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Influence of EGR and oxygen-enriched air on diesel engine NO–Smoke emission and combustion characteristic
Applied Energy 107 (2013) 304–314
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Applied Energy journal homepage: www.elsevier.com/locate/apenergy
Influence of EGR and oxygen-enriched air on diesel engine NO–Smoke emission and combustion characteristic Wei Zhang a, Zhaohui Chen a, Weidong Li a, Gequn Shu b,⇑, Biao Xu b, Yinggang Shen a a b
Yunnan Key Laboratory of Internal Combustion Engine, Kunming University of Science and Technology, Kunming 650500, China State Key Laboratory of Engines, Tianjin University, Tianjin 300072, China
h i g h l i g h t s " Oxygen-enriched intake air and proper ratio of low-temperature cold EGR were combined to achieve low NO–Smoke emission. " A combined n-heptane model containing a detailed PAHs mechanism was developed. " The mechanism explains diameter growth of soot particles effectively suppressed by oxygen-enriched combustion. " The mechanism coupled with CFD model well predict NO and Smoke emissions.
a r t i c l e
i n f o
Article history: Received 23 June 2012 Received in revised form 29 January 2013 Accepted 6 February 2013 Available online 16 March 2013 Keywords: Diesel engine EGR Oxygen-enriched intake air Oxygen-enriched combustion Emission characteristic
a b s t r a c t The oxygen enriched combustion of diesel engines can reduce smoke emission and increase engine thermal efficiency; however it can also lead to an increase of NO emission. In this paper, experiment was conducted on a turbocharged direct injection diesel engine, and oxygen-enriched and EGR techniques were used to produce lower NO–Smoke emission than the unmodified engine under the same fuel supply rate curve and fuel supply quantity. The specific fuel consumption and the power loss were lower than 5% compared to the unmodified engine. The effect of oxygen enrichment on the particle size distribution was tested and analyzed. The results revealed that the optimal NO–Smoke emission can be achieved at these conditions: 1600 rpm of engine speed, full load, 30–40% EGR rate and 21.5–22.5% of intake oxygen density; 2200 rpm of engine speed, full load, 20–45% EGR rate and 22–24% of intake oxygen density. The result of particle size distribution tests revealed that oxygen enriched combustion can effectively suppress the diameter growth of particles and lead to fewer large particles with a diameter larger than 100 nm emissions; however it did lead to an increase of 15 nm small particles. A reduced n-heptane kinetic model was also developed in this research which contained NO and PAHs formation mechanisms, and the model was coupled with a CFD model to simulate the oxygen-enriched combustion of a diesel engine. The calculated results demonstrated that the coupled model can accurately predict ignition time and the change of in-cylinder pressure when the combined oxygen-enriched and EGR technique was used. The computed NO change with in-cylinder oxygen density agreed well with experiment results, and the computed result of the growth experience of PAHs showed that oxygen-enriched combustion can effectively suppress HACA reaction during PAHs formation, which leads to the reduction of large molecule PAHs, and this result agreed well with the observed situation that particle size diameter decreases with the increase of intake oxygen density. Ó 2013 Elsevier Ltd. All rights reserved.
1. Introduction The intake oxygen-enriched combustion research of an internal combustion engine began in the 1960s [1], with an original purpose to reduce emissions. The studies found that intake oxygen-enriched combustion will not only improve the emission characteristics, but also significantly reduce particulate emissions, ⇑ Corresponding author. Tel.: +86 22 27409558. E-mail address: [email protected] (G. Shu). 0306-2619/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.apenergy.2013.02.024
while the smoke capacity can significantly decrease 35–73% through the increase of intake oxygen concentration [2]. Researches also showed that intake oxygen-enriched combustion can increase the heat of combustion efficiency at some loads, and the total fuel consumption is reduced by 2–10% [3]. Engine power density can also be increased by applied intake oxygen-enriched combustion, which means the engine can be downsized with the same power output. The intake of oxygen concentration increased to 32% at same power output, which means the displacement of the engine can be reduced to 73% [4], to shorten the ignition delay
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Nomenclature V InCO2 V ExCO2 Mcycle Hfuel R Ncylinder Nnozzle
CO2 volume fraction in the intake gas CO2 volume fraction in the exhaust gas fuel supply quantity of each hole per cycle (kg/cycle) fuel consumption per hour (kg/h) engine speed (r/m) the number of cylinder the number of holes in each nozzle
Abbreviations EGR exhaust gas recirculation NO nitrogen oxide PM particulate matter CFD computational fluid dynamic
period and allow the use of low-quality fuel [3]. The increase in NOx emissions is a key issue for further development and application of constraints to internal combustion engines of oxygenenriched combustion technology. A large number of studies have shown that oxygen-enriched combustion for reducing particulate matter and soot is very significant, but the case did not take special precautionary measures, as NOx emissions will significantly increase [5–8], however, the emission characteristics of engine fueled with oxygenated fuel oxygen-enriched combustion is similar to that of oxygen-enriched intake combustion [9,10]. The diesel engine harmful emissions of NOx and PM purification methods contradict each other, thus making it difficult to fully meet the emissions regulations. EGR can reduce NOx emissions [11], but the EGR is too high, resulting in lower O2 concentration of the intake, combustion deterioration, and diesel engine performance degradation, increase in CO, HC and PM emissions levels. In this research, experimental and numerical simulation methods were combined, and appropriate EGR rate and oxygen-enriched air were used to suppress the high NO emission caused by oxygen-enriched combustion and achieved lower temperature combustion (compared to the unmodified engine’s combustion). The oxygenenriched intake air and proper ratio of low-temperature cold EGR can be combined to achieve low NO–Smoke emission and the oxygen concentration and EGR rate was delicately adjusted to achieve the goal in experimental research. In addition, the exhaust gas particle size distribution under different oxygen concentrations was detected. Polycyclic aromatic hydrocarbons (PAHs), as important soot precursors, have significant impact on smoke opacity in exhaust gas [12]. In this research, a CFD model was established using AVL FIRE software, and a new n-heptane mechanism containing the formation and oxidation mechanism of benzene, naphthalene, phenanthrene, pyrene was established. The CFD model was coupled with a chemical kinetic model to simulate oxygen-enriched combustion of a diesel engine. The PAHs evolution progress under Table B.1 Main parameters of 4100QBZL-2 diesel engine. Engine
Main parameters
Engine type Bore/stroke Compression ratio Power Displacement Injection timing Injection duration Combustion chamber Intake air type
Inline 4 cylinder, water-cooled, 4 stroke 100 mm/115 mm 17.5 80 kW/3200 rpm 3.6 L 12.5° BTDC 22° CA x bowl in piston Turbocharged, inter-cooler
PAHs HACA BSFC O2 CO2 rpm CA BTDC FSN ppm IVC EVO
polycyclic aromatic hydrocarbons hydrogen abstraction C2H2 addition reaction brake specific fuel consumption oxygen carbon dioxide revolutions per minute crank angle before top dead center filter smoke number part per million intake valve close exhaust valve open
oxygen enriched conditions was analyzed in this research. Experimental and numerical simulation methods were combined to explore the effects of intake oxygen-enriched combustion on emissions and combustion characteristics of a diesel engine. 2. Experimental setup and test method 2.1. Test engine and experiment setup Experiment was conducted on an inline four cylinder watercooled turbocharged 4100QBZL-2 diesel engine. The main parameters of the engine are shown in Table B.1. The exhaust gas was measured by a DIGAS 4000 exhaust gas analyzer from AVL. Smoke opacity was measured by a AVL 415S smoke meter. Power output was measured by a electric dynamometer produced by AVL, and a WaveBook data acquisition system was used to collect data. DMS500 fast particle analyzer produced by Combustion Company was used to measure the exhaust gas particle size distribution under different intake oxygen densities. The schematic of experimental setup is shown in Fig. C.1. The oxygen used in the experiment was supplied from lowtemperature adiabatic-storage liquid oxygen. Liquid oxygen was gasified though a liquid gas vaporizer. The pressure of the oxygen was very high, but reduced to match the pressure of the after compressor by a pressure-reducing valve. Then the oxygen mixed well with the air from the compressor and cooled EGR exhaust gas in a intake air buffer/mixing tank as intake gas. The experiment used a low-pressure EGR, and exhaust gas through the turbine into the exhaust buffer tank. Exhaust not containing EGR will emit to atmosphere through exhaust gas valve, while the other part of exhaust gas will enter the combustion chamber through the EGR valve. Exhaust gas flowed to the EGR cooler through the EGR valve, and the temperature will drop to less than 100 °C. The cooled EGR gas will benefit from reducing the combustion temperature and reduced NO emission. The cooled exhaust gas went into the intake buffer tank to mix well with oxygen and air. EGR rate can be controlled by changing the EGR valve and exhaust gas valves’ opening. The EGR rate can be calculated by Eq. (A.1).
