4

C2H6

C3H8

87.65

7.22

1.65

C4H10 C5H12 3.30

0.05

H2, N2, CO2 0.13 (%)

to calculate the EGR rate at the intake manifold and at the EGR return channel respectively as shown in Fig.1 The charging rate of natural gas in the intake fresh dry air, defined by xCNG [kg/kg], was varied from 0 to 0.024 which corresponds to the equivalence ratio of about 0.40, and the EGR rate of xEGR [kg/kg], defined by the total intake charge, was changed from 0 to 0.18. The test engine was operated at a constant speed of 1,700 ± 5 rpm under two kinds of mean effective pressure condition; Pme=0.33 MPa for the low load and 0.66 MPa for the high load. The suction air temperature at the engine inlet “TIN” was regulated at 40 or 80 ± 0.5°C by the electric heater even in the case with EGR. The suction air pressure at the engine inlet was also adjusted and fixed at the standard atmospheric condition by using the motor-driven blower. The composition of tested natural gas is shown in Table 1, which is the urban gas fuel named “13A”, and the gas oil for ignition is a ordinary one having a cetane index of 57. The net calorific values of these fuels are 49.55 and 42.91 MJ/kg respectively. Furthermore, fuel consumption “be (g/ kWh)” shown in the figures is indicated by the reduced one based on the net calorific value of gas oil. The time histories of combustion pressure, fuel nozzle pressure and the needle valve lift were measured using the respective sensors, and those outputs were sampled every one-fourth degree of crank angle simultaneously by means of the four-channel combustion analyzer “CB-467” manufactured by Ono Sokki Co. Ltd. The measured time histories of the experimental results are the ensemble average values sampled over continuous 350 engine cycles. Those data were transmitted to the personal computer and recorded on floppy disks. 2

6

60

2

40 20 0.0 0.010 0.015 0.019 0.023

XCNG [kg/kg]

-20 -10

0

10

20

30

40 20

P [MPa]

60

2

0 0.0 0.011 0.016 0.021

XCNG [kg/kg]

0

10

20

30

40

10

20

30

40

50

CA [deg]

6 80

4

60

2

40

0 Lift [mm]

P [MPa]

4

0

(a) TIN=40°C

8

0 Lift [mm]

-20 -10

80

-20 -10

0 0.0 0.011 0.015 0.019 0.024

XCNG [kg/kg]

50

6

0.3 0.2 0.1 0

20

CA [deg]

(a) TIN =40°C

8

40

40

0.3 0.2 0.1 0

Lift [mm]

0

0.3 0.2 0.1 0

60

2 0

dQ/dθ [J/deg]

Lift [mm]

0

80

4

20 0

0.3 0.2 0.1 0

0.0 0.012 0.017 0.022

XCNG [kg/kg]

-20 -10

50

dQ/dθ [J/deg]

80

4

dQ/dθ [J/deg]

6

P [MPa]

8

dQ/dθ [J/deg]

P [MPa]

8

0

10

20

30

40

50

CA [deg]

CA [deg]

0

1

T (°C ) IN

0.8

40

0.6

80

1

0.4 0.2

0.5 1.5

0

0

1.2 1 0.8 0.6

1

T (°C) IN

0.4

40 80

0.5 0

1.5 be / be0

be / be0

25 1.4

Smoke / Smoke0

NOx / NOx0

1.2

50

1

Smoke / Smoke0

50

NOx / NOx0

100

THC / THC0

(b) TIN=80°C Fig.3 Change in combustion histrory due to CNG rate (Pme=0.66 MPa, θinj=5°BTDC)

THC / THC0

(b) TIN=80°C Fig.2 Change in combustion histrory due to CNG rate (Pme=0.33 MPa, θinj=5°BTDC)

1 0.5

0.5 0

0.01

0.02

0

0.03

0.01

0.02

0.03

XCNG[kg/kg]

XCNG[kg/kg]

