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Combustion improvement for reducing exhaust emissions in IDI diesel engine
ELSEVIER
JSAE Review 18 (1997) 19-31
Combustion improvement for reducing exhaust emissions in IDI diesel engine Yoshihiro Hotta a, Kiyomi Nakakita a Minaji Inayoshi a, Takashi Ogawa b, Takeshi Sato b, Mitsumasa Yamada b a Mechanical Engineering Div. 1, Toyota Central Res. & Develop. Labs., Inc., 41-1 Yokomichi, Nagakute, Nagakute-cho, Aichi-gun, Aichi, 480-11 Japan b Engine Engineering Div. Ill, Toyota Motor Corp., 1200 Mishuku, Susono-shi, Shizuoka, 410-11 Japan
Received 10 May 1996
Abstract Means for reducing PM from a swirl chamber type diesel engine were examined and the mechanisms of the PM reduction were investigated using both a multi-cylinder and an optically accessible single-cylinder engine. The following points were clarified. (1) At light load, suppression of initial injection rate reduces PM, because of both the change of ignition point to reduce SOF, and the retarded flowout of dense soot from the swirl chamber to reduce smoke. (2) Over medium load, the main cause of the exhaust smoke is hard spray-wall impingement which leads to both fuel adhesion on the wall and the stagnation of rich fuel-air mixture.
1. Introduction
As swirl chamber type diesel engines are widely used for both passenger cars and commercial cars of up to medium-sized class, the reduction of their exhaust emissions is an urgent problem to be tackled. Therefore at present, various researches have been progressing [1-3] from the aspect of both combustion improvement and aftertreatment. This study aims at the simultaneous reduction of NOx and particulate matter (PM) by improving the combustion itself. First, some means for simultaneous reduction of NO x and PM were obtained by investigating the effects of injection characteristics such as injection rate and injection pressure and of swirl chamber volume ratio, using a multi-cylinder engine. Then, the mechanism of emission improvement was examined in detail using an optically accessible, single-cylinder engine.
(1) NO x is reduced by decreasing the pre-mixed burning rate caused by the suppression of initial injection rate. (2) Smoke is reduced by promoting both fuel atomization and air entrainment into the fuel spray caused by the increase of injection pressure [4]. (3) Smoke is reduced by decreasing equivalence ratio in the swirl chamber caused by the increase of swirl chamber volume ratio. The smoke reduction in (2) and (3) is expected to lead not only to the PM reduction, but also to the NO x reduction caused by the possible increase of EGR rate. Whether these hypothetical mechanisms can be actually realised is examined through multi-cylinder engine experiments in Section 4.
I Low Initial
}
Injection Rate I
= Reduction of L ~ Premix Burning I ~ EGR
2. Concept for simultaneous reduction of NO x and PM
A concept for simultaneous reduction of NO x and PM was hypothesized as shown in Fig. 1. This consists of the following NOx/PM reduction mechanism.
High Injection ~ Pressure I Swirl Chamber
Volume Ratio
Fuel Spray L Increase/ I Atomization, etc I ~ ] ' ~ " ~
NOx ...... I /I ........... I
~" Equivalence Ratio [
in Swirl Chamber I
I Reduction J
"
-
Fig. 1. Concept for NO~ and PM reduction.
0389-4304/97/$17.00 Copyright © 1997 Society of Automotive Engineers of Japan, Inc. and Elsevier Science B.V. All rights reserved Pll S 0 3 8 9 - 4 3 0 4 ( 9 6 ) 0 0 0 5 0 - 1 JSAE9730038
20
Y. Hotta et al./JSAE Review 18 (1997) 19-31 1450 rpm 20 %load
Table 1 Specifications
Bore X stroke (ram) Number of cylinders Displacement (cc) Air aspiration system Compression ratio Swirl chamber volume ratio (%) Injection pump Injection nozzle type Nozzle-opening pressure (MPa) Prelift (mm)
2000 rpm 50 %/load
of multi-cylinder e n g i n e 96 × 103 4 2982 Turbocharged 21.2 58, 62
-. a~~
5
2
_~
E
om
~
= ' 0 ~-20 Distributor type Throttling pintle nozzle 1 Spring 2 Spring t5 (base) -
15/20 20/25 0.2
0 15/25 25/30
3. Experimental apparatus and conditions using multicylinder engine The specification of the multi-cylinder engine and the injection system used in this study are listed in Table I, and the schematic diagram of the swirl chamber configuration is shown in Fig. 2. An ordinary single-spring injector and also several double-spring injectors were used, in order to change both the injection rate and the injection pressure. Engine running conditions were selected as 20% load at 1450 rpm as the light load condition and 50% load at 2000 rpm as the medium load condition, which are representative of conditions in actual use. Injection timing was adjusted so as to optimize the NOx-PM trade-off characteristics for each case.
4. Results and discussion on multi-cylinder engine experiment
~£ " Fig. 2. Schematic diagram of swirl chamber (multi-cylinder engine).
0 10 20 30 Crank Angle deg
Fig, 3. Comparison of initial injection characteristics,
increased up to 20 or 25 MPa, while the first opening pressure was maintained at the same pressure as that of the single-spring injector (hereafter Po 15 / 20 MPa, Po 15/25 MPa respectively). The duration of suppressed initial injection rate increases in the order of Po15 MPa, P o 1 5 / 2 0 MPa and Po15/25 MPa, due to the function of the double-spring nozzle holder. Figure 4 shows the effects of initial injection rate suppression on exhaust emissions under both light load and medium load. Against expectation, but not NO X, both smoke and PM are reduced with the injection rate suppression. Under light load, both smoke and PM decrease remarkably depending on the duration of the initial-injection-rate suppression. This cause is examined in detail in Section 6. Under medium load, the tendency is slightly different. In the suppressed initial-injection-rate case of Po15/20 MPa, both smoke and PM are clearly reduced compared with the base case of Po15 MPa. However, in the further suppressed initial-injection-rate case of Po15/25 MPa, smoke and PM are hardly reduced. One reason for this is consid-
4.1. Effects of suppressing initial injection rate In this section, the hypothetical mechanism (1) in Section 2 is evaluated. The evaluated injection characteristics are shown in Fig. 3. The single-spring injector with nozzle opening pressure of 15 MPa (hereafter designated as Po15 MPa) was selected as the base injection system. For the double-spring injectors, the second opening pressure was
I 0 10 20 30 Crank Angle deg
E / o:Pol5MPa o.. 200 I- zx : 15•20 o_ | o: 15/25 0"r'
O3
2000 rpm 50 %load
1450 rpm 20 %load
300 j
-
I
~ I
I
20
30
0
0
10
EGR %
40
0
2
4
6
8
EGR %
Fig. 4. Effects of initial injection rate suppression on e x h a u s t e m i s s i o n s .
Y. Hotta et al./ JSAE Review 18 (1997) 19-31
21
1450 rpm 20 %load 300 r / o : Po15/20MPa & 200 ~- zx: 20/25 o_ / n: 25/30
E
_
Cq
o 100~ . . _ . . . , ~ , , ~ 0 i
t
t
2000 rpm 50 %load
t
/
O I---
1400 rpm
2000 rpm
Fig. 5. Effectsof initial injection rate on full load torque.
oE ered to be that the injection duration of Po15/25 MPa under medium load becomes too long to fully oxidize the fuel injected at the latter part of the injection period. Figure 5 shows the maximum torque in each case under the limited smoke level of 1.5 BSU. The maximum torque decreases with suppression of the initial injection rate. This is caused by the fact that the injection amount must be reduced when suppressing the initial injection rate to keep the limited smoke level, because the excess elongation of suppressed-injection-rate duration increases the smoke due to the excessively long injection duration, especially under high load. In the case of a double-spring injector, injection pressure at the latter part of the injection period becomes higher due to both the high second nozzle opening pressure and the use of the higher injection rate range on the pump cam. Therefore, the decrease of smoke and PM mentioned above are obtained by the effects of both the initial injection rate suppression and the injection pressure increase. This effect of injection pressure is clarified in the next section. 4.2. Effects o f increasing injection pressure
In this section, the hypothetical mechanism (2) is evaluated. P o 1 5 / 2 0 MPa was selected as the base injector, because it has both the remarkable effect of PM reduction under light and medium loads and the least increase of smoke under high load. In order to evaluate the effects of injection pressure increase, exhaust emissions of Po20/25 MPa and P o 2 5 / 3 0 MPa were compared to that of the base.
1450 rpm 20 %load
2000 rpm 50 %/load
0
10
20
30
EGR %
40
0
2
4
6
Fig. 7. Effectsof injection pressureon exhaust emissions.
The difference between the first and the second nozzle opening pressures of these injectors were kept constant so as to prevent the effects of initial injection rate suppression from changing. As shown in Fig. 6, the injection pressure increases throughout the injection duration while the suppressed-injection-rate duration is maintained as the same as that of the base. Exhaust emissions under both light and medium loads are shown in Fig. 7. Against expectation, under medium load, both smoke and PM increase remarkably as the injection pressure increases. Furthermore, even under light load, smoke and PM reductions are hardly achieved. This smoke increase tendency is more remarkable under heavy load. This causes a remarkable decrease of the maximum torque under the limited smoke level of 1.5 BSU as shown in Fig. 8. This mechanism of smoke and PM increase is examined in detail in Section 7. Based on the results of this section, the following points are revealed. (1) Smoke and PM reductions by suppressing initial injection rate shown in the previous section are caused by overcoming the increase of smoke and PM by the injection pressure increase at the latter part of the injection period. (2) Smoke and PM increases under the excess suppression of initial injection rate at full load condition (Fig. 5)
~rr~E "e E ,- ~ 4
0
[
E
~
Zr
O
o o
~'~"
I/ 01
~ P o 15/20 MPa t
i
t
x
t
0 10 20 30 CrankAngle deg
0 10 20 30 Crank Angle deg
Fig. 6. Comparisonof injection pressurecharacteristics.
8
EGR %
O F-
1400 rpm
2000 rpm
Fig. 8. Effects of injection pressureon full load torque.
22
Y. Hotta et al./JSAE Review 18 (1997) 19-31 1450 rpm 20 %load 300 I Po 15/20 MPa
2000 rpm 50 %load
|
200 I- o: a58% /
0~
z~: 62
f
~'/
i
i
i
Smoke1.5BSU Po15/20MPa #' 58% 62
E ¢~
[--
g
m.......
I---
1400 rpm i A
v
Fig. 10. Effects of swirl chamber volume ratio ( / 3 ) on full load torque.
I
0
10
20 EGR %
2000 rpm
30
40
0
4 6 EGR %
10
Fig. 9. Effects of swirl chamber volume ratio ( f l ) on exhaust emissions.
are considered to be caused not only by the excessively elongated injection period, but also by the increased injection pressure at the latter part of the injection period as mentioned in Section 4.1.