Table B.2 Test condition. Engine speed (rpm) Fuel consumption per hour (kg/h) EGR rate (%) O2 volume fraction (%) Fuel injection strategy
1600 10 0, 21, 32, 48, 50 21, 22, 23, 24 Single injection
2200 14.2 0, 21, 32, 48, 50 21, 22, 23, 24 Single injection
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W. Zhang et al. / Applied Energy 107 (2013) 304–314 Table B.3 The formation mechanism of the benzene ring in SKLE + PAHs model. NO.
Reactions
1 2 3 4 5 6
C3H3 + C3H3 = A1 nC4H3 + C2H2 = A1 nC4H5 + C2H2 = A1 + H lC6H4 + H = A1 nC6H5 = A1 lC6H6 + H = A1 + H
Table B.4 The PAHs molecular structure of the one ring benzene to four rings Pyrene. Code
A1
A2
A3
A4
Name Molecular Structure
Benzene C6H6
Naphthalene C10H8
Phenanthrene C14H10
Pyrene C16H10
Table B.5 Main parameters of simulation. Speed
1600 rpm
1600 rpm
1600 rpm
Oxygen concentration EGR rate Pressure Temperature Mcycle N2 mole fraction O2 mole fraction CO2 mole fraction
21% 0% 0.22 MPa 145 °C 1.25E–5 kg 0.79 0.21 0
22% 32% 0.22 MPa 145 °C 1.25E–5 kg 0.738 0.22 0.042
24% 0% 0.22 MPa 145 °C 1.25E–5 kg 0.76 0.24 0
emission area is on the left side of the dash line, while low-carbon smoke emission area is on the right side of the dash line. As seen in Fig. C.4, the trend of the smoke emission changes with EGR rate at 2200 rpm is similar with the trend at 1600 rpm. Comparing Figs. C.3 and C.5, it is observed that the low carbon area at 2200 rpm is smaller than that at 1600 rpm. The increased smoke emission caused by using EGR can be effectively suppressed by the combination of intake oxygen-enriched intake and EGR enabling low temperature combustion under high oxygen concentration. Fig. C.6 shows NO emission change with EGR rate and intake oxygen concentration. At different EGR rates, NO emission rises rapidly with increasing oxygen concentration in the intake gas. The NO emission reduced with increasing EGR rate. With a 50% EGR rate and 24% intake oxygen concentration, combustion will deteriorate, which lead to a NO emission equal to the NO emission when no EGR is used, and 24% concentration of O2. The dash line in Fig. C.7 represents the NO emission of the unmodified engine at 1600 rpm. The upper left corner of the dash line of oxygen concentration and EGR rate combination can lead to NO emission lower than the unmodified engine, and adversely when moving towards the right area of the dash line. When comparing Figs. C.8 and C.6, the same NO emission trend shows at 2200 rpm and 1600 rpm; that is, the NO emission increases with increased intake oxygen concentration. However, at test engine speed of 2200 rpm the NO emission was lower than
2.2. Experiment method In the study, the engine was operated at a constant speed of 1600 rpm and 2200 rpm, 100% fuel throttle and full load condition. The fuel consumption and fuel supply rate curve were stable. During the tests, the dynamometer was adjusted to the full load at the setting speed, and real-time monitoring of the CO2 and O2 volume fractions in the intake gas through the AVL gas analyzer. Control of CO2 and O2 volume fractions in the intake gas achieved the concentration of test requirements through adjustment of the EGR and O2 throttle valves. When the engine operation was stable (which shows by speed and loads), data collection instruments were used to collect data. To compare the result, 1600 rpm and 2200 rpm of speed, full load, EGR rates were set at 0%, 21%, 32%, 48% and 50%. At each EGR rate, 21%, 22%, 23%, 24% O2 concentrations were tested. The test condition is shown in Table B.2. 3. Results and discussion 3.1. The effect of EGR and oxygen-enriched on smoke and NO emissions It can be seen from Fig. C.2, the smoke showed a rapid downward trend with the increase of O2 concentration in the intake gas. The trend was more obvious when the EGR rate is high. When the oxygen concentration increased to 24%, soot emission is much lower than that of original engine in the exhaust under each EGR rate. It can be seen from Fig. C.3, smoke emission decreased rapidly with the increase of O2 concentration. The dash line in Fig. C.3 represents the smoke emission of the unmodified engine (which is air for combustion and no EGR involved). High carbon smoke
Fig. C.1. Layout of experimental setup. 1. EGR cooler, 2. EGR temperature sensor, 3. EGR pressure sensor, 4. inter-cooler, 5. intake air buffer/mixing tank, 6. throttle valve, 7. pressure reducing valve, 8. liquid oxygen vaporizer, 9. liquid oxygen storage tank, 10. diesel, 11. control PC, 12. oil consumption meter, 13. data acquisition instrument, 14. diesel engine, 15. dynamometer, 16. compressor, 17. combustion DMS500, 18. exhaust pipe, 19. AVL gas analyzer, 20. exhaust pressure valve, 21. EGR valve, 22. AVL 415S smoke meter, 23. turbine, 24. over flow valve, 25. air flow throttle valve, 26. optical shaft encoder, 27. intake air temperature sensor, 28. intake air flow meter, 29. intake air pressure sensor, 30. AVL4000 gas analyzer and 31. oxygen sensor.