(b) Pme=0.66MPa (a) Pme=0.33 MPa Fig.4 Change in fuel consumption and exhaust emissions due to CNG rate (θinj=5°BTDC, TIN =40, 80°C)

the low and high loads, the number of gas oil droplets is much more in the high load than the low load, thus, the combustion rate of natural gas seems to be much larger in the high load case compared with the low load case. Figures 4(a) and (b) show the effect of the CNG charging

rate on exhaust emissions and fuel consumption, where the parameter is the intake pre-mixture temperature. Each ordinate is normalized by the measured value obtained under the zero CNG condition. In both cases of low and high loads, smoke was reduced remarkably by about 80% by the CNG 3

4 XCNG = 0 T IN = 40°C

θinj = TDC

0.4

Smoke [Bosch]

Smoke [Bosch]

0.5

θinj = 5°BTDC

0.3 0.2

θinj = 10°BTDC

XCNG = 0

0.1 XCNG = 0.023

0

XCNG = 0.01 XCNG = 0.015 XCNG = 0.019

3 2

θinj = TDC XCNG = 0 θ inj = 5°BTDC XCNG = 0.011

1 0

XCNG = 0.023

be [g/kWh]

XCNG = 0.015 XCNG = 0.01 θ inj = TDC

250

θinj = 5°BTDC

θinj = TDC X CNG = 0

XCNG = 0.019

300

θinj = 10°BTDC

XCNG = 0.015 XCNG = 0.019 XCNG = 0.024

240

350

be [g/kWh]

XCNG = 0 T IN = 40°C

θinj = 10°BTDC

220

200

XCNG = 0 .011 XCNG = 0.015

θinj = 5°BTDC θinj = 10°BTDC

X CNG = 0.019 X CNG = 0.024

XCNG = 0

180 0 2 4 6 4 6 8 10 NOx [g/kWh] NOx [g/kWh] (a) Pme=0.33 MPa (b) Pme=0.66MPa Fig.5 Improvement of trade-off due to CNG rate

8

2

amount of 0.023 [kg/kg]. NOx was decreased markedly by about 80% at the low load because the combustion temperature was decreased markedly by a marked increase in the unburned natural gas having high specific heat which is about three times of air. On the other hand, NOx was not decreased at the high load because the pre-mixture of natural gas burned very fast and almost completely. With respect to fuel consumption, it increases significantly at the low load, which results from a large increase in the unburned hydrocarbon emission as shown by “THC” in the figures, on the other hand, it is decreased slightly at the high load. It is noticeable that, by increasing the pre-mixture temperature, the rate of increase in fuel consumption shown in the low load case becomes smaller and that of decrease in the high load case becomes larger, at the same time, the rate of increase in THC becomes small. This indicates that fuel consumption can be improved by the increase of intake pre-mixture temperature as suggested by Kusaka, et al.[1]. Figures 5(a) and (b) show changes in trade-off relationships between smoke and NOx, and between fuel consumption and NOx at the low and high loads respectively. The trade-off between smoke and NOx was improved significantly at both low and high loads by the CNG charge, on the other hand, the trade-off between fuel consumption and NOx was markedly improved at the high load, however, it is not at the low load because of the remarkable increase in the unburned hydrocarbon.

10

P [MPa]

8 6 80

4

60

2

40

Lift [mm]

0

20 0

0.3 0.2 0.1 0

X

P [MPa]

CNG

0 0 0 0

-20 -10

8

0

10

20

30

, X

EGR

, , , ,

40

0 0.07 0.11 0.17

50

CA [deg] (a) Without CNG (TIN=40°C)

6 80

4

60

2

40

Lift [mm]

0

EFFECT OF “EGR” RATE ON COMBUSTION

dQ/dθ [J/deg]

0

20

dQ/dθ [J/deg]

200

0

0.3 0.2 0.1 0

XCNG , 0 , 0.022 , 0.024 , 0.024 , 0.024 ,

-20 -10

0

10

20

30

40

XEGR

0 0 0.06 0.10 0.17

50

CA [deg] (b) With CNG (TIN =80°C)

Figures 6(a) and (b) and Figs.7(a) and (b) show changes in time histories of combustion due to EGR. Figs.6(a) and (b) are the cases without and with the CNG charge at the low load, and Figs.7(a) and (b) are those at the high load. Figure