4.3. Effects of increasing swirl chamber volume ratio In this section, the hypothetical mechanism (3) is evaluated. Swirl chamber volume ratio (hereafter designated as /3) was increased up to 62% from 58% of the base condition while keeping the same compression ratio. Concerning the injector, Po15/20 MPa was selected considering the above results. Figure 9 shows the comparison of exhaust emissions between /3 = 58% and /3 = 62% under both light and medium loads. It is clear that the reductions of smoke and PM by increasing /3 at medium load are much larger than that at light load. Furthermore, these reductions become larger with increasing EGR rate. These results indicate that the effect of increasing /3 on smoke and PM reductions becomes greater as the equivalence ratio increases. Concerning the PM reduction in this case, it is considered to be caused not only by soot reduction but also by SOF reduction because HC is also reduced. NO x increases with increasing /3 in the range of low EGR rate at medium load. However, simultaneous reduction of NOx and PM is consequently obtained because this NO~ increase can be cancelled out by increasing EGR rate as follows. When the emissions at several EGR rate conditions of /3 = 62% are compared with the emission at 2.5% EGR rate of /3 = 58% regarded as the base condition, then (1) at the EGR rate of 2.5%, smoke and PM are remarkably reduced, while NO x is higher. (2) with increasing EGR rate up to 4%, smoke and PM are reduced without increasing NO x.
(3) during the range of EGR rate from 4% to 8%, simultaneous reduction of smoke, PM and NO x is obtained. Additionally as shown in Fig. 10, maximum torque under the limited smoke level of 1.5 BSU increases by increasing /3 due to the smoke decrease. Therefore, increasing /3 enables the increase of maximum torque in spite of using the double-spring injector which decreases the maximum torque as shown in Section 4.1 (Fig. 10, /362%-Po15/20 MPa vs. /358%-Po15 MPa).
4.4. Simultaneous reduction of NO, and PM The results in Sections 4.1 to 4.3 show that simultaneous reduction of NO x and PM by combining the suppression of initial injection rate with the increase of swirl chamber volume ratio was confirmed, as shown in Fig. 11. In this measurement, the EGR rate was adjusted so as to keep the exhaust smoke level of 0.5 BSU. A is the base case, and B is the case of suppressing initial injection rate. C is the case of combining the increase of swirl chamber volume ratio with B. Comparing C with A, it is clear that the reductions are approximately 40% of NO x and 20% of PM at light load, 40% of NO, and 35% of PM at medium load, respectively.
35 -1450 - rpm - 20 - %load
~om3Ot rsmoke0.SBSU , ~ ~
10 2000 rpm 50 %load
] ~ ~
5f-~° k ~
~ 1.0 •~ 0.5
{,op osF
1I
A B C A B C A: Base B: Suppression of Initial Injection Rate (Po15/20 MPa) C: B Jr Increase of Swirl Chamber Volume Ratio (/362 %) *)SOF : Soluble Organic Fraction *)ISOF: In-Soluble Organic Fraction
Fig. 11. NO x and PM reduction with combination of suppressing initial injection rate and increasing swirl chamber volume ratio.
Y. Hotta et al./JSAE Review 18 (1997) 19-31
Mirror
Table 2 Specifications of single-cylinderengine Bore× stroke Displacement Compression ratio Swirl chambervolumeratio
23
92 mm× 92 mm 612 cc 20.8 52%
Half Mirror
u It is also clear that the cause of NO x reduction is the possible increase of EGR rate due to the decrease of PM. Therefore in the following sections, the mechanisms of these PM reductions are examined in detail mainly by in-cylinder visualization analysis.
5. Apparatus and procedure for in-cylinder observation
Convex,ens
V High Speed Camera
Fig. 13. Schematic diagram of shadow-graph photography.
lyzed based on the in-cylinder pressure which was recorded by a mini-computer every 0.5 ° CA throughout 200 cycles.
5.1. Optically accessible single-cylinder engine 5.2. Apparatus for visualization The specification of the optically accessible engine is listed in Table 2, and the schematic diagram of the swirl chamber configuration is shown in Fig. 12. The swirl chamber is constructed as a removable unit in order to evaluate several chamber designs, and therefore the connecting passage is longer than that of the multi-cylinder engine used in the previous section. Also, this swirl chamber is designed to be two-dimensional, with flat side walls for visualization by the shadow-graph method. A piezo type pressure transducer and a glow-plug shaped thermocouple are attached to measure gas pressure and representative atmospheric temperature in the swirl chamber respectively. In the visualization experiments, the engine with an elongated piston was kept motoring with control of both coolant and lubricant at 80°C and with control of intake air at 55°C. Just when the representative temperature of the swirl chamber reached 410°C, fuel injection was started. The exhaust gases and the combustion characteristics were also examined under steady-state running condition with an ordinary short piston instead of the elongated piston. Combustion characteristics were ana-
Shadow-graph photographs were taken with the apparatus shown in Fig. 13, at a frame speed of 9000 fps using a high-speed camera. A tungsten halogen lamp was used for lighting. The observation field is shown in Fig. 14. Both the whole swirl chamber area and the center area of the main chamber whose diameter is 75% of the cylinder bore were simultaneously observed. 5.3. EGR method In the visualization experiments, if real exhaust gas was recirculated just like ordinary EGR, the observation windows would be dirtied by the formation of dense soot before reaching the required EGR rate, so that clear observation would be impossible. Hence, to simulate the required EGR condition in this study, the exhaust gas under preliminary steady engine running was transferred into a 900-liter sampling bag after removing soot, oil-mist and water. Then, the stored clean exhaust gas was mixed into the intake air in the visualization or measurement of emissions.
Main Chamber
Swirl Chamber
. . - ~ 7---'~Val, res
~x
~
iI I
I //
Throat
~Crank Angle
-,2-'_2-"
Throat Location Fig. 12. Schematicdiagramof swirlchamber(single-cylinderengine).
Fig. 14. Observationfield.
24
Y. Hotta et a l . / J S A E Review 18 (1997) 19-31
Table 3 Experimental conditions for analysis of the effect of initial injection rate suppression (light load) Engine speed Volumetric efficiency Fuel amount Equivalence ratio Injection timing EGR rate Nozzle-opening pressure
1600 rpm 100% 20 mm3/st 0.34 - 4° ATDC 0-24% 1 Spring 15 MPa
3
2 Spring 20/30MPa
3 t/)
=
6. P M reduction m e c h a n i s m o f initial injection rate suppression
2
,
,
.
|
|
|
Conventional Inj. Rate ,.,.,~
o
E
1
Low Initial Inj. Rate |
6.1. Exhaust emissions and combustion characteristics
o
1'o EGR
The smoke and PM reduction mechanisms by the initial injection rate suppression were investigated under the light load conditions listed in Table 3, where the largest reduction of PM was obtained in Section 4.1. A single-spring injector with nozzle opening pressure of 15 MPa was used for the representative of conventional injection rate (hereafter Conventional), and a double-spring injector with nozzle opening pressures of 2 0 / 3 0 MPa was used for the representative of the suppressed initial injection rate (hereafter Suppressed). A comparison of the injection pattern between Conventional and Suppressed is shown in Fig. 15. The characteristics of Suppressed are as follows. (1) The injection rate at the first half of the injection period is suppressed quite low and kept nearly constant, and its peak value at the latter part is only 54% of the Conventional. (2) The injection period is extended to 24 ° CA from 14° CA of the Conventional. (3) The peak injection pressure increases to 40 MPa from 22 MPa of the Conventional. Comparisons of exhaust emissions between Conventional and Suppressed are shown in Fig. 16. --
~..~ .g
Conventional Inj. Rate Low Initial Inj.Rate
2
10[ 0.5
z
0
.-
-4°f
E "" ~" 2 0
o
Fig. 15. Comparison of injection characteristics at light load.
30 %
Fig. 16. Effects of initial injection rate suppression on exhaust smoke.
It was confirmed that the initial injection rate suppression reduces smoke and HC but not NO x, the same results as for the experiments with multi-cylinder engine in Section 4.1. It is considered that reduction of NO x was not obtained due to the increased injection pressure, which canceled the NO x reduction effect caused by the suppressed initial injection rate and the extended injection period. Additional characteristics of the emissions are as follows. (1) In the case of Conventional, smoke level is fairly high and increases rapidly with EGR rate. In the case of Suppressed, smoke level is remarkably low and hardly increases with increasing EGR rate. (2) HC level of Suppressed is approximately half of that of Conventional. Comparisons of the heat release pattern and the combustion characteristics are shown at the top of Fig. 18, and in Fig. 17, respectively. Figure 18 shows that the heat release of Suppressed is lower at the first stage and higher at the second. On the other hand, Fig. 17 shows that the initial injection rate suppression increases the ignition delay period only by 2 degrees, and extends the combustion period only by 1 degree. These observations indicate that the initial injection rate suppression retards main combustion time while keeping almost the same ignition delay and combustion period. Also injection timing retard causes the retard of main combustion time, but it increases fuel consumption due to the excessively elongated ignition delay and combustion period, even though having an advantage of smoke reduction. The initial injection rate suppression does not cause such undesirable combustion and therefore the fuel consumption is maintained at the same level as that of Conventional as shown in Fig. 17. This characteristic is con-
25
Y. Hotta et al.//JSAE Reoiew 18 (1997) 19-31 Zx Conventional Inj. Rate O Low Initial Inj. Rate
~..= 180 -~.== 160 I
I'~
~ 8
•
,
I
|
•
8o
6.2. PM reduction mechanism
i
i
i
|
m
•
|
•
i
!
=
=
Mixture formation and flame development processes in the swirl chamber and in the main chamber are shown in Fig. 18 and Fig. 19 respectively. The differences between Conventional and Suppressed, and PM reduction mechanisms are as follows.
|
t
31 29
sidered to be one of the main advantages of the initial injection rate suppression.
I
33}
E P~ 120 •~ 0. z=
,
~
o
-O ~
i
i
i
0
10
20
(1) SOF (Soluble Organic Fraction) reduction mechanism. In the case of Conventional, a lot of fuel vapor is formed even from the early stage and the ignition occurs near the injection nozzle (at 1.3 ° ATDC of Conventional, in Fig. 18), so that unburned mixture exists ahead of the flame. This unburned mixture flows out to the main chamber prior to the flame. In the main chamber, there exist some areas where flames do not spread throughout the
|
t
30
EGR %
Fig. 17. Effects of initial injection rate suppression on combustion characteristics.