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4.0
3.2 2.8
2.0 1.8
2.5
0.5
2.4 Original Smoke emission 2.0 1.6
3.0
40
EGR Rate
3.6
Smoke / FSN
50
0% EGR 21% EGR 32% EGR 48% EGR 50% EGR
1.0
30
1.5
SPEED=2200 rpm LOAD=100% Smoke / FSN
20
SPEED=1600 r/m LOAD=100%
1.2
10 0.5
0.8 0.4 21.0
21.5
22.0
22.5
23.0
23.5
0 21.0
24.0
21.5
O2 Concentration / % Fig. C.2. Influence of O2 concentration and EGR rate on smoke emission at 1600 rpm, 100% load.
3.0 1.5
EGR Rate
40
1.1 1.8
30
20 2.6
SPEED=1600 rpm LOAD=100% Smoke / FSN
2.2
10 2.7 0 21.0
21.5
22.0
22.5
23.0
22.5
23.0
23.5
24.0
23.5
Fig. C.5. Influence of O2 concentration and EGR rate on smoke emission at 2200 rpm, 100% load (contour map).
engine speed of 1600 rpm, with the same EGR rate and intake oxygen concentration. This is because the reaction time of NO formation is reduced due to increase of engine speed. Therefore, the different combinations of EGR rate and intake oxygen concentration are needed to control NO emissions under the different engine speeds. In Fig. C.9, the dash line is the contour map of NO emission at 2200 r/min of speed and full load of the unmodified engine emission. The upper left of the figure is O2 concentration-EGR rate combination of low NO emission area, the bottom right corner of the figure is the high NO emission area.
50 2.6
22.0
O2 Concentration / %
3.2. The effects of EGR and enriched oxygen on in-cylinder pressure and heat release rate
24.0
O2 Concentration / % Fig. C.3. Influence of O2 concentration and EGR rate on smoke emission at 1600 rpm, 100% load (contour map).
Regarding the O2 concentration and EGR rate distribution, we selectively chose the 2200 rpm and 1600 rpm of speed and different oxygen concentrations and EGR rates to draw the in-cylinder pressures and calculated heat release curves to compare and analyze.
2500
0% EGR
SPEED=2200 rpm LOAD=100%
3.5 3.0
Smoke / FSN
SPEED=1600 rpm LOAD=100%
32% EGR
0% EGR 21% EGR 32% EGR 48% EGR 50% EGR
2.5
48% EGR 50% EGR
1500
NO / ppm
4.0
21% EGR
2000
2.0 1.5
1000
500
Original NO emission
1.0 0.5 0.0
Original Smoke emission
21.0
21.5
22.0
0
22.5
23.0
23.5
24.0
O2 Concentration / % Fig. C.4. Influence of O2 concentration and EGR rate on smoke emission at 2200 rpm, 100% load.
21.0
21.5
22.0
22.5
23.0
23.5
24.0
O2 Concentration / % Fig. C.6. Influence of O2 concentration and EGR rate on NO emission at 1600 rpm, 100% load.
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50
12.0M
240
10.0M
200
1275
EGR rate
30
20
Cylinder pressure / Pa
SPEED=1600 rpm LOAD=100% NO emission / ppm 790
1275 994
1838
8.0M
O2:21% EGR:0% O2:21% EGR:48%
6.0M 4.0M
SPEED=1600 rpm LOAD=100%
0 21.0
21.5
22.0
22.5
O2:22% EGR:32%
2.0M
23.5
40
O2:23% EGR:48%
2119
23.0
120 80
10 1556
160
Heat release / (J/deg)
431
40
0.0 -10
-5
0
10
15
20
0 25
Crank angle / º CA
24.0
O2 Concentration / %
5
Fig. C.10. Influence of O2 concentration and EGR rate on cylinder pressure and heat release rate at 1600 rpm, 100% load.
Fig. C.7. Influence of O2 Concentration and EGR rate on NO.
2000
0% EGR 21% EGR 32% EGR 48% EGR 50% EGR
1800 1600
NO / ppm
1400
SPEED=2200 rpm LOAD=100%
1200 1000 800 600 400
Original NO emission
200 21.0
21.5
22.0
22.5
23.0
23.5
24.0
O2 Concentration / % Fig. C.8. Influence of O2 concentration and EGR rate on NO emission at 2200 rpm, 100% load.
Four test conditions, 21% O2 and 0% EGR rate, 21% O2 and 48% EGR rate, 22% O2 and 32% EGR rate, 23% O2 and 48% EGR rate, were chosen to be compared and analyzed. The results are shown in Fig. C.10. As Fig. C.10 shows, when the EGR rate is 48%, the in-cylinder pressure is lower and the ignition delay time is longer due to more CO2 getting into each cylinder involved in combustion. In the case of 22% O2 and 32% EGR rate, there is no significant difference shown in the in-cylinder pressure and heat release curve compared with the unmodified engine, which indicated that although the oxygen concentration increased 1%, nearly the same in-cylinder pressure and heat release can be reached by using 32% EGR. Four conditions are chosen to be compared and analyzed: 21% O2 and 0% EGR, 22% O2 and 21% EGR, 23% O2 and 32% EGR, 24% O2 and 50% EGR. The in-cylinder pressure and heat release of these four conditions are shown in Fig. C.11. As Fig. C.11 shows, in-cylinder pressure of 22% O2 and 20% EGR is very close to that of the unmodified engine, and the heat release delayed a little. In-cylinder pressure in the case of 23% O2 and 32% EGR decreased noticeably and ignition time was further delayed. In the case of 24% O2 and 50% EGR, ignition time was delayed by 4° CA, the maximum pressure was reduced by 2 MPa, and the combustion deteriorated. 3.3. The effects of EGR and enriched oxygen on brake specific fuel consumption and power Fig. C.12 shows the effects of oxygen concentration and EGR rate on brake specific fuel consumption. The effect of oxygen con-
50 276
12.0M
30
Cylinder pressure / Pa
SPEED=2200 rpm LOAD=100% NO / ppm
20
10
513 775
0 21.0
985 21.5
22.0
1221 22.5
SPEED=2200 rpm LOAD=100%
280
10.0M
240 200
8.0M
160 6.0M 4.0M
O2 :21% EGR:0%
120
O2 :22% EGR:21%
80
2.0M
O2 :23% EGR:32%
1458 23.0
23.5
24.0
O2 Concentration / % Fig. C.9. Influence of O2 concentration and EGR rate on NO emission at 2200 rpm, 100% load (contour map).
0.0 -10
O2 :24% EGR:50% -5
0
5
10
15
20
Heat release / (J/deg)
EGR rate
40
40 0 25
Crank angle / º CA Fig. C.11. Influence of O2 concentration and EGR rate on cylinder pressure and heat release rate at 2200 rpm, 100% load.
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BSFC / ( g. (kw. h) -1)
200
160
120
SPEED=1600 rpm LOAD=100%
80
0% EGR 21% EGR 32% EGR 48% EGR
40
50% EGR 0
21
22
23
24
O2 Concentration / % Fig. C.12. Influence of O2 concentration and EGR rate on brake specific fuel consumption at 1600 rpm, 100% load.