Fig.6 Change in combustion histrory due to EGR rate (Pme=0.33 MPa, θinj=5°BTDC)

4

1.5

X

0.3 0.2 0.1 0

CNG

EGR

0

10

20

30

, , , ,

40

△

3.15

80

1.5

40

0.5

50

1 0.8 0

0.05

6 120

4 2

80

0

40

0.3 0.2 0.1 0

X

CNG

0.022 0.022 0.023 0.024 0.025

0

10

20

30

40

, X

, , , , ,

1

0 EGR

0 0.06 0.12 0.15 0.17

1.2 1 0.8 0.6 0.4 0.2 0

0.5

GCNG (kg/h)

TIN (°C)

○

0

80

△

3.20

80

0

1.2 be/be0

1 0.8 0

G CNG[kg/h] 0

3.15

0

4

0.05

0.1

0.1 0.15 XEGR[kg/kg]

0.2

Fig.9 Change in fuel consumption and exhaust emissions due to EGR rate (θinj=5°BTDC, TIN =40, 80°C)

80 80 80 40

0

0.05

(b) Pme=0.66 MPa

T IN[°C]

6

4 2

10 0.66 0.33

6

50

Fig.7 Change in combustion histrory due to EGR rate (Pme=0.66 MPa, θinj=5°BTDC)

Pme[MPa]

0.2

1.5

CA [deg] (b) With CNG (TIN=80°C)

8

0.1 0.15 XEGR[kg/kg]

(a) Pme=0.33 MPa

NOx /NOx 0

P [MPa]

2

1.2

CA [deg] (a) Without CNG (TIN =80°C)

-20 -10

Ignition Delay [deg.CA]

0

1

0 0.07 0.13 0.19

be/be0

0 0 0 0

8

2

TIN (°C)

GCNG (kg/h) ○

0 , 0 (TIN=40)

-20 -10

Lift [mm]

0

, X

dQ/dθ [J/deg]

Lift [mm]

20

0.5

Smoke /Smoke 0

40

0

1

THC/THC0

2

1.2 1 0.8 0.6 0.4 0.2 0

Smoke /Smoke0

60

NOx /NOx0

80

4

dQ/dθ [J/deg]

P [MPa]

6

THC/THC 0

8

both cases with and without CNG, especially in Fig.7(b), this phenomenon is quite clear in the case with CNG. The end of combustion delays as the ignition delay increases, thus, a slight increase in fuel consumption results. Figures 9(a) and (b) show the effect of the EGR rate on exhaust emissions and fuel consumption in the cases without and with the CNG charge. Each ordinate is normalized by the measured value obtained under the zero EGR condition. By increasing the EGR rate, NOx was obviously reduced and smoke was increased at both low and high loads. With respects to fuel consumption and the unburned hydrocarbon emission, their variations due to EGR are small at both low and high loads. It is noticed that the reduction rate in NOx due to EGR is different depending on the load, that is, the NOx reduction rate in the high load case is about twice of the low load case. With respect to the effect of EGR on NOx reduction, Kusaka, et al.[1] have mentioned that NOx is reduced by the lower combustion temperature due to the inert gas brought by EGR. The effect of the inert gas is, however, not quantitative

0.15 0.2 XEGR [kg/kg]

Fig.8 Change in ignition delay due to EGR rate

8 shows a change in ignition delay due to the EGR rate. In the cases of low load, the history of heat release rate is hardly changed by the EGR rate although it varied markedly by the CNG charge as shown in Fig.6(b). Under the high temperature of pre-mixture at the low load, ignition delay increases slightly as shown in Fig.8, and the burning rate of natural gas is suppressed slightly in the former half of the combustion duration, which might be caused by the suppression of flame propagation velocity due to the induced inert gas. On the other hand, in the cases of high load, ignition delay is increased more by EGR as shown in Fig.8, and, at the same time, the initial combustion heat release rate is suppressed by EGR in 5