Conventional Inj. Rate [Conventional Inj. Rat [ EGR 0% EGR 24"~ I
/Jk\t
-40
0
40
=
80 ~
Crank Angle "ATDC
i
"I"
"
-40
0
5.0"
-40
15.5"
~ 9 . 8 "
0
40
/ i
5.5" Outflow of Flame
~ 1 o . ~
15.8" ~
80 :~
Crank Angle "ATDC
: .
Outflow of Dense Soot
9.5"
°
40"-~'-"'~0
Crank Angle "ATD~
m
4.7"
Low Initial Inj.Rate-] EGR 240/,
16.3"
I
"Outflow o f
[]
Dense Soot
Fig. 18. Heat release rate and combustion process in swirl chamber.
26
Y. Hotta et al./JSAE Review 18 (1997) 19-31
Conventional Inj. Rate EGR 24% Luminou., flame 8.5*
L o w Initial Inj. Rate EGR 24%
ATDC Outflow of Dense sool
11.5"
13.8"
16.9"
22.0*
Fig. 19. Comparisonof flame developmentin main chamber(EGR 24%).
combustion period under this light load condition (at 16.9 ° ATDC of Conventional, in Fig. 19). Hence, unburned mixtures that spread into these areas are considered to be not oxidized completely and exhausted. In the case of Suppressed, the amount of formed fuel vapor is fairly small at the early stage and the ignition occurs near the spray tip (at 1.8 ° ATDC of Suppressed, in Fig. 18), so that there are no unbumed mixtures ahead of the flame. This is considered to be the cause of HC and SOF reduction. The cause of shifting the ignition point is considered to be as follows. In the case of Conventional, injection rate increases sharply at the beginning of the injection period as shown in Fig. 4, so that the momentum of the fuel spray also increases sharply. Therefore, the fuel injected at the beginning is overtaken by the fuel injected later, and carried near the vicinity of the nozzle. The ignition is considered to occur at the mixture formed by the fuel that is injected at the beginning and stayed for a long time in a high-temperature atmosphere, so the ignition occurs at the vicinity of the nozzle. In the case of Suppressed, it is considered that the overtaking does not occur on a large scale, because the injection rate is suppressed to quite low and is nearly constant during the suppression period, so
that the fuels are sequentially carried by the strong swirl. Therefore, the fuel which is injected at the beginning remains near the spray tip to cause ignition.
(2) Soot (Insoluble Organic Fraction) reduction mechanism. In the case of Conventional, fuel spray impinges against the chamber wall opposite t o t h e injection nozzle from the beginning due to the high injection rate. Therefore in Fig. 18, a dense soot cloud shown by the dark area, which is peculiar to the case of high exhaust smoke and is considered to be generated by the cooling of the dense soot on the quartz window [5], appears near the spray-wall impinging point from the early stage of the combustion period, and some part of this dense soot flows out to the main chamber just when the initial flame flows out (at 5.0 ° ATDC of Conventional, in Fig. 18). This is confirmed by the fact that the dense soot is observed in the main chamber at almost the same time of the flowout of the flame (at 8.5 ° ATDC of Conventional, in Fig. 19). As the formation of the dense soot continues throughout the injection period due to the continuous hard fuel-spray impinging, the dense soot continuously flows out to the main chamber and can not be fully oxidized. This incomplete oxidation is shown by the fact that the dense soot does not disappear and spreads to the whole area of the main chamber (at 16.9 ° ATDC of Conventional, in Fig. 19). In the case of Suppressed, the flame develops along the swirl chamber wall without the spray-wall impinging during the suppression period due to the quite low injection rate (at 3.0 ° ATDC of Suppressed, in Fig. 18). Therefore, the flame with a sparse rather than a dense soot cloud flows out into the main chamber, unlike in the case of Conventional (at 5.5 ° ATDC of Suppressed, in Fig. 18). This luminous flame changes into either burned gas or nonluminous flame by sufficient oxygen as soon as it flows out into the main chamber. This burned gas or nonluminous flame is shown as the shadow image caused by the density gradients near the throat (at 8.3 °, 1 1.7 °, and 14.0 ° ATDC of Suppressed, in Fig. 19). The dense soot cloud is formulated also in this case due to the hard spray-wall impinging during the main injection period which corresponds to the second stage of nozzle needle lift (at 16.3 ° ATDC of Suppressed, in Fig. 18). However, the quantity of the dense soot is less than that in Conventional, because the main injection period is shorter under this light load condition. Additionally, when this dense soot cloud flows out to the main chamber (at 17.2 ° ATDC of Suppressed, in Fig. 19), the temperature in the main chamber has already been raised by the burned gas and the flame which flowed out prior to the dense soot. In other words, the potential for oxidizing soot in the main chamber has already been raised to the level enough to oxidize the dense soot cloud easily. Additionally, the cause that smoke rapidly increases with increasing EGR rate in the case of Conventional but
27
Y. Hona et aL / J S A E Review 18 (1997) 19-31
Table 4 Experimental conditions for analysis of the effects of injection pressure and swirl chamber volume ratio (medium load) Engine speed Volumetric efficiency Fuel amount Equivalence ratio Injection timing Nozzle-opening pressure
1800 rpm 100% 30 m m 3 / s t 0.52 - 8° ATDC 2 Spring 15/20 MPa
~ero~
';o-:E'~'0.6 -w I/I ' ~ ~0.4 ~'~ 0.21 , "~
2 Spring 25/30MPa
7. S m o k e formation m e c h a n i s m under m e d i u m load condition
7.1. Main cause of smoke increase with increased injection pressure
The mechanism of the smoke increase with increased injection pressure was examined under medium load conditions, where the smoke remarkably increases as shown in Section 4.2. The experimental conditions are listed in Table 4. The nozzle opening pressures of the double-spring
'
t
• 15/20MPaI
.
~
hardly increases in the case of Suppressed, is considered to be as follows. It is clear by comparing the left and middle lines in Fig. 18 that EGR delays flame development. Accordingly, in the case of Conventional, at the beginning of the combustion period under EGR condition, the quantity of the flame is relatively too small to fully oxidize the dense soot cloud, so that the exhaust smoke increases. On the other hand, in the case of Suppressed, flame development delays under EGR conditions quite similarly to the case of Conventional. However, flowout of the dense soot cloud begins sufficiently later than that of the flame, so that the slight delay of the flame development does not have much influence on the soot oxidization process. Therefore, exhaust smoke is kept to a low level.
'
46 4
'
'
'
0
10
20
30
EGR % Fig. 21. Effect of nozzle opening pressure on combustion characteristics.
injectors are 15/20 MPa and 2 5 / 3 0 MPa. The comparison of the exhaust emissions are confirmed as shown in Fig. 20. It is clear that the increased injection pressure remarkably increases both smoke and HC just as in the case of the multi-cylinder engine in Section 4.2. Figure 21 shows the comparison of the combustion characteristics. It is clear that the injection pressure has little influence on the combustion characteristics such as ignition delay and the combustion period, except the maximum rate of pressure rise which increases approximately 35% with the increase of injection pressure. Little difference is recognized also in the heat release rate patterns in Fig. 22. In conclusion, it is revealed that the macroscopic combustion characteristics are hardly changed by the injection pressure increase, in spite of the remarkable smoke increase. Therefore, the combustion process was examined in detail by in-cylinder observation. The results are shown in Fig. 23. The photographs in the left and center line of Fig. 23 correspond to the case of P o 1 5 / 2 0 MPa and Po25/30 MPa, respectively. From these results, the following points are observed. During the initial-injection-rate suppression period (before 5 ° ATDC), there are no differences in both cases.
_~0.1.
X ol
I
|
|
I"
3
i:t
•
=~"~
°I 3
I
o.4t |
i
I
Z~,.,.,. ,.Z~ 25/30MPa
I
/
a: O.Ot
E~~MPa
0
.
o,8[
i
I
®. 90[ i~60
i
/
Fig. 20. Effect of nozzle opening pressure on exhaust emissions.
--
25130MPa
-- 15/20MPa
30
t
0 EGR 10 %20 30
/
"r
0
°
-40 0 Crank Angle
]
40 80 ° ATDC
Fig. 22. Heat release pattern for each nozzle opening pressure.
28
Y. Hotta et al./JSAE Review 18 (1997) 19-31
Namely, the ignition occurs near the spray tip, so that the flame successfully envelops the whole mixture area. Also, the fuel spray does not impinge on the chamber wall. Therefore, the dense soot cloud is not produced in both cases (2.5 ° ATDC). Thus, the required conditions for keeping the low PM level at light load shown in Section 6 are maintained in both cases. However, remarkable differences appear during the main injection period corresponding to the second needle lift (5°-17 ° ATDC). In this period, the fuel spray impinges on the chamber wall due to the increased injection rate, so that the dense soot cloud represented by the dark area is formed near the spray-wall impinging point in both cases (9.8 °, 9.7 ° ATDC). The difference is in the quantity of this dense soot cloud. In the case of Po25/30 MPa, the quantity of the dense soot is so large that the dense soot clouds are carried along the chamber wall (14.5 ° ATDC), and spread in the whole swirl chamber (32.4 ° ATDC). Thus, a large quantity of soot particles are observed in the latter part of the combustion period (57.7 ° ATDC). On the other hand in the case of P o 1 5 / 2 0 MPa, the quantity of the dense soot formed near the spray-wall impinging point is not so large as that in P o 2 5 / 3 0 MPa, so that far less soot clouds are carried along the chamber wall and spread in the chamber (17 ° to 32.6 ° ATDC). Thus, the soot clouds are easily oxidized to disappear (58.0 ° ATDC).
15,
2.4 °
ATDC
spray-wall ImpingementI • Less
spray-wailImpingement
I • Furious
spray-wallImpingement • Far less
..... .d~n..s.a.-soot .I.or..m.a.t.i.o. n..I ..... .d.a.n~?: .sp.ot1o.rma!l.o.n.......... .d.a.n~?: _s_ _o9_t_l_qr_.m_~]1_on__ 17.0"
19.8"
Fig. 23. Effects of nozzle opening pressure and spray impinging plate on combustion processes ( E G R 11%).
12
•
•
•
•
E
o80t
t.O
I
~:~ 40 0
•
-Oot 80
t
•
. . . .
-20 -10
0
10
20
30
Crank Angle •ATDC Fig. 24. C o m p a r i s o n of spray m o m e n t u m between high and low nozzle opening pressure.