52 50
Power / KW
0% EGR 21% EGR
48
32% EGR 48% EGR
46
centration on brake specific fuel consumption is not obvious, however, the specific fuel consumption increase significantly when EGR is more than 32%. Fig. C.13 shows the influence of oxygen concentration and EGR rate on power, similar with brake specific fuel consumption, the intake oxygen concentration has little influence on power, and power reduces with increasing of EGR. When EGR rate reached 32%, the power reduces a little (<3%), when EGR rate reached 50%, the decline reached about 8 kW (>15%) comparing with 0% EGR. As Fig. C.14 shows, at 1600 rpm of engine speed, full load, the power of original engine is 50.7 kW. The upper limit of power loss was set as 5% (48 kW), which was shown as a dash line at the contour map of EGR, intake oxygen concentration and power output. The area under dash line in the figure is the area of power loss less than 5% and above the dash line area the power loss is more than 5%. It can be observed that the power loss under different oxygen concentration can be controlled within 5% when EGR rate was controlled within 40%. Fig. C.15 shows the brake specific fuel consumption at 2200 rpm. Similar to 1600 rpm, the brake specific fuel consumption under the same EGR rate and different intake oxygen concentration has little difference, and the brake specific fuel consumption increases significantly when the EGR rate is greater than 32%. Fig. C.16 shows the power output at different oxygen concentration and EGR rates at 2200 rpm. The power output is reduced 5% when the EGR rate is 48%, and the power loss increases to 6 kW (>9%) when 50% EGR. The dash line 61 kW in Fig. C.17 was the upper limit of 5% power loss when engine at 2200 rpm, full load. Power loss can be controlled within 5% under various oxygen concentrations with 45% EGR.
50% EGR
3.4. The NO–Smoke emission area at different intake oxygen concentration and EGR rates
44
SPEED=1600 r/m LOAD=100%
42 40 21.0
21.5
22.0
22.5
23.0
23.5
24.0
O2 Concentration / % Fig. C.13. Influence of O2 concentration and EGR rate on power at 1600 rpm, 100% load.
Concerning the smoke emission, NO emission, brake specific fuel consumption, and power at 1600 rpm and 2200 rpm, the oxygen concentration from 21% to 24% and the EGR rate from 0% to 50% can be matched to realize lower Smoke–NO emissions than the unmodified engine with loss of brake specific fuel consumption and power no more than 5%. The contour map of Smoke emission and NO emission at 1600 rpm and 2200 rpm, with the 5% power loss line, results in a NO–Smoke emission and power loss line contour map as shown
50
240
45 46
200
40
BSFC / ( g. (kw. h) -1)
48 49
EGR rate
50 30
51
20
SPEED=1600 rpm LOAD=100% Power / kW
10
0 21.0
160
120
SPEED=2200 rpm LOAD=100%
0% EGR 21% EGR
80
32% EGR 48% EGR
40
50% EGR
0 21.5
22.0
22.5
23.0
23.5
24.0
O2 Concentration / % Fig. C.14. Influence of O2 concentration and EGR rate on power of 1600 rpm, 100% load (contour map).
21
22
23
24
O2 Concentration / % Fig. C.15. Influence of O2 concentration and EGR rate on brake specific fuel consumption at 2200 rpm, 100% load.
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64
Power / kW
60 56 0% EGR
SPEED=2200 rpm LOAD=100%
52
21% EGR 32% EGR 48% EGR
48
50% EGR
44 40 21.0
21.5
22.0
22.5
23.0
23.5
24.0
O2 Concentration / % Fig. C.16. Influence of O2 concentration and EGR rate on power at 2200 rpm, 100% load.
50
4. The construction of numerical models and the analysis of simulation results
61 62
EGR Rate
40
63
4.1. The combined model containing detailed PAHs and NOx formation mechanisms
30 64 20
10
0 21.0
SPEED=2200 rpm LOAD=100% Power / kW
65
21.5
22.0
22.5
23.0
23.5
24.0
O2 Concentration / % Fig. C.17. Influence of O2 concentration and EGR rate on power of 2200 rpm, 100% load (contour map).
(a) 1600rpm
in Fig. C.18. The white area in the figure is lower NO–Smoke emission area. Lower NO–Smoke emission than the unmodified engine with no significant power loss in this area can be attained by matching EGR and O2 in this area. In Fig. C.18a, the white area represents 1600 rpm, full load, 30–40% EGR rate, 21.5–22.5% O2. Matching the EGR rate and oxygen concentration in this area can achieve lower NO–Smoke emission than the unmodified engine. Similarly, the white area in Fig. C.18b is the lower NO–Smoke emission area of 2200 rpm, full load, 20–45% EGR rate, 22–24% O2. The matches of EGR rate and oxygen concentration can achieve lower NO–Smoke emission. When comparing Fig. C.18a and b, the size and shape of lower NO–Smoke emission areas are different in the engine’s various work conditions and produces different experiment results. The changes trends of combustion and emissions with oxygen concentration and EGR at 1600 rpm were similar with the case of those at 2200 rpm. In the next section, the simulation researches were conducted to study the effects of oxygen-enriched combustion and EGR on NO formation and PAH growth experience at 1600 rpm and full load.
The composition of diesel fuel is complex and difficult to accurately simulate. n-Heptane is usually used to simulate the combustion and emission characteristics, because its cetane number is close to diesel fuel [13]. The detailed chemical model developed by the Lawrence Livermore National Laboratory (LLNL) contains 631 species and 2827 reactions, and was validated by shock tube and rapid compression machine experiments, and was proved to be reliable [14]. This research used a reduced skeleton model (SKLE model) based on the LLNL detailed model, and contains 41 species and 63 reactions [15]. The new SKLE + PAHs detail model contains the detailed PAHs mechanism proposed by Wang containing 99 species and 527 reactions [16]. The detailed PAHs mechanism of Wang was validated by premixed laminar flame [17,18].
(b) 2200rpm
Fig. C.18. NO–Smoke emission zone distribution at 1600 rpm and 2200 rpm, 100% load. (1) Red Zone: Power Loss zone, (2) Green Zone: High Smoke Emission Zone, (3) Blue Zone: High Smoke and NO Emission Zone, (4) Cyan Zone: High NO Emission Zone and (5) White Zone: Low Smoke and NO Emission Zone. (For interpretation of the references to colour in this figure legend, the reader is referred to the web version of this article.)
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Fig. C.19. The growth of PAHs with HACA reactions.
Fuel Injection Rate
0.0012
0.0008
0.0004
0.0000 0
4
8
12
16
20
Injection Duration / deg Fig. C.20. Spray curve and combustion chamber 1/6 mesh.