0.8

1.5

XEGR = 0 XCNG = 0 TIN=40°C

F7 F7 F3 F3 F7

1

40 80 80 80 40

0 3.15 0 3.2 0

Smoke [Bosch]

NOx / NOx0

TIN[° C] GCNG[kg/h]

Port water injection

θ inj = TDC

0.6

θinj = 5°BTDC

0.4 0.2

θinj = 10°BTDC

0.5 0

EGR

350 0 0.01

0.02

0.03

be [g/kWh]

0

∆XW [kg/kg] Fig.10 NOx reduction rate due to water content in EGR

XCNG (kg/kg) XEGR (kg/kg) △ ◇ ▽ □

300

θinj = TDC EGR+CNG

200

0.66 0.33

1.1 ηvair/η v0

EGR

2

θ = 5°BTDC inj

θinj = 10°BTDC

4

6 8 NOx [g/kWh] (a) Pme=0.33 MPa

10

12

4

0

0.05

0.1

0.15

Smoke [Bosch]

0.9 0.8

0

ηv0

1

0 0.06 0.10 0.17

250

1.2 Pme[MPa]

0.022 0.024 0.024 0.027

0.2

X∗EGR + X∗CNG Fig.11 Change in volumetric efficiency due to CNG and EGR rates

explanation. In order to make it more clear, the NOx reduction rate due to EGR was analyzed from the viewpoint of water content included in the recirculated exhaust gas. Figure 10 shows change in NOx due to water content included in the EGR gas. The abscissa ∆x W[kg/kg] denotes the water content brought by EGR, which is the same definition of the absolute humidity of the air. As seen in the figure, the NOx reduction rate due to the water content is nearly equal between the data measured at both low and high loads. This is due to the fact that the water vapor concentration in the exhaust gas is about twice in the high load case compared with the low load case even in the cases with the equal EGR rate. For comparison, the NOx data measured by changing the amount of water injected at the port are shown by the open square marks in the figure. As shown in this figure, the equivalent absolute humidity of ∆xW=0.02 [kg/kg] results in 35% NOx reduction in the case of port water injection, which is almost the same NOx reduction rate shown by Ishida, et al.[4], on the other hand, it results in about 70% NOx reduction in the case of EGR, that is, the NOx reduction rate due to EGR is about twice of the one due to the port water injection. The second factor of the NOx reduction due to EGR seems to be the carbon dioxide content in the EGR gas, however, it is difficult to confirm that it is due to the thermal dissociation from carbon dioxide to carbon monoxide. According to the Zel'dovich mechanism on NO formation, a decrease in the oxygen content due to EGR might be considered as another factor, which has to be confirmed further.

3

θinj = TDC θinj = 5°BTDC

2

θinj = 10°BTDC

1 0 300

be [g/kWh]

XEGR = 0 XCNG = 0 TIN=40°C

XCNG (kg/kg) XEGR (kg/kg)

250

θinj = TDC

△ ◇ ▽ □

0.022 0.022 0.023 0.025

0 0.06 0.11 0.17

θinj = 5°BTDC θinj = 10°BTDC

200

150

0

2

4 6 NOx [g/kWh]

8

10

(b) Pme=0.66 MPa Fig.12 Improvement of trade-off in combination with CNG and EGR

Figure 11 shows the change in volumetric efficiency “ηV” due to EGR and/or CNG charge, which is normalized by ηV0 of the reference case without both EGR and CNG charge. The volumetric efficiency was defined only with respect to the intake fresh air excluding EGR and CNG, and the abscissa of (x*EGR+ x*CNG) denotes the mole fraction of EGR and CNG in the pre-mixture charged into the cylinder. The data show somewhat scattering but locate near the broken line which is calculated assuming that the total amount of intake charge is unchanged by introducing CNG and EGR. Figures 12(a) and (b) show changes in trade-off relationships between smoke and NOx, and between fuel 6

ηV θinj

consumption and NOx due to CNG plus EGR at the low and high loads respectively. In the low load case, the significant improvement in trade-off between smoke and NOx is obtained mainly by the CNG charging, however, the trade-off between fuel consumption and NOx is deteriorated, in addition, the effect of EGR on both trade-offs is small in comparison with that of CNG. On the other hand, the trade-off improvement is obtained mainly by EGR at the high load, especially in the trade-off between fuel consumption and NOx. According to the present experiments, it seems that a better trade-off between smoke and NOx at the high load can be obtained by increasing the CNG charging rate in combination with the appropriate EGR rate, and a better trade-off between fuel consumption and NOx at the low load may be obtained by the better combination between the preheating and the EGR rate as suggested by Kusaka, et al.[1].