In conclusion, the mechanism of this dense soot formation near the spray-wall impinging point is considered to be as follows. The hard spray impingement causes fuel adhesion on the wall near this point. This adherent fuel is not quickly evaporated and formed fuel vapor is hardly carried out of this area, because the stagnation zone is formed here due to the chamber shape. Thus, the rich fuel-air mixture stagnates in this zone under the condition of high temperature and insufficient oxygen to form the dense soot cloud. Though the hard spray-wall impingement is considered to be a main cause of the dense soot formation, the difference in formed soot amount between both cases is quite large. Therefore, the momentum of the fuel spray in each case was numerically analyzed by the method of characteristics [6] based on the experimental value of the nozzle needle lift and the pressure at the entrance of the injector. The results are shown in Fig. 24 as well as the pressure at the nozzle sac and the injection rate. It is clear that the momentum of the fuel spray at the main injection period of Po25/30 MPa is approximately twice as large as that of Po15/20 MPa. This is caused not only by the increase of injection pressure, but also by the increase of the injected fuel mass due to the increased effective flow area with increased nozzle needle lift. Thus, increased nozzle opening pressure causes drastic increase of spray momentum due to the characteristics of the throttle nozzle, so that the hard spray-wall impingement is caused to form a great deal of dense soot cloud. For really promoting fuel atomization and air entrainment while keeping the prevention of spray-wall impingement, it is considered that, for instance, a small-hole-diameter nozzle should be combined with this increased injection pressure just as in the case of the direct injection diesel engine [7].
29
Y. Hotta et al./JSAE Review 18 (1997) 19-31
7.2. Effect of spray impinging plate
WRh
Plate
A
In order to confirm the smoke formation mechanism mentioned above, the effect of a spray impinging plate shown in Fig. 25 was examined. This plate is intended to prevent the injected fuel droplets from adhering on the wall near the stagnation area by destroying the spray core and promoting the spray atomization. Figure 26 shows the influence of the spray impinging plate on NOx-Smoke trade-off characteristics with the nozzle opening pressures of both Po15/20 MPa and Po25/30 MPa. In the case of Po25/30 MPa, the smoke is reduced by 70% as expected. In the case of Po15/20 MPa, the smoke reduction is not so large as that of Po25/30 MPa. This is considered to be so because the amount of formed dense soot is originally small in Po15/20 MPa due to the relatively weak spray-wall impingement. Figure 27 shows the changes of combustion characteristics caused by the spray impinging plate. It is clear that spray impinging plate remarkably extends the combustion period and increases fuel consumption and also decreases the temperature in the swirl chamber. These changes indicate that the spray impinging plate deteriorates combustion. This deterioration of combustion is considered to be caused by the attenuating swirl flow velocity due to the spray impinging plate, which is explained later based on the observation results. However, from another point of view, the large smoke reduction in spite of the deterioration of microscopic combustion is considered to prove that the fuel adhesion on the chamber wall and the rich mixture stagnation are the main cause of the smoke formation. The observation results in the case of this spray impinging plate are shown in the right line of Fig. 23. In this observation, the injector of Po25/30 MPa was used, because the smoke reduction by the spray impinging plate with Po25/30 MPa is larger than that with Po15/20 MPa as shown in Fig. 26. With the impinging plate, the fuel spray impinges mainly to the plate and fuel droplets spread in the wide area during the main injection period. Therefore, the spray-wall impinging is prevented to suppress the formation of dense soot (11.3 ° ATDC).
3
i
°:t 0
3
Thus, compared to the case without the spray impinging plate (center line in Fig. 23), it is clear that the quantity of soot cloud which is carried along the chamber wall and spreads to the whole swirl chamber (19.8 ° to 32.8 ° ATDC) is remarkably reduced. Also from Fig. 23, it is shown that the spray impinging plate decelerates flame development by approximate 2 ° CA. This indicates the decrease of the swirl as mentioned above. This attenuation of the swirl flow not only deteriorates the combustion but also decelerates the soot oxidization process, so that the soot cloud does not fully disappear (56.8 ° ATDC) in spite of the remarkable decrease of the dense soot formation.
8. Smoke reduction mechanism of increasing swirl chamber volume ratio
Smoke reduction mechanism of increasing swirl chamber volume ratio (hereafter designated as /3 up) is also examined from the aspect of the dense soot formation near
15/20MPa Conve(~Uonal With• Plate 2~OMF,,,
,4
~.
~,
/ r 6001 ~
"~
Fig. 25. Schematic diagram of spray impinging plate.
i
2
Fig. 26. Improvement of NO~ -Smoke trade-off by spray impinging plate.
=
44 1901 ,
Impinging Plate
I
1
NOx g/kwh
8
With
|
170 /
0
.
,
10
20
EGR
%
30
Fig. 27. Comparison of combustion characteristics between with and without spray impinging plate.
30
Y. Hotta et al./JSAE Review 18 (1997) 19-31 Base 25/30Mpa 15/20MP,
~ A
3t '
!
I I
Large
,
• A.
!
100 200 300 400 NOx ppm
mixture formed near the spray-wall impinging point also decreases. Second, the distance from the injector to the spray-wall impinging point is increased by 2 mm due to the enlarged swirl chamber, so that not only the amount of oxygen entrained into the fuel spray increases before impinging the wall, but also the momentum of the fuel spray at the impinging point decreases. Therefore, both the decrease of the equivalence ratio and the suppression of the fuel adhesion to the chamber wall are achieved at the same time. Hence, the formation of the dense soot is suppressed.
Fig. 28. Improvement of NO~ -Smoke trade-off by enlarged swirl chamber.
9. Conclusion the spray-wall impinging point. In this experiment, /3 was increased up to 60% from 52% of the base condition. The whole swirl chamber is enlarged while keeping a similar shape to the base, and the depth of the whole clover-shaped main chamber is decreased, in order to maintain the same compression ratio. Comparison of NOx-Smoke trade-off characteristics is shown in Fig. 28. It is shown that smoke is reduced by 36% in the case of Po25/30 MPa, and 33% in the case of Po15/20 MPa respectively. Thus, it is confirmed that /3 up is quite effective to decrease the exhaust smoke. The change in the combustion peocesses caused by /3 up is shown in Fig. 29. It is clear that (1) during the main injection period, the amount of dense soot formed near the spray-wall impinging point is remarkably decreased, and that (2) the oxidation still continues at the later time of the combustion period. It is considered that the phenomenon of (2) is caused by the decrease of the fuel-air equivalence ratio in the swirl chamber as was expected. On the other hand, the phenomenon of (1) is considered to be caused by the following two mechanisms. First, the amount of the oxygen entrained into the fuel spray increases due to the decreased fuel-air equivalence ratio in the swirl chamber, especially at the later part of the injection period when the fuel is injected into the flame. Therefore, the equivalence ratio of the excess rich fuel-air
Fig. 29. Comparison of c o m b u s t i o n p r o c e s s e s between conventional and large swirl chamber (Po25/30MPa, EGR 11%).
Some means for simultaneous reduction of NO x and PM from swirl chamber type diesel engine were obtained using a multi-cylinder engine. The mechanisms of the emission improvement were examined using an optically accessible, single-cylinder engine. Through this study, the following points are clarified. (1) Suppression of the initial injection rate reduces both smoke and PM but not NOx. This tendency is opposite to that in direct injection diesel engines, and is peculiar to the swirl chamber type diesel engines. The smoke and PM reduction is caused by the following mechanism. (a) The fuel-air mixture formed during the ignition delay is decreased, and the ignition point moves from the vicinity of the injector to the spray tip. Therefore, the flame envelops the whole mixture area. This prevents the unburned fuel from flowing out into the main chamber prior to the flame and also from spreading into the lowtemperature areas in the main chamber which are not covered with flame throughout the combustion period. This leads to the SOF reduction. (b) The hard spray-wall impinging at the beginning of the injection period is prevented, so that the soot formation near the spray-wall impinging point is prevented at the beginning of the combustion period. Therefore, flowout of the dense soot occurs sufficiently after the flowout of the flame which raises the temperature of the main chamber. Thus, the soot cloud is easily oxidized in the main chamber. (c) The quantity of the formed dense soot itself is reduced due to the decrease of the main injection period in which the hard spray-wall impingement takes place. Therefore, exhaust smoke is reduced. (2) Increased injection pressure by using high nozzle opening pressure increases exhaust smoke. This is because a great deal of dense soot is formed by the hard spray-wall impinging due to the increased fuel spray momentum based on the characteristics of throttle nozzle. Therefore, not increasing but decreasing the injection pressure is effective for smoke reduction as long as a conventional throttle type injection nozzle is used. (3) Increasing swirl chamber volume ratio decreases exhaust smoke and PM especially over medium load con-
Y. Hotta et aL / J S A E Review 18 (1997) 19-31
dition. This is caused not only by the promotion of soot oxidation due to the decreased equivalence ratio in the swirl chamber, but also by the decrease of the formed dense soot near the spray-wall impinging point. In addition, NO x reduction is also achieved when increasing the EGR rate, which is possible due to the great reduction of smoke in this case. (4) Over medium load conditions, the main cause of the exhaust smoke is both the fuel adhesion on the chamber wall, and the stagnation of rich fuel-air mixture. Therefore, for example, the prevention of the spray-wall impingement by destroying the fuel spray core using a spray impinging plate is effective to reduce exhaust smoke. (5) The methods for simultaneous reduction of NO x and PM clarified through this study are summarized in Fig. 30. •
EGR
Injection Rate
ction I
Increase of [ Swid Chamber ~ VoumeRatio I
[
Decrease of ~ E ~ O F [ inSwirlChamber I
~eu~t~c~ioun I .......... =
~l Reduction I '
References [1] Yoshida, T. et al., Development of a Low Exhaust Emission and High Performance Turbocharged Diesel Engine (in Japanese with English summary), Journal of JSAE, 47-10 pp. 10-16 (1993). [2] Shinzawa, M. et al., Effects of Combustion Improvement on Reducing NOx Emissions of IDI Diesel Engines (in Japanese with English summary), Proceedings of the 1 lth Internal Combust. Engine Symp., Japan, pp. 357-362 (1993). [3] Walker, A. et al., High Performance Diesel Catalysts for Europe Beyond 1996, SAE Paper 950750 (1995). [4] Kawamura, K. et al., Analysis of Fuel Spray Characteristics under High Pressure Injection (in Japanese with English summary), Proceedings of JSAE, No. 911076 (1991). [5] Nakakita, K. et al., Optimization of Pilot Injection Pattern and Its Effect on Diesel Combustion with High-Pressure Injection, JSME International Journal, Series B, Vol. 37, No. 4 (1994). [6] Streeter, V.L. and Wylie, E.B., Hydraulic Transients, McGraw-Hill, New York (1967). [7] Nakakita, K. et al., A Study on Diesel Combustion with High-Pressure Fuel Injection (Improvement of Combustion and Exhaust Emissions Using Small-Hole-Diameter Nozzles) (in Japanese with English summary), Trans. Jpn. Soc. Mech. Eng., Series B, Vol. 60, No. 577, pp. 3198-3206 (1994).
'
Fig. 30. Mechanism of NO x and PM reduction.