The NOx formation mechanism in this research is the Zeldovich NOx mechanism validated by Dr. Golovitchev from Chalmers technology university, and contains 4 species and 9 reactions [19]. The acetylene formation and oxidation reactions and the C2Hx ? C6Hx formation and oxidation reactions added with benzene and phenyl formation reactions to form a complete description of the benzene rings formation. The reactions related to the formation of benzene and benzene radical are list in Table B.3. Once the formation of benzene ring, naphthalene, phenanthrene, pyrene can form through hydrogen abstraction C2H2 addition reactions (HACA) [17]. The molecular structures of benzene, naphthalene, phenanthrene, pyrene are shown in Table B.4. The formation paths from benzene to pyrene are shown in Fig. C.19. Through the combination of ethylene reaction in the SKLE model, acetylene and propargyl precursor reactions, ethylene decomposition follow-up reactions, and the formation and oxidation of PAHs in Wang’s model, so n-heptane SKLE model combines well with Wang’s PAHs model. When the NOx formation model joined the combined model, the new SKLE + PAHs model contains 101 components, 329 elementary reactions, reflecting PAHs and NOx changes in the n-heptane combustion process, and can be applied to coupling calculation with the CFD model. 4.2. The establishment of the CFD model and set up of initial condition In this research, ESE Diesel in FIRE software was used to establish the combustion chamber model. Construction parameters of
the test engine were set in the general engine parameter and piston movement specification. Combustion chamber type was chosen and modified according to the geometric data of the piston in Sketcher, as were the fuel injector position, the number of injector holes, the spray angle and other parameters. The 2D grids of the combustion chamber were created in Mesh, and then generated 3D solid models based on the 2D grid. Due to the long calculation time of the full combustion chamber grid, the 1/6 model shown at Fig. C.22 was established according to the injector holes in the nozzle to save the computing time. In the initial condition set of the coupling calculation, only time scale between intake valve closing (IVC) to exhaust valve opening (EVO) was calculated. The top dead center (TDC) was set as 0° CA, and the calculated crank angle range of 132–115° CA, total 247° CA, which was determined by IVC and EVO timing. The fuel injection time was set as 12.5° CA before TDC. The fuel supply curve shown in Fig. C.20 was maintained during the calculation. The fuel supply quantity was measured by a fuel consumption meter and the fuel each cycle supplied of the 1/6 model can be calculated by Eq. (A.2). The initial turbulent kinetic energy was calculated according to the engine speed and the intake air flow. In AVL Fire’s solver steering file, the k–epsilon model was selected as the gas turbulent model, the turbulent diffusion model was activated in spray module, and the Dukowicz fuel evaporation model and the Wave fuel broken model should be involved in numerical simulation at the same time.
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2500
25.0M
Mean pressure / Pa
20.0M
Mean temperature / K
Detail model SKLE model SKLE+PAHs model
15.0M
10.0M
Detail model SKLE model SKLE+PAHs model
2000
1500
1000
5.0M 500 0.0 -100
-50
0
50
-100
100
-50
Crack angle / º CA
0
50
100
Crack angle / º CA
(a) pressure
(b) temperature
12.0M
SPEED=1600 rpm LOAD=100%
10.0M
O2 :21% EGR:0%
0.005
NO mass fraction
Cylinder pressure / Pa
Fig. C.21. The simulation result comparison with detailed model, SKLE model and SKLE + PAHs model.
Experimental Calculated
8.0M 6.0M
SPEED=1600 rpm LOAD=100%
0.004
0.003
0.002
Calculated O2 :21% EGR:0%
4.0M
O2 :22% EGR:32%
0.001
O2 :24% EGR:0%
2.0M 0.000 0
0.0 -100
-50
0
50
10
100
20
30
40
Crank angle / º CA
Crank angle / º CA Fig. C.22. Experimental and calculated cylinder pressure on 1600 rpm, 100% load.
Fig. C.24. Influence of O2 concentration and EGR rate on NO mass fraction at 1600 rpm, 100% load.
12.0M
1800
10.0M
1600 8.0M 1400 Calculated 1200
6.0M
O2 :21% EGR:0% O2 :22% EGR:32% O2 :24% EGR:0%
1000
4.0M -20
-10
0
10
20
30
40
Crank angle / º CA
Size spectral density / (dN/ dl ogDp/ cc)
2000
5x108
SPEED=1600 rpm LOAD=100%
Cylinder pressure / Pa
Cylinder temperature / K
2200
SPEED=1600 rpm LOAD=100%
4x108
3x10
Experimental O2 :21% EGR: 0%
8
O2 :22% EGR: 0% O2 :23% EGR: 0%
2x108
O2 :24% EGR: 0%
1x108
0 10
Fig. C.23. Influence of O2 concentration and EGR rate on pressure and temperature at 1600 rpm, 100%.
100
1000
Dp / nm Fig. C.25. Influence of O2 concentration on particle size distribution (experimental).
4.3. Validation of the chemical kinetic model and the coupled CFD calculation With the chemical kinetic analysis software CHEMKIN, developed by Sandia National Laboratory, a calculated result comparison
was carried out on three models: the specifically LLNL n-heptane detail model, reduced SKLE model, and SKLE + PAHs detail model, to verify the accuracy of the new built SKLE + PAHs detail model
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4.0x10-6
SPEED=1600 rpm LOAD=100%
1.2x10
SPEED=1600 rpm LOAD=100%
1.0x10-4
Calculated O2:21% EGR:0% O2:22% EGR:32% O2:24% EGR:0%
8.0x10-5 6.0x10-5
Calculated O2 :21% EGR:0% O2 :22% EGR:32% O2 :24% EGR:0%
3.0x10-6
A2 Mass fraction
A1 Mass fraction
-4
4.0x10-5
2.0x10-6
1.0x10-6
-5
2.0x10
0.0 0
5
10
15
25
30
0.0
35
-5
5
10
15
20
25
Crank angle / º CA
(a) A1 Mass fraction
(b) A2 Mass fraction
35
SPEED=1600 rpm LOAD=100% 2.0x10-6
Calculated O2 :21% EGR:0% O2 :22% EGR:32% O2 :24% EGR:0%
6.0x10-6
30
2.5x10-6
SPEED=1600 rpm LOAD=100%
8.0x10
4.0x10-6
2.0x10-6
Calculated O2 :21% EGR:0% O2 :22% EGR:32% O2 :24% EGR:0%
1.5x10-6
1.0x10-6
5.0x10-7
0.0 -5
0
Crank angle / º CA
-6
A3 Mass fraction
20
A4 Mass fraction
-5
0
5
10
15
20
25
30
35
0.0 -5
0
5
10
15
20
25
Crank angle / º CA
Crank angle / º CA
(c) A3 Mass fraction
(d) A4 Mass fraction
30
35
Fig. C.26. Influence of different O2 concentration on mean PAHs mass fraction.
in the research. A zero-dimension single region IC engine quick compressor model was applied in the chemistry kinetic calculation, with the boundary and initial conditions the same as the measured value of the diesel engine at 2000 rpm and full load. As can be seen from the chemical kinetic calculation result in Fig. C.21, difference exists in the SKLE model and detail model at low-temperature combustion (<800 K) and high-temperature (>1000 K), but can fulfill the calculation demands as a whole. With the SKLE + PAHs model and SKLE model added, the peak values of the pressure and temperature curve have a similar drop, but the shapes of the curve agree with the detail model totally thus the combustion state can be reflected in the cylinder. The comparison between experimental results at full load, 1600 rpm and the simulation results of the CFD model coupled with the chemical kinetic model are shown in Fig. C.22. The simulation results agree well with the experimental results indicating that the calculated result using the SKLE + PAHs detail model and coupled with CFD model in FIRE software can fulfill the real measured simulation demand.