Volumetric efficiency Injection timing of gas oil [° Crank angle]

ACKNOWLEDGMENT The authors wish to thank to Mr. Nitta, Y. of ISUZU Motors Ltd., and Mr. Fujita, H. of ZEXEL Corp. for their supports on experimental apparatus, and Mr. Kudo, K. and Mr. Matsuoka, T. of the graduate school of Nagasaki University and the undergraduate students in the Energy System Laboratory, Nagasaki University for securing the experiment.

REFERENCES [1] Kusaka, J., Daisho, Y., Kihara, R., Saito, T. and Nakayama, S., "Combustion and Exhaust Gas Emissions Characteristics of a Diesel Engine Dual-Fueled with Natural Gas", Proc. of the 4th International Symposium COMODIA 98, pp.555-560 (1998) [2] Miyamoto, N., Ogawa, H., Nurun, N. M., Obata, K. and Arima, T., "Smokeless, Low NOx, High Thermal Efficiency and Low Noise Diesel Combustion with Oxygenated Agents as Main Fuel", Trans. SAE, J. Fuels & Lubricants, Sec.4, Paper No.980506, pp.171-177 (1998) [3] Ishida, M., Ueki, H., Sakaguchi, D. and Imaji, H., "Simultaneous Reduction of NOx and Smoke by Port Injection of Methanol/Water Blend in a DI Diesel Engine", Proc. of 15th Internal Combustion Engine Symposium (International), Paper No.9935202, pp.93-98 (1999) [4] Ishida, M., Ueki, H., Sakaguchi, D. and Izumi, S., "Significant NOx Reduction in Diesel Engine Based on Electronically Controlled Port Water Injection", Proc. of the 22nd CIMAC International Congress on Combustion Engines, Paper No.07.09, Vol.4, pp.879-893 (1998)

CONCLUSION In order to improve the trade-off between smoke and NOx without deteriorating fuel consumption in a DI diesel engine, the effect of pre-mixed natural gas on combustion and exhaust emissions has been investigated experimentally. The effects of the CNG charging rate and the EGR rate on ignition and combustion rate are mainly examined, in addition, the effects of CNG and EGR on NOx reduction was analyzed from the viewpoint of water content in the EGR gas. Concluding remarks are as follows. (1) The burning rate of natural gas pre-mixture is larger under the higher temperature condition, thus, the larger burning rate results in shortening the combustion duration and leads to lower fuel consumption. (2) Ignition and the burning rate of natural gas premixture are suppressed by EGR, which might be caused by suppression of flame propagation due to the inert gas induced in the pre-mixture. (3) Significant NOx reduction is obtained by charging natural gas under the low load condition because of the large amount of unburned natural gas having high specific heat. (4) The NOx reduction rate due to EGR is about twice of the one due to the port water injection. The first factor of NOx reduction due to EGR is the water content brought by EGR, and the second factor might be a decrease in the oxygen content due to EGR.

NOMENCLATURE be Brake specific fuel consumption [g/kWh] Lift Needle valve lift [mm] NOx Brake specific nitrogen oxides emission [g/kWh] P Cylinder pressure [MPa] Pme Brake mean effective pressure [MPa] dQ/dθ Heat release rate [J/deg] Smoke Smoke density [Bosch] TIN Intake pre-mixture temperature [°C] THC Total unburned hydrocarbon emission [g/kWh] xCNG CNG charging rate (=GCNG/GAIR) [kg/kg] xEGR EGR rate (=GEGR /GINTAKE-CHARGE) [kg/kg] xW Equivalent absolute humidity [kg/kg] 7