Overseas Business Divl . . . . 3 - 1 0 , Fukuura, Kanazawa-ku, Yokohama 236 U RL: http://www, nhkspg, co .jp/
4
31
JSAE Review 18 (1997) 19-31
Combustion improvement for reducing exhaust emissions in IDI diesel engine Yoshihiro Hotta a, Kiyomi Nakakita a Minaji Inayoshi a, Takashi Ogawa b, Takeshi Sato b, Mitsumasa Yamada b a Mechanical Engineering Div. 1, Toyota Central Res. & Develop. Labs., Inc., 41-1 Yokomichi, Nagakute, Nagakute-cho, Aichi-gun, Aichi, 480-11 Japan b Engine Engineering Div. Ill, Toyota Motor Corp., 1200 Mishuku, Susono-shi, Shizuoka, 410-11 Japan
Received 10 May 1996
Abstract Means for reducing PM from a swirl chamber type diesel engine were examined and the mechanisms of the PM reduction were investigated using both a multi-cylinder and an optically accessible single-cylinder engine. The following points were clarified. (1) At light load, suppression of initial injection rate reduces PM, because of both the change of ignition point to reduce SOF, and the retarded flowout of dense soot from the swirl chamber to reduce smoke. (2) Over medium load, the main cause of the exhaust smoke is hard spray-wall impingement which leads to both fuel adhesion on the wall and the stagnation of rich fuel-air mixture.
1. Introduction
As swirl chamber type diesel engines are widely used for both passenger cars and commercial cars of up to medium-sized class, the reduction of their exhaust emissions is an urgent problem to be tackled. Therefore at present, various researches have been progressing [1-3] from the aspect of both combustion improvement and aftertreatment. This study aims at the simultaneous reduction of NOx and particulate matter (PM) by improving the combustion itself. First, some means for simultaneous reduction of NO x and PM were obtained by investigating the effects of injection characteristics such as injection rate and injection pressure and of swirl chamber volume ratio, using a multi-cylinder engine. Then, the mechanism of emission improvement was examined in detail using an optically accessible, single-cylinder engine.
(1) NO x is reduced by decreasing the pre-mixed burning rate caused by the suppression of initial injection rate. (2) Smoke is reduced by promoting both fuel atomization and air entrainment into the fuel spray caused by the increase of injection pressure [4]. (3) Smoke is reduced by decreasing equivalence ratio in the swirl chamber caused by the increase of swirl chamber volume ratio. The smoke reduction in (2) and (3) is expected to lead not only to the PM reduction, but also to the NO x reduction caused by the possible increase of EGR rate. Whether these hypothetical mechanisms can be actually realised is examined through multi-cylinder engine experiments in Section 4.
I Low Initial
}
Injection Rate I
= Reduction of L ~ Premix Burning I ~ EGR
2. Concept for simultaneous reduction of NO x and PM
A concept for simultaneous reduction of NO x and PM was hypothesized as shown in Fig. 1. This consists of the following NOx/PM reduction mechanism.
High Injection ~ Pressure I Swirl Chamber
Volume Ratio
Fuel Spray L Increase/ I Atomization, etc I ~ ] ' ~ " ~
NOx ...... I /I ........... I
~" Equivalence Ratio [
in Swirl Chamber I
I Reduction J
"
-
Fig. 1. Concept for NO~ and PM reduction.
0389-4304/97/$17.00 Copyright © 1997 Society of Automotive Engineers of Japan, Inc. and Elsevier Science B.V. All rights reserved Pll S 0 3 8 9 - 4 3 0 4 ( 9 6 ) 0 0 0 5 0 - 1 JSAE9730038
20
Y. Hotta et al./JSAE Review 18 (1997) 19-31 1450 rpm 20 %load
Table 1 Specifications
Bore X stroke (ram) Number of cylinders Displacement (cc) Air aspiration system Compression ratio Swirl chamber volume ratio (%) Injection pump Injection nozzle type Nozzle-opening pressure (MPa) Prelift (mm)
2000 rpm 50 %/load
of multi-cylinder e n g i n e 96 × 103 4 2982 Turbocharged 21.2 58, 62
-. a~~
5
2
_~
E
om
~
= ' 0 ~-20 Distributor type Throttling pintle nozzle 1 Spring 2 Spring t5 (base) -
15/20 20/25 0.2
0 15/25 25/30
3. Experimental apparatus and conditions using multicylinder engine The specification of the multi-cylinder engine and the injection system used in this study are listed in Table I, and the schematic diagram of the swirl chamber configuration is shown in Fig. 2. An ordinary single-spring injector and also several double-spring injectors were used, in order to change both the injection rate and the injection pressure. Engine running conditions were selected as 20% load at 1450 rpm as the light load condition and 50% load at 2000 rpm as the medium load condition, which are representative of conditions in actual use. Injection timing was adjusted so as to optimize the NOx-PM trade-off characteristics for each case.
4. Results and discussion on multi-cylinder engine experiment
~£ " Fig. 2. Schematic diagram of swirl chamber (multi-cylinder engine).
0 10 20 30 Crank Angle deg
Fig, 3. Comparison of initial injection characteristics,
increased up to 20 or 25 MPa, while the first opening pressure was maintained at the same pressure as that of the single-spring injector (hereafter Po 15 / 20 MPa, Po 15/25 MPa respectively). The duration of suppressed initial injection rate increases in the order of Po15 MPa, P o 1 5 / 2 0 MPa and Po15/25 MPa, due to the function of the double-spring nozzle holder. Figure 4 shows the effects of initial injection rate suppression on exhaust emissions under both light load and medium load. Against expectation, but not NO X, both smoke and PM are reduced with the injection rate suppression. Under light load, both smoke and PM decrease remarkably depending on the duration of the initial-injection-rate suppression. This cause is examined in detail in Section 6. Under medium load, the tendency is slightly different. In the suppressed initial-injection-rate case of Po15/20 MPa, both smoke and PM are clearly reduced compared with the base case of Po15 MPa. However, in the further suppressed initial-injection-rate case of Po15/25 MPa, smoke and PM are hardly reduced. One reason for this is consid-
4.1. Effects of suppressing initial injection rate In this section, the hypothetical mechanism (1) in Section 2 is evaluated. The evaluated injection characteristics are shown in Fig. 3. The single-spring injector with nozzle opening pressure of 15 MPa (hereafter designated as Po15 MPa) was selected as the base injection system. For the double-spring injectors, the second opening pressure was
I 0 10 20 30 Crank Angle deg
E / o:Pol5MPa o.. 200 I- zx : 15•20 o_ | o: 15/25 0"r'
O3
2000 rpm 50 %load
1450 rpm 20 %load
300 j
-
I
~ I
I
20
30
0
0
10
EGR %
40
0
2
4
6
8
EGR %
Fig. 4. Effects of initial injection rate suppression on e x h a u s t e m i s s i o n s .
Y. Hotta et al./ JSAE Review 18 (1997) 19-31
21
1450 rpm 20 %load 300 r / o : Po15/20MPa & 200 ~- zx: 20/25 o_ / n: 25/30
E
_
Cq
o 100~ . . _ . . . , ~ , , ~ 0 i
t
t
2000 rpm 50 %load
t
/
O I---
1400 rpm
2000 rpm
Fig. 5. Effectsof initial injection rate on full load torque.
oE ered to be that the injection duration of Po15/25 MPa under medium load becomes too long to fully oxidize the fuel injected at the latter part of the injection period. Figure 5 shows the maximum torque in each case under the limited smoke level of 1.5 BSU. The maximum torque decreases with suppression of the initial injection rate. This is caused by the fact that the injection amount must be reduced when suppressing the initial injection rate to keep the limited smoke level, because the excess elongation of suppressed-injection-rate duration increases the smoke due to the excessively long injection duration, especially under high load. In the case of a double-spring injector, injection pressure at the latter part of the injection period becomes higher due to both the high second nozzle opening pressure and the use of the higher injection rate range on the pump cam. Therefore, the decrease of smoke and PM mentioned above are obtained by the effects of both the initial injection rate suppression and the injection pressure increase. This effect of injection pressure is clarified in the next section. 4.2. Effects o f increasing injection pressure
In this section, the hypothetical mechanism (2) is evaluated. P o 1 5 / 2 0 MPa was selected as the base injector, because it has both the remarkable effect of PM reduction under light and medium loads and the least increase of smoke under high load. In order to evaluate the effects of injection pressure increase, exhaust emissions of Po20/25 MPa and P o 2 5 / 3 0 MPa were compared to that of the base.
1450 rpm 20 %load
2000 rpm 50 %/load
0
10
20
30
EGR %
40
0
2
4
6
Fig. 7. Effectsof injection pressureon exhaust emissions.
The difference between the first and the second nozzle opening pressures of these injectors were kept constant so as to prevent the effects of initial injection rate suppression from changing. As shown in Fig. 6, the injection pressure increases throughout the injection duration while the suppressed-injection-rate duration is maintained as the same as that of the base. Exhaust emissions under both light and medium loads are shown in Fig. 7. Against expectation, under medium load, both smoke and PM increase remarkably as the injection pressure increases. Furthermore, even under light load, smoke and PM reductions are hardly achieved. This smoke increase tendency is more remarkable under heavy load. This causes a remarkable decrease of the maximum torque under the limited smoke level of 1.5 BSU as shown in Fig. 8. This mechanism of smoke and PM increase is examined in detail in Section 7. Based on the results of this section, the following points are revealed. (1) Smoke and PM reductions by suppressing initial injection rate shown in the previous section are caused by overcoming the increase of smoke and PM by the injection pressure increase at the latter part of the injection period. (2) Smoke and PM increases under the excess suppression of initial injection rate at full load condition (Fig. 5)
~rr~E "e E ,- ~ 4
0
[
E
~
Zr
O
o o
~'~"
I/ 01
~ P o 15/20 MPa t
i
t
x
t
0 10 20 30 CrankAngle deg
0 10 20 30 Crank Angle deg
Fig. 6. Comparisonof injection pressurecharacteristics.
8
EGR %
O F-
1400 rpm
2000 rpm
Fig. 8. Effects of injection pressureon full load torque.
22
Y. Hotta et al./JSAE Review 18 (1997) 19-31 1450 rpm 20 %load 300 I Po 15/20 MPa
2000 rpm 50 %load
|
200 I- o: a58% /
0~
z~: 62
f
~'/
i
i
i
Smoke1.5BSU Po15/20MPa #' 58% 62
E ¢~
[--
g
m.......
I---
1400 rpm i A
v
Fig. 10. Effects of swirl chamber volume ratio ( / 3 ) on full load torque.
I
0
10
20 EGR %
2000 rpm
30
40
0
4 6 EGR %
10
Fig. 9. Effects of swirl chamber volume ratio ( f l ) on exhaust emissions.
are considered to be caused not only by the excessively elongated injection period, but also by the increased injection pressure at the latter part of the injection period as mentioned in Section 4.1.