It can be seen from Fig. C.23, the calculated in-cylinder pressure and temperature in the case of 22% O2 and 32% EGR rate are lower than those at the 21% O2 and 0% EGR rate. The in-cylinder pressure and temperature are higher than those at the 24% O2 and 0% EGR rate. Comparing Figs. C.23 and C.10, it can be seen that the calculated in-cylinder pressure agrees well with the experiment results, indicating that the model simulation can reflect the temperature and in-cylinder pressure changes when O2 and CO2 concentration change. As shown in Fig. C.24, the mass fraction of NO at 22% O2, 32% EGR is lower than that at the 21% O2, 0% EGR rate. The NO emission of 22% O2 can be lower than the unmodified engine when 32% EGR is used, however, the NO mass fraction in the case of 24% O2, with no EGR involved, is almost double that of the NO mass fraction at 21% O2, 0% EGR. The simulation result of NO emission is identical with the experimental results in Fig. C.6.
4.4. Simulation results
PAHs are the main precursor of soot, they play a key role in the amount of soot [20], soot particles aggregation and growth [21,22], therefore the probabilistic prediction of soot formation by PAHs via the corresponding relationship between generation amount of PAHs and soot is possible [23,24]. So numerical simulation of PAHs combined with soot experimental measurements were used to
Three conditions were chosen to simulate: 21% O2 and 0% EGR, 22% O2 and 32% EGR, 24% O2 and 0% EGR, with 1600 rpm of speed and full load. The calculation results were compared and analyzed. The main parameters of simulation are shown in Table B.5.
4.5. The effect of intake oxygen concentration on particle size distribution in exhaust gas
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study the effect of intake oxygen concentration on soot particle size distribution in the exhaust. Fig. C.25 shows the effects of intake oxygen concentration on particle number and size distribution in exhaust gas. Particle size distribution ranges from 5 to 1000 nm. When intake oxygen concentration is 21%, most particles concentrate at about 100 nm. The large particle number reduces and small particles increase at increased intake oxygen concentration. When oxygen concentration increases to 24%, most particle diameters reduce to about 15 nm. The results indicated that oxygen-enriched combustion has significant suppression effect on diameter growth of particles. Fig. C.26 shows the mass fractions of A1–A4 (benzene–pyrene) change with the crank angle, at 22% of O2 concentration and 32% of EGR rate. The formation and oxidation of PAHs were delayed with the extension of ignition delay time. The PAHs formation at 24% O2 and no EGR are obviously lower than at 21% O2, no EGR and 22% O2, 32% EGR. These results indicated that with an increase of O2 concentration in the chamber, the formation of PAHs was significantly suppressed by the oxygen-enriched environment. 5. Conclusion (1) Using a high EGR rate can achieve low NO emission when oxygen-enriched combustion is applied; however, an excessively high EGR rate will lead to combustion deterioration, plus increase of brake specific fuel consumption, and higher smoke emission even at an oxygen-enriched combustion. (2) The proper combination of oxygen concentration and EGR rate can achieve low NO–Smoke emission. The size and shape of the low NO–Smoke emission area are different with different engine speeds and loads as shown in the specific experiment results. (3) A combined n-heptane model containing a detailed PAHs mechanism was developed in this research. The calculation results using the combined model, coupled with the CFD model predict well the NO and smoke emissions, and are suitable for qualitative research of oxygen-enriched combustion of diesel engines. (4) The oxygen-enriched combustion of diesel engines can reduce smoke opacity effectively and suppress the diameter growth of soot particles, while reducing the number of large soot particles; however, this leads to the increase of small particles with diameters of about 15 nm. (5) The numerical simulation results indicated that oxygenenriched combustion can suppress HACA reactions and reduce the formation of large molecule PAHs (naphthalene, phenanthrene, pyrene), while also reducing the mass fraction of benzenes at the same time.
Acknowledgements This work was supported by the National Natural Science Foundation of China (No. 50976077). The authors also grateful to Department of Education, Yunnan Province, China (KKJA201202025, KKJA201256011) and Kunming University of Science and Technology, Yunnan Province, China (KKSY201202143, KKSY201256142). Appendix A
EGR% ¼ ðV InCO2 =V ExCO2 Þ 100%
ðA:1Þ
M cycle ¼ Hfuel =ð60 ðR=2Þ Ncylinder Nnozzle Þ
ðA:2Þ
Appendix B Tables B.1–B.5. Appendix C Figs. C.1–C.26. References [1] Karim GA, Ward G. The examination of the combustion processes in a compression-ignition engine by changing the partial pressure of oxygen in the intake charge, SAE Paper 680767; 1968. [2] Byun H, Hong B, Lee B. The effect of oxygen enriched air obtained by gas separation membranes from the emission gas of diesel engines. Desalination 2006;193(1–3):73–81. [3] Poola RB, Sekar R. Reduction of NOx and particulate emissions by using oxygen-enriched combustion air in a locomotive diesel engine. J Eng Gas Turbines Power 2003;125(2):524–33. [4] Caton Jerald A. The effects of oxygen enrichment of combustion air for sparkignition engines using a thermodynamic cycle simulation. In: ASME 2005 internal combustion engine division spring technical conference; 2005. p. 135– 47. [5] Poola RB, Sekar R. Reduction of NOx and particulate emissions by using oxygen-enriched combustion air in a locomotive diesel engine. J Eng Gas Turb Power – Trans ASME 2003;125(2):524–33. [6] Song Juhun, Zello Vince, Boehman André L, Waller Francis J. Comparison of the impact of intake oxygen enrichment and fuel oxygenation on diesel combustion and emissions. Energy Fuels 2004;18(5):1282–90. [7] Subramanian KA, Ramesh A. Experimental investigation on the use of water diesel emulsion with oxygen enriched air in a DI diesel engine. SAE Paper 2001-01-0205; 2001. [8] Bedoya Iván D, Saxena Samveg, Cadavid Francisco J, Dibble Robert W, Wissink Martin. Experimental evaluation of strategies to increase the operating range of a biogas-fueled HCCI engine for power generation. Appl Energy 2012;97:618–29. [9] Liang Chen, Ji Changwei, Liu Xiaolong. Combustion and emissions performance of a DME-enriched spark-ignited methanol engine at idle condition. Appl Energy 2011;88(11):3704–11. [10] Kegl Breda. Influence of biodiesel on engine combustion and emission characteristics. Appl Energy 2011;88(5):1803–12. [11] Maiboom Alain, Tauzia Xavier, Hétet Jean-François. Experimental study of various effects of exhaust gas recirculation (EGR) on combustion and emissions of an automotive direct injection diesel engine. Energy 2008;33(1):22–34. [12] Dobbins Richard A, Fletcher Robert A, Benner Jr Bruce A, Hoeftc Stephen. Polycyclic aromatic hydrocarbons in flames, in diesel fuels, and in diesel emissions. Combustion and Flame 2006;144(4):773–81. [13] Ahmed SS, Mauss F, Zeuch T. The generation of a compact n-Heptane toluene reaction mechanism using the chemistry guided reduction (CGR) technique. J Res Phys Chem Chem Phys 2009;223(4-5):551–63. [14] https://www-pls.llnl.gov/?url=science_and_technology-chemistry-combustionn_heptane_version_3. [15] wan-hua SU, Hao-zhong Huang. Development and calibration of a reduced chemical kinetic model of a reduced chemical kinetic model of n-heptane for HCCI engine combustion. Fuel 2005(84):1029–40. [16] Wang Hai, Frenklach Michael. A detailed kinetic modeling study of aromatics formation in laminar premixed acetylene and ethylene flames. Combust Flame 1997;110(1-2):173–221. [17] Kennedy Ian M. Models of soot formation and oxidation. Prog Energy Combust Sci 1997;23(2):95–132. [18] D’Anna A, D’Alessio A, Kent J. A computational study of hydrocarbon growth and the formation of aromatics in coflowing laminar diffusion flames of ethylene. Combust Flame 2001;125(3):1196–206. [19] Golovitchev VI, Atarashiya K, Tanaka K, Yamada S, Towards universal EDCbased combustion model for compression ignited engine simulations, SAE Paper 2003-01-1849; 2003. [20] Dobbins Richard A, Fletcher Robert A, Benner Jr Bruce A, Hoeft Stephen. Polycyclic aromatic hydrocarbons in flames, in diesel fuels, and in diesel emissions. Combust Flame 2006;144(4):773–81. [21] Blacha Thomas, Domenico Massimiliano Di, Gerlinger Peter, Aigner Manfred. Soot predictions in premixed and non-premixed laminar flames using a sectional approach for PAHs and soot. Combust Flame 2012;159:181–93. [22] Appel Jörg, Bockhorn Henning, Frenklach Michael. Kinetic modeling of soot formation with detailed chemistry and physics: laminar premixed flames of C2 hydrocarbons. Combust Flame 2000;121(1-2):122–36. [23] Hwang JY, Chung SH. Growth of soot particles in counter flow diffusion flames of ethylene. Combust Flame 2001;125:752–62. [24] Richter H, Howard JB. Formation of polycyclic aromatic hydrocarbons and their growth to soot – a review of chemical reaction pathways. Prog Energy Combust Sci 2000;26:565–608.
Contents lists available at SciVerse ScienceDirect
Applied Energy journal homepage: www.elsevier.com/locate/apenergy
Influence of EGR and oxygen-enriched air on diesel engine NO–Smoke emission and combustion characteristic Wei Zhang a, Zhaohui Chen a, Weidong Li a, Gequn Shu b,⇑, Biao Xu b, Yinggang Shen a a b
Yunnan Key Laboratory of Internal Combustion Engine, Kunming University of Science and Technology, Kunming 650500, China State Key Laboratory of Engines, Tianjin University, Tianjin 300072, China
h i g h l i g h t s " Oxygen-enriched intake air and proper ratio of low-temperature cold EGR were combined to achieve low NO–Smoke emission. " A combined n-heptane model containing a detailed PAHs mechanism was developed. " The mechanism explains diameter growth of soot particles effectively suppressed by oxygen-enriched combustion. " The mechanism coupled with CFD model well predict NO and Smoke emissions.
a r t i c l e
i n f o
Article history: Received 23 June 2012 Received in revised form 29 January 2013 Accepted 6 February 2013 Available online 16 March 2013 Keywords: Diesel engine EGR Oxygen-enriched intake air Oxygen-enriched combustion Emission characteristic
a b s t r a c t The oxygen enriched combustion of diesel engines can reduce smoke emission and increase engine thermal efficiency; however it can also lead to an increase of NO emission. In this paper, experiment was conducted on a turbocharged direct injection diesel engine, and oxygen-enriched and EGR techniques were used to produce lower NO–Smoke emission than the unmodified engine under the same fuel supply rate curve and fuel supply quantity. The specific fuel consumption and the power loss were lower than 5% compared to the unmodified engine. The effect of oxygen enrichment on the particle size distribution was tested and analyzed. The results revealed that the optimal NO–Smoke emission can be achieved at these conditions: 1600 rpm of engine speed, full load, 30–40% EGR rate and 21.5–22.5% of intake oxygen density; 2200 rpm of engine speed, full load, 20–45% EGR rate and 22–24% of intake oxygen density. The result of particle size distribution tests revealed that oxygen enriched combustion can effectively suppress the diameter growth of particles and lead to fewer large particles with a diameter larger than 100 nm emissions; however it did lead to an increase of 15 nm small particles. A reduced n-heptane kinetic model was also developed in this research which contained NO and PAHs formation mechanisms, and the model was coupled with a CFD model to simulate the oxygen-enriched combustion of a diesel engine. The calculated results demonstrated that the coupled model can accurately predict ignition time and the change of in-cylinder pressure when the combined oxygen-enriched and EGR technique was used. The computed NO change with in-cylinder oxygen density agreed well with experiment results, and the computed result of the growth experience of PAHs showed that oxygen-enriched combustion can effectively suppress HACA reaction during PAHs formation, which leads to the reduction of large molecule PAHs, and this result agreed well with the observed situation that particle size diameter decreases with the increase of intake oxygen density. Ó 2013 Elsevier Ltd. All rights reserved.