4.3. Effects of increasing swirl chamber volume ratio In this section, the hypothetical mechanism (3) is evaluated. Swirl chamber volume ratio (hereafter designated as /3) was increased up to 62% from 58% of the base condition while keeping the same compression ratio. Concerning the injector, Po15/20 MPa was selected considering the above results. Figure 9 shows the comparison of exhaust emissions between /3 = 58% and /3 = 62% under both light and medium loads. It is clear that the reductions of smoke and PM by increasing /3 at medium load are much larger than that at light load. Furthermore, these reductions become larger with increasing EGR rate. These results indicate that the effect of increasing /3 on smoke and PM reductions becomes greater as the equivalence ratio increases. Concerning the PM reduction in this case, it is considered to be caused not only by soot reduction but also by SOF reduction because HC is also reduced. NO x increases with increasing /3 in the range of low EGR rate at medium load. However, simultaneous reduction of NOx and PM is consequently obtained because this NO~ increase can be cancelled out by increasing EGR rate as follows. When the emissions at several EGR rate conditions of /3 = 62% are compared with the emission at 2.5% EGR rate of /3 = 58% regarded as the base condition, then (1) at the EGR rate of 2.5%, smoke and PM are remarkably reduced, while NO x is higher. (2) with increasing EGR rate up to 4%, smoke and PM are reduced without increasing NO x.
(3) during the range of EGR rate from 4% to 8%, simultaneous reduction of smoke, PM and NO x is obtained. Additionally as shown in Fig. 10, maximum torque under the limited smoke level of 1.5 BSU increases by increasing /3 due to the smoke decrease. Therefore, increasing /3 enables the increase of maximum torque in spite of using the double-spring injector which decreases the maximum torque as shown in Section 4.1 (Fig. 10, /362%-Po15/20 MPa vs. /358%-Po15 MPa).
4.4. Simultaneous reduction of NO, and PM The results in Sections 4.1 to 4.3 show that simultaneous reduction of NO x and PM by combining the suppression of initial injection rate with the increase of swirl chamber volume ratio was confirmed, as shown in Fig. 11. In this measurement, the EGR rate was adjusted so as to keep the exhaust smoke level of 0.5 BSU. A is the base case, and B is the case of suppressing initial injection rate. C is the case of combining the increase of swirl chamber volume ratio with B. Comparing C with A, it is clear that the reductions are approximately 40% of NO x and 20% of PM at light load, 40% of NO, and 35% of PM at medium load, respectively.
35 -1450 - rpm - 20 - %load
~om3Ot rsmoke0.SBSU , ~ ~
10 2000 rpm 50 %load
] ~ ~
5f-~° k ~
~ 1.0 •~ 0.5
{,op osF
1I
A B C A B C A: Base B: Suppression of Initial Injection Rate (Po15/20 MPa) C: B Jr Increase of Swirl Chamber Volume Ratio (/362 %) *)SOF : Soluble Organic Fraction *)ISOF: In-Soluble Organic Fraction
Fig. 11. NO x and PM reduction with combination of suppressing initial injection rate and increasing swirl chamber volume ratio.
Y. Hotta et al./JSAE Review 18 (1997) 19-31
Mirror
Table 2 Specifications of single-cylinderengine Bore× stroke Displacement Compression ratio Swirl chambervolumeratio
23
92 mm× 92 mm 612 cc 20.8 52%
Half Mirror
u It is also clear that the cause of NO x reduction is the possible increase of EGR rate due to the decrease of PM. Therefore in the following sections, the mechanisms of these PM reductions are examined in detail mainly by in-cylinder visualization analysis.
5. Apparatus and procedure for in-cylinder observation
Convex,ens
V High Speed Camera
Fig. 13. Schematic diagram of shadow-graph photography.
lyzed based on the in-cylinder pressure which was recorded by a mini-computer every 0.5 ° CA throughout 200 cycles.
5.1. Optically accessible single-cylinder engine 5.2. Apparatus for visualization The specification of the optically accessible engine is listed in Table 2, and the schematic diagram of the swirl chamber configuration is shown in Fig. 12. The swirl chamber is constructed as a removable unit in order to evaluate several chamber designs, and therefore the connecting passage is longer than that of the multi-cylinder engine used in the previous section. Also, this swirl chamber is designed to be two-dimensional, with flat side walls for visualization by the shadow-graph method. A piezo type pressure transducer and a glow-plug shaped thermocouple are attached to measure gas pressure and representative atmospheric temperature in the swirl chamber respectively. In the visualization experiments, the engine with an elongated piston was kept motoring with control of both coolant and lubricant at 80°C and with control of intake air at 55°C. Just when the representative temperature of the swirl chamber reached 410°C, fuel injection was started. The exhaust gases and the combustion characteristics were also examined under steady-state running condition with an ordinary short piston instead of the elongated piston. Combustion characteristics were ana-
Shadow-graph photographs were taken with the apparatus shown in Fig. 13, at a frame speed of 9000 fps using a high-speed camera. A tungsten halogen lamp was used for lighting. The observation field is shown in Fig. 14. Both the whole swirl chamber area and the center area of the main chamber whose diameter is 75% of the cylinder bore were simultaneously observed. 5.3. EGR method In the visualization experiments, if real exhaust gas was recirculated just like ordinary EGR, the observation windows would be dirtied by the formation of dense soot before reaching the required EGR rate, so that clear observation would be impossible. Hence, to simulate the required EGR condition in this study, the exhaust gas under preliminary steady engine running was transferred into a 900-liter sampling bag after removing soot, oil-mist and water. Then, the stored clean exhaust gas was mixed into the intake air in the visualization or measurement of emissions.
Main Chamber
Swirl Chamber
. . - ~ 7---'~Val, res
~x
~
iI I
I //
Throat
~Crank Angle
-,2-'_2-"
Throat Location Fig. 12. Schematicdiagramof swirlchamber(single-cylinderengine).
Fig. 14. Observationfield.
24
Y. Hotta et a l . / J S A E Review 18 (1997) 19-31
Table 3 Experimental conditions for analysis of the effect of initial injection rate suppression (light load) Engine speed Volumetric efficiency Fuel amount Equivalence ratio Injection timing EGR rate Nozzle-opening pressure
1600 rpm 100% 20 mm3/st 0.34 - 4° ATDC 0-24% 1 Spring 15 MPa
3
2 Spring 20/30MPa
3 t/)
=
6. P M reduction m e c h a n i s m o f initial injection rate suppression
2
,
,
.
|
|
|
Conventional Inj. Rate ,.,.,~
o
E
1
Low Initial Inj. Rate |
6.1. Exhaust emissions and combustion characteristics
o
1'o EGR
The smoke and PM reduction mechanisms by the initial injection rate suppression were investigated under the light load conditions listed in Table 3, where the largest reduction of PM was obtained in Section 4.1. A single-spring injector with nozzle opening pressure of 15 MPa was used for the representative of conventional injection rate (hereafter Conventional), and a double-spring injector with nozzle opening pressures of 2 0 / 3 0 MPa was used for the representative of the suppressed initial injection rate (hereafter Suppressed). A comparison of the injection pattern between Conventional and Suppressed is shown in Fig. 15. The characteristics of Suppressed are as follows. (1) The injection rate at the first half of the injection period is suppressed quite low and kept nearly constant, and its peak value at the latter part is only 54% of the Conventional. (2) The injection period is extended to 24 ° CA from 14° CA of the Conventional. (3) The peak injection pressure increases to 40 MPa from 22 MPa of the Conventional. Comparisons of exhaust emissions between Conventional and Suppressed are shown in Fig. 16. --
~..~ .g
Conventional Inj. Rate Low Initial Inj.Rate
2
10[ 0.5
z
0
.-
-4°f
E "" ~" 2 0
o
Fig. 15. Comparison of injection characteristics at light load.
30 %
Fig. 16. Effects of initial injection rate suppression on exhaust smoke.
It was confirmed that the initial injection rate suppression reduces smoke and HC but not NO x, the same results as for the experiments with multi-cylinder engine in Section 4.1. It is considered that reduction of NO x was not obtained due to the increased injection pressure, which canceled the NO x reduction effect caused by the suppressed initial injection rate and the extended injection period. Additional characteristics of the emissions are as follows. (1) In the case of Conventional, smoke level is fairly high and increases rapidly with EGR rate. In the case of Suppressed, smoke level is remarkably low and hardly increases with increasing EGR rate. (2) HC level of Suppressed is approximately half of that of Conventional. Comparisons of the heat release pattern and the combustion characteristics are shown at the top of Fig. 18, and in Fig. 17, respectively. Figure 18 shows that the heat release of Suppressed is lower at the first stage and higher at the second. On the other hand, Fig. 17 shows that the initial injection rate suppression increases the ignition delay period only by 2 degrees, and extends the combustion period only by 1 degree. These observations indicate that the initial injection rate suppression retards main combustion time while keeping almost the same ignition delay and combustion period. Also injection timing retard causes the retard of main combustion time, but it increases fuel consumption due to the excessively elongated ignition delay and combustion period, even though having an advantage of smoke reduction. The initial injection rate suppression does not cause such undesirable combustion and therefore the fuel consumption is maintained at the same level as that of Conventional as shown in Fig. 17. This characteristic is con-
25
Y. Hotta et al.//JSAE Reoiew 18 (1997) 19-31 Zx Conventional Inj. Rate O Low Initial Inj. Rate
~..= 180 -~.== 160 I
I'~
~ 8
•
,
I
|
•
8o
6.2. PM reduction mechanism
i
i
i
|
m
•
|
•
i
!
=
=
Mixture formation and flame development processes in the swirl chamber and in the main chamber are shown in Fig. 18 and Fig. 19 respectively. The differences between Conventional and Suppressed, and PM reduction mechanisms are as follows.
|
t
31 29
sidered to be one of the main advantages of the initial injection rate suppression.
I
33}
E P~ 120 •~ 0. z=
,
~
o
-O ~
i
i
i
0
10
20
(1) SOF (Soluble Organic Fraction) reduction mechanism. In the case of Conventional, a lot of fuel vapor is formed even from the early stage and the ignition occurs near the injection nozzle (at 1.3 ° ATDC of Conventional, in Fig. 18), so that unburned mixture exists ahead of the flame. This unburned mixture flows out to the main chamber prior to the flame. In the main chamber, there exist some areas where flames do not spread throughout the
|
t
30
EGR %
Fig. 17. Effects of initial injection rate suppression on combustion characteristics.
Conventional Inj. Rate [Conventional Inj. Rat [ EGR 0% EGR 24"~ I
/Jk\t
-40
0
40
=
80 ~
Crank Angle "ATDC
i
"I"
"
-40
0
5.0"
-40
15.5"
~ 9 . 8 "
0
40
/ i
5.5" Outflow of Flame
~ 1 o . ~
15.8" ~
80 :~
Crank Angle "ATDC
: .
Outflow of Dense Soot
9.5"
°
40"-~'-"'~0
Crank Angle "ATD~
m
4.7"
Low Initial Inj.Rate-] EGR 240/,
16.3"
I
"Outflow o f
[]
Dense Soot
Fig. 18. Heat release rate and combustion process in swirl chamber.
26
Y. Hotta et al./JSAE Review 18 (1997) 19-31
Conventional Inj. Rate EGR 24% Luminou., flame 8.5*
L o w Initial Inj. Rate EGR 24%
ATDC Outflow of Dense sool
11.5"
13.8"
16.9"
22.0*
Fig. 19. Comparisonof flame developmentin main chamber(EGR 24%).
combustion period under this light load condition (at 16.9 ° ATDC of Conventional, in Fig. 19). Hence, unburned mixtures that spread into these areas are considered to be not oxidized completely and exhausted. In the case of Suppressed, the amount of formed fuel vapor is fairly small at the early stage and the ignition occurs near the spray tip (at 1.8 ° ATDC of Suppressed, in Fig. 18), so that there are no unbumed mixtures ahead of the flame. This is considered to be the cause of HC and SOF reduction. The cause of shifting the ignition point is considered to be as follows. In the case of Conventional, injection rate increases sharply at the beginning of the injection period as shown in Fig. 4, so that the momentum of the fuel spray also increases sharply. Therefore, the fuel injected at the beginning is overtaken by the fuel injected later, and carried near the vicinity of the nozzle. The ignition is considered to occur at the mixture formed by the fuel that is injected at the beginning and stayed for a long time in a high-temperature atmosphere, so the ignition occurs at the vicinity of the nozzle. In the case of Suppressed, it is considered that the overtaking does not occur on a large scale, because the injection rate is suppressed to quite low and is nearly constant during the suppression period, so
that the fuels are sequentially carried by the strong swirl. Therefore, the fuel which is injected at the beginning remains near the spray tip to cause ignition.
(2) Soot (Insoluble Organic Fraction) reduction mechanism. In the case of Conventional, fuel spray impinges against the chamber wall opposite t o t h e injection nozzle from the beginning due to the high injection rate. Therefore in Fig. 18, a dense soot cloud shown by the dark area, which is peculiar to the case of high exhaust smoke and is considered to be generated by the cooling of the dense soot on the quartz window [5], appears near the spray-wall impinging point from the early stage of the combustion period, and some part of this dense soot flows out to the main chamber just when the initial flame flows out (at 5.0 ° ATDC of Conventional, in Fig. 18). This is confirmed by the fact that the dense soot is observed in the main chamber at almost the same time of the flowout of the flame (at 8.5 ° ATDC of Conventional, in Fig. 19). As the formation of the dense soot continues throughout the injection period due to the continuous hard fuel-spray impinging, the dense soot continuously flows out to the main chamber and can not be fully oxidized. This incomplete oxidation is shown by the fact that the dense soot does not disappear and spreads to the whole area of the main chamber (at 16.9 ° ATDC of Conventional, in Fig. 19). In the case of Suppressed, the flame develops along the swirl chamber wall without the spray-wall impinging during the suppression period due to the quite low injection rate (at 3.0 ° ATDC of Suppressed, in Fig. 18). Therefore, the flame with a sparse rather than a dense soot cloud flows out into the main chamber, unlike in the case of Conventional (at 5.5 ° ATDC of Suppressed, in Fig. 18). This luminous flame changes into either burned gas or nonluminous flame by sufficient oxygen as soon as it flows out into the main chamber. This burned gas or nonluminous flame is shown as the shadow image caused by the density gradients near the throat (at 8.3 °, 1 1.7 °, and 14.0 ° ATDC of Suppressed, in Fig. 19). The dense soot cloud is formulated also in this case due to the hard spray-wall impinging during the main injection period which corresponds to the second stage of nozzle needle lift (at 16.3 ° ATDC of Suppressed, in Fig. 18). However, the quantity of the dense soot is less than that in Conventional, because the main injection period is shorter under this light load condition. Additionally, when this dense soot cloud flows out to the main chamber (at 17.2 ° ATDC of Suppressed, in Fig. 19), the temperature in the main chamber has already been raised by the burned gas and the flame which flowed out prior to the dense soot. In other words, the potential for oxidizing soot in the main chamber has already been raised to the level enough to oxidize the dense soot cloud easily. Additionally, the cause that smoke rapidly increases with increasing EGR rate in the case of Conventional but
27
Y. Hona et aL / J S A E Review 18 (1997) 19-31
Table 4 Experimental conditions for analysis of the effects of injection pressure and swirl chamber volume ratio (medium load) Engine speed Volumetric efficiency Fuel amount Equivalence ratio Injection timing Nozzle-opening pressure
1800 rpm 100% 30 m m 3 / s t 0.52 - 8° ATDC 2 Spring 15/20 MPa
~ero~
';o-:E'~'0.6 -w I/I ' ~ ~0.4 ~'~ 0.21 , "~
2 Spring 25/30MPa
7. S m o k e formation m e c h a n i s m under m e d i u m load condition
7.1. Main cause of smoke increase with increased injection pressure
The mechanism of the smoke increase with increased injection pressure was examined under medium load conditions, where the smoke remarkably increases as shown in Section 4.2. The experimental conditions are listed in Table 4. The nozzle opening pressures of the double-spring
'
t
• 15/20MPaI
.
~
hardly increases in the case of Suppressed, is considered to be as follows. It is clear by comparing the left and middle lines in Fig. 18 that EGR delays flame development. Accordingly, in the case of Conventional, at the beginning of the combustion period under EGR condition, the quantity of the flame is relatively too small to fully oxidize the dense soot cloud, so that the exhaust smoke increases. On the other hand, in the case of Suppressed, flame development delays under EGR conditions quite similarly to the case of Conventional. However, flowout of the dense soot cloud begins sufficiently later than that of the flame, so that the slight delay of the flame development does not have much influence on the soot oxidization process. Therefore, exhaust smoke is kept to a low level.
'
46 4
'
'
'
0
10
20
30
EGR % Fig. 21. Effect of nozzle opening pressure on combustion characteristics.
injectors are 15/20 MPa and 2 5 / 3 0 MPa. The comparison of the exhaust emissions are confirmed as shown in Fig. 20. It is clear that the increased injection pressure remarkably increases both smoke and HC just as in the case of the multi-cylinder engine in Section 4.2. Figure 21 shows the comparison of the combustion characteristics. It is clear that the injection pressure has little influence on the combustion characteristics such as ignition delay and the combustion period, except the maximum rate of pressure rise which increases approximately 35% with the increase of injection pressure. Little difference is recognized also in the heat release rate patterns in Fig. 22. In conclusion, it is revealed that the macroscopic combustion characteristics are hardly changed by the injection pressure increase, in spite of the remarkable smoke increase. Therefore, the combustion process was examined in detail by in-cylinder observation. The results are shown in Fig. 23. The photographs in the left and center line of Fig. 23 correspond to the case of P o 1 5 / 2 0 MPa and Po25/30 MPa, respectively. From these results, the following points are observed. During the initial-injection-rate suppression period (before 5 ° ATDC), there are no differences in both cases.
_~0.1.
X ol
I
|
|
I"
3
i:t
•
=~"~
°I 3
I
o.4t |
i
I
Z~,.,.,. ,.Z~ 25/30MPa
I
/
a: O.Ot
E~~MPa
0
.
o,8[
i
I
®. 90[ i~60
i
/
Fig. 20. Effect of nozzle opening pressure on exhaust emissions.
--
25130MPa
-- 15/20MPa
30
t
0 EGR 10 %20 30
/
"r
0
°
-40 0 Crank Angle
]
40 80 ° ATDC
Fig. 22. Heat release pattern for each nozzle opening pressure.
28
Y. Hotta et al./JSAE Review 18 (1997) 19-31
Namely, the ignition occurs near the spray tip, so that the flame successfully envelops the whole mixture area. Also, the fuel spray does not impinge on the chamber wall. Therefore, the dense soot cloud is not produced in both cases (2.5 ° ATDC). Thus, the required conditions for keeping the low PM level at light load shown in Section 6 are maintained in both cases. However, remarkable differences appear during the main injection period corresponding to the second needle lift (5°-17 ° ATDC). In this period, the fuel spray impinges on the chamber wall due to the increased injection rate, so that the dense soot cloud represented by the dark area is formed near the spray-wall impinging point in both cases (9.8 °, 9.7 ° ATDC). The difference is in the quantity of this dense soot cloud. In the case of Po25/30 MPa, the quantity of the dense soot is so large that the dense soot clouds are carried along the chamber wall (14.5 ° ATDC), and spread in the whole swirl chamber (32.4 ° ATDC). Thus, a large quantity of soot particles are observed in the latter part of the combustion period (57.7 ° ATDC). On the other hand in the case of P o 1 5 / 2 0 MPa, the quantity of the dense soot formed near the spray-wall impinging point is not so large as that in P o 2 5 / 3 0 MPa, so that far less soot clouds are carried along the chamber wall and spread in the chamber (17 ° to 32.6 ° ATDC). Thus, the soot clouds are easily oxidized to disappear (58.0 ° ATDC).
15,
2.4 °
ATDC
spray-wall ImpingementI • Less
spray-wailImpingement
I • Furious
spray-wallImpingement • Far less
..... .d~n..s.a.-soot .I.or..m.a.t.i.o. n..I ..... .d.a.n~?: .sp.ot1o.rma!l.o.n.......... .d.a.n~?: _s_ _o9_t_l_qr_.m_~]1_on__ 17.0"
19.8"
Fig. 23. Effects of nozzle opening pressure and spray impinging plate on combustion processes ( E G R 11%).
12
•
•
•
•
E
o80t
t.O
I
~:~ 40 0
•
-Oot 80
t
•
. . . .
-20 -10
0
10
20
30
Crank Angle •ATDC Fig. 24. C o m p a r i s o n of spray m o m e n t u m between high and low nozzle opening pressure.
In conclusion, the mechanism of this dense soot formation near the spray-wall impinging point is considered to be as follows. The hard spray impingement causes fuel adhesion on the wall near this point. This adherent fuel is not quickly evaporated and formed fuel vapor is hardly carried out of this area, because the stagnation zone is formed here due to the chamber shape. Thus, the rich fuel-air mixture stagnates in this zone under the condition of high temperature and insufficient oxygen to form the dense soot cloud. Though the hard spray-wall impingement is considered to be a main cause of the dense soot formation, the difference in formed soot amount between both cases is quite large. Therefore, the momentum of the fuel spray in each case was numerically analyzed by the method of characteristics [6] based on the experimental value of the nozzle needle lift and the pressure at the entrance of the injector. The results are shown in Fig. 24 as well as the pressure at the nozzle sac and the injection rate. It is clear that the momentum of the fuel spray at the main injection period of Po25/30 MPa is approximately twice as large as that of Po15/20 MPa. This is caused not only by the increase of injection pressure, but also by the increase of the injected fuel mass due to the increased effective flow area with increased nozzle needle lift. Thus, increased nozzle opening pressure causes drastic increase of spray momentum due to the characteristics of the throttle nozzle, so that the hard spray-wall impingement is caused to form a great deal of dense soot cloud. For really promoting fuel atomization and air entrainment while keeping the prevention of spray-wall impingement, it is considered that, for instance, a small-hole-diameter nozzle should be combined with this increased injection pressure just as in the case of the direct injection diesel engine [7].
29
Y. Hotta et al./JSAE Review 18 (1997) 19-31
7.2. Effect of spray impinging plate
WRh
Plate
A
In order to confirm the smoke formation mechanism mentioned above, the effect of a spray impinging plate shown in Fig. 25 was examined. This plate is intended to prevent the injected fuel droplets from adhering on the wall near the stagnation area by destroying the spray core and promoting the spray atomization. Figure 26 shows the influence of the spray impinging plate on NOx-Smoke trade-off characteristics with the nozzle opening pressures of both Po15/20 MPa and Po25/30 MPa. In the case of Po25/30 MPa, the smoke is reduced by 70% as expected. In the case of Po15/20 MPa, the smoke reduction is not so large as that of Po25/30 MPa. This is considered to be so because the amount of formed dense soot is originally small in Po15/20 MPa due to the relatively weak spray-wall impingement. Figure 27 shows the changes of combustion characteristics caused by the spray impinging plate. It is clear that spray impinging plate remarkably extends the combustion period and increases fuel consumption and also decreases the temperature in the swirl chamber. These changes indicate that the spray impinging plate deteriorates combustion. This deterioration of combustion is considered to be caused by the attenuating swirl flow velocity due to the spray impinging plate, which is explained later based on the observation results. However, from another point of view, the large smoke reduction in spite of the deterioration of microscopic combustion is considered to prove that the fuel adhesion on the chamber wall and the rich mixture stagnation are the main cause of the smoke formation. The observation results in the case of this spray impinging plate are shown in the right line of Fig. 23. In this observation, the injector of Po25/30 MPa was used, because the smoke reduction by the spray impinging plate with Po25/30 MPa is larger than that with Po15/20 MPa as shown in Fig. 26. With the impinging plate, the fuel spray impinges mainly to the plate and fuel droplets spread in the wide area during the main injection period. Therefore, the spray-wall impinging is prevented to suppress the formation of dense soot (11.3 ° ATDC).
3
i
°:t 0
3
Thus, compared to the case without the spray impinging plate (center line in Fig. 23), it is clear that the quantity of soot cloud which is carried along the chamber wall and spreads to the whole swirl chamber (19.8 ° to 32.8 ° ATDC) is remarkably reduced. Also from Fig. 23, it is shown that the spray impinging plate decelerates flame development by approximate 2 ° CA. This indicates the decrease of the swirl as mentioned above. This attenuation of the swirl flow not only deteriorates the combustion but also decelerates the soot oxidization process, so that the soot cloud does not fully disappear (56.8 ° ATDC) in spite of the remarkable decrease of the dense soot formation.
8. Smoke reduction mechanism of increasing swirl chamber volume ratio
Smoke reduction mechanism of increasing swirl chamber volume ratio (hereafter designated as /3 up) is also examined from the aspect of the dense soot formation near
15/20MPa Conve(~Uonal With• Plate 2~OMF,,,
,4
~.
~,
/ r 6001 ~
"~
Fig. 25. Schematic diagram of spray impinging plate.
i
2
Fig. 26. Improvement of NO~ -Smoke trade-off by spray impinging plate.
=
44 1901 ,
Impinging Plate
I
1
NOx g/kwh
8
With
|
170 /
0
.
,
10
20
EGR
%
30
Fig. 27. Comparison of combustion characteristics between with and without spray impinging plate.
30
Y. Hotta et al./JSAE Review 18 (1997) 19-31 Base 25/30Mpa 15/20MP,
~ A
3t '
!
I I
Large
,
• A.
!
100 200 300 400 NOx ppm
mixture formed near the spray-wall impinging point also decreases. Second, the distance from the injector to the spray-wall impinging point is increased by 2 mm due to the enlarged swirl chamber, so that not only the amount of oxygen entrained into the fuel spray increases before impinging the wall, but also the momentum of the fuel spray at the impinging point decreases. Therefore, both the decrease of the equivalence ratio and the suppression of the fuel adhesion to the chamber wall are achieved at the same time. Hence, the formation of the dense soot is suppressed.
Fig. 28. Improvement of NO~ -Smoke trade-off by enlarged swirl chamber.
9. Conclusion the spray-wall impinging point. In this experiment, /3 was increased up to 60% from 52% of the base condition. The whole swirl chamber is enlarged while keeping a similar shape to the base, and the depth of the whole clover-shaped main chamber is decreased, in order to maintain the same compression ratio. Comparison of NOx-Smoke trade-off characteristics is shown in Fig. 28. It is shown that smoke is reduced by 36% in the case of Po25/30 MPa, and 33% in the case of Po15/20 MPa respectively. Thus, it is confirmed that /3 up is quite effective to decrease the exhaust smoke. The change in the combustion peocesses caused by /3 up is shown in Fig. 29. It is clear that (1) during the main injection period, the amount of dense soot formed near the spray-wall impinging point is remarkably decreased, and that (2) the oxidation still continues at the later time of the combustion period. It is considered that the phenomenon of (2) is caused by the decrease of the fuel-air equivalence ratio in the swirl chamber as was expected. On the other hand, the phenomenon of (1) is considered to be caused by the following two mechanisms. First, the amount of the oxygen entrained into the fuel spray increases due to the decreased fuel-air equivalence ratio in the swirl chamber, especially at the later part of the injection period when the fuel is injected into the flame. Therefore, the equivalence ratio of the excess rich fuel-air
Fig. 29. Comparison of c o m b u s t i o n p r o c e s s e s between conventional and large swirl chamber (Po25/30MPa, EGR 11%).
Some means for simultaneous reduction of NO x and PM from swirl chamber type diesel engine were obtained using a multi-cylinder engine. The mechanisms of the emission improvement were examined using an optically accessible, single-cylinder engine. Through this study, the following points are clarified. (1) Suppression of the initial injection rate reduces both smoke and PM but not NOx. This tendency is opposite to that in direct injection diesel engines, and is peculiar to the swirl chamber type diesel engines. The smoke and PM reduction is caused by the following mechanism. (a) The fuel-air mixture formed during the ignition delay is decreased, and the ignition point moves from the vicinity of the injector to the spray tip. Therefore, the flame envelops the whole mixture area. This prevents the unburned fuel from flowing out into the main chamber prior to the flame and also from spreading into the lowtemperature areas in the main chamber which are not covered with flame throughout the combustion period. This leads to the SOF reduction. (b) The hard spray-wall impinging at the beginning of the injection period is prevented, so that the soot formation near the spray-wall impinging point is prevented at the beginning of the combustion period. Therefore, flowout of the dense soot occurs sufficiently after the flowout of the flame which raises the temperature of the main chamber. Thus, the soot cloud is easily oxidized in the main chamber. (c) The quantity of the formed dense soot itself is reduced due to the decrease of the main injection period in which the hard spray-wall impingement takes place. Therefore, exhaust smoke is reduced. (2) Increased injection pressure by using high nozzle opening pressure increases exhaust smoke. This is because a great deal of dense soot is formed by the hard spray-wall impinging due to the increased fuel spray momentum based on the characteristics of throttle nozzle. Therefore, not increasing but decreasing the injection pressure is effective for smoke reduction as long as a conventional throttle type injection nozzle is used. (3) Increasing swirl chamber volume ratio decreases exhaust smoke and PM especially over medium load con-
Y. Hotta et aL / J S A E Review 18 (1997) 19-31
dition. This is caused not only by the promotion of soot oxidation due to the decreased equivalence ratio in the swirl chamber, but also by the decrease of the formed dense soot near the spray-wall impinging point. In addition, NO x reduction is also achieved when increasing the EGR rate, which is possible due to the great reduction of smoke in this case. (4) Over medium load conditions, the main cause of the exhaust smoke is both the fuel adhesion on the chamber wall, and the stagnation of rich fuel-air mixture. Therefore, for example, the prevention of the spray-wall impingement by destroying the fuel spray core using a spray impinging plate is effective to reduce exhaust smoke. (5) The methods for simultaneous reduction of NO x and PM clarified through this study are summarized in Fig. 30. •
EGR
Injection Rate
ction I
Increase of [ Swid Chamber ~ VoumeRatio I
[
Decrease of ~ E ~ O F [ inSwirlChamber I
~eu~t~c~ioun I .......... =
~l Reduction I '
References [1] Yoshida, T. et al., Development of a Low Exhaust Emission and High Performance Turbocharged Diesel Engine (in Japanese with English summary), Journal of JSAE, 47-10 pp. 10-16 (1993). [2] Shinzawa, M. et al., Effects of Combustion Improvement on Reducing NOx Emissions of IDI Diesel Engines (in Japanese with English summary), Proceedings of the 1 lth Internal Combust. Engine Symp., Japan, pp. 357-362 (1993). [3] Walker, A. et al., High Performance Diesel Catalysts for Europe Beyond 1996, SAE Paper 950750 (1995). [4] Kawamura, K. et al., Analysis of Fuel Spray Characteristics under High Pressure Injection (in Japanese with English summary), Proceedings of JSAE, No. 911076 (1991). [5] Nakakita, K. et al., Optimization of Pilot Injection Pattern and Its Effect on Diesel Combustion with High-Pressure Injection, JSME International Journal, Series B, Vol. 37, No. 4 (1994). [6] Streeter, V.L. and Wylie, E.B., Hydraulic Transients, McGraw-Hill, New York (1967). [7] Nakakita, K. et al., A Study on Diesel Combustion with High-Pressure Fuel Injection (Improvement of Combustion and Exhaust Emissions Using Small-Hole-Diameter Nozzles) (in Japanese with English summary), Trans. Jpn. Soc. Mech. Eng., Series B, Vol. 60, No. 577, pp. 3198-3206 (1994).
'
Fig. 30. Mechanism of NO x and PM reduction.
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