1. Introduction The intake oxygen-enriched combustion research of an internal combustion engine began in the 1960s [1], with an original purpose to reduce emissions. The studies found that intake oxygen-enriched combustion will not only improve the emission characteristics, but also significantly reduce particulate emissions, ⇑ Corresponding author. Tel.: +86 22 27409558. E-mail address: [email protected] (G. Shu). 0306-2619/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.apenergy.2013.02.024
while the smoke capacity can significantly decrease 35–73% through the increase of intake oxygen concentration [2]. Researches also showed that intake oxygen-enriched combustion can increase the heat of combustion efficiency at some loads, and the total fuel consumption is reduced by 2–10% [3]. Engine power density can also be increased by applied intake oxygen-enriched combustion, which means the engine can be downsized with the same power output. The intake of oxygen concentration increased to 32% at same power output, which means the displacement of the engine can be reduced to 73% [4], to shorten the ignition delay
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Nomenclature V InCO2 V ExCO2 Mcycle Hfuel R Ncylinder Nnozzle
CO2 volume fraction in the intake gas CO2 volume fraction in the exhaust gas fuel supply quantity of each hole per cycle (kg/cycle) fuel consumption per hour (kg/h) engine speed (r/m) the number of cylinder the number of holes in each nozzle
Abbreviations EGR exhaust gas recirculation NO nitrogen oxide PM particulate matter CFD computational fluid dynamic
period and allow the use of low-quality fuel [3]. The increase in NOx emissions is a key issue for further development and application of constraints to internal combustion engines of oxygenenriched combustion technology. A large number of studies have shown that oxygen-enriched combustion for reducing particulate matter and soot is very significant, but the case did not take special precautionary measures, as NOx emissions will significantly increase [5–8], however, the emission characteristics of engine fueled with oxygenated fuel oxygen-enriched combustion is similar to that of oxygen-enriched intake combustion [9,10]. The diesel engine harmful emissions of NOx and PM purification methods contradict each other, thus making it difficult to fully meet the emissions regulations. EGR can reduce NOx emissions [11], but the EGR is too high, resulting in lower O2 concentration of the intake, combustion deterioration, and diesel engine performance degradation, increase in CO, HC and PM emissions levels. In this research, experimental and numerical simulation methods were combined, and appropriate EGR rate and oxygen-enriched air were used to suppress the high NO emission caused by oxygen-enriched combustion and achieved lower temperature combustion (compared to the unmodified engine’s combustion). The oxygenenriched intake air and proper ratio of low-temperature cold EGR can be combined to achieve low NO–Smoke emission and the oxygen concentration and EGR rate was delicately adjusted to achieve the goal in experimental research. In addition, the exhaust gas particle size distribution under different oxygen concentrations was detected. Polycyclic aromatic hydrocarbons (PAHs), as important soot precursors, have significant impact on smoke opacity in exhaust gas [12]. In this research, a CFD model was established using AVL FIRE software, and a new n-heptane mechanism containing the formation and oxidation mechanism of benzene, naphthalene, phenanthrene, pyrene was established. The CFD model was coupled with a chemical kinetic model to simulate oxygen-enriched combustion of a diesel engine. The PAHs evolution progress under Table B.1 Main parameters of 4100QBZL-2 diesel engine. Engine
Main parameters
Engine type Bore/stroke Compression ratio Power Displacement Injection timing Injection duration Combustion chamber Intake air type
Inline 4 cylinder, water-cooled, 4 stroke 100 mm/115 mm 17.5 80 kW/3200 rpm 3.6 L 12.5° BTDC 22° CA x bowl in piston Turbocharged, inter-cooler
PAHs HACA BSFC O2 CO2 rpm CA BTDC FSN ppm IVC EVO
polycyclic aromatic hydrocarbons hydrogen abstraction C2H2 addition reaction brake specific fuel consumption oxygen carbon dioxide revolutions per minute crank angle before top dead center filter smoke number part per million intake valve close exhaust valve open
oxygen enriched conditions was analyzed in this research. Experimental and numerical simulation methods were combined to explore the effects of intake oxygen-enriched combustion on emissions and combustion characteristics of a diesel engine. 2. Experimental setup and test method 2.1. Test engine and experiment setup Experiment was conducted on an inline four cylinder watercooled turbocharged 4100QBZL-2 diesel engine. The main parameters of the engine are shown in Table B.1. The exhaust gas was measured by a DIGAS 4000 exhaust gas analyzer from AVL. Smoke opacity was measured by a AVL 415S smoke meter. Power output was measured by a electric dynamometer produced by AVL, and a WaveBook data acquisition system was used to collect data. DMS500 fast particle analyzer produced by Combustion Company was used to measure the exhaust gas particle size distribution under different intake oxygen densities. The schematic of experimental setup is shown in Fig. C.1. The oxygen used in the experiment was supplied from lowtemperature adiabatic-storage liquid oxygen. Liquid oxygen was gasified though a liquid gas vaporizer. The pressure of the oxygen was very high, but reduced to match the pressure of the after compressor by a pressure-reducing valve. Then the oxygen mixed well with the air from the compressor and cooled EGR exhaust gas in a intake air buffer/mixing tank as intake gas. The experiment used a low-pressure EGR, and exhaust gas through the turbine into the exhaust buffer tank. Exhaust not containing EGR will emit to atmosphere through exhaust gas valve, while the other part of exhaust gas will enter the combustion chamber through the EGR valve. Exhaust gas flowed to the EGR cooler through the EGR valve, and the temperature will drop to less than 100 °C. The cooled EGR gas will benefit from reducing the combustion temperature and reduced NO emission. The cooled exhaust gas went into the intake buffer tank to mix well with oxygen and air. EGR rate can be controlled by changing the EGR valve and exhaust gas valves’ opening. The EGR rate can be calculated by Eq. (A.1).
Table B.2 Test condition. Engine speed (rpm) Fuel consumption per hour (kg/h) EGR rate (%) O2 volume fraction (%) Fuel injection strategy
1600 10 0, 21, 32, 48, 50 21, 22, 23, 24 Single injection
2200 14.2 0, 21, 32, 48, 50 21, 22, 23, 24 Single injection
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W. Zhang et al. / Applied Energy 107 (2013) 304–314 Table B.3 The formation mechanism of the benzene ring in SKLE + PAHs model. NO.
Reactions
1 2 3 4 5 6
C3H3 + C3H3 = A1 nC4H3 + C2H2 = A1 nC4H5 + C2H2 = A1 + H lC6H4 + H = A1 nC6H5 = A1 lC6H6 + H = A1 + H
Table B.4 The PAHs molecular structure of the one ring benzene to four rings Pyrene. Code
A1
A2
A3
A4
Name Molecular Structure
Benzene C6H6
Naphthalene C10H8
Phenanthrene C14H10
Pyrene C16H10
Table B.5 Main parameters of simulation. Speed
1600 rpm
1600 rpm
1600 rpm
Oxygen concentration EGR rate Pressure Temperature Mcycle N2 mole fraction O2 mole fraction CO2 mole fraction
21% 0% 0.22 MPa 145 °C 1.25E–5 kg 0.79 0.21 0
22% 32% 0.22 MPa 145 °C 1.25E–5 kg 0.738 0.22 0.042
24% 0% 0.22 MPa 145 °C 1.25E–5 kg 0.76 0.24 0
emission area is on the left side of the dash line, while low-carbon smoke emission area is on the right side of the dash line. As seen in Fig. C.4, the trend of the smoke emission changes with EGR rate at 2200 rpm is similar with the trend at 1600 rpm. Comparing Figs. C.3 and C.5, it is observed that the low carbon area at 2200 rpm is smaller than that at 1600 rpm. The increased smoke emission caused by using EGR can be effectively suppressed by the combination of intake oxygen-enriched intake and EGR enabling low temperature combustion under high oxygen concentration. Fig. C.6 shows NO emission change with EGR rate and intake oxygen concentration. At different EGR rates, NO emission rises rapidly with increasing oxygen concentration in the intake gas. The NO emission reduced with increasing EGR rate. With a 50% EGR rate and 24% intake oxygen concentration, combustion will deteriorate, which lead to a NO emission equal to the NO emission when no EGR is used, and 24% concentration of O2. The dash line in Fig. C.7 represents the NO emission of the unmodified engine at 1600 rpm. The upper left corner of the dash line of oxygen concentration and EGR rate combination can lead to NO emission lower than the unmodified engine, and adversely when moving towards the right area of the dash line. When comparing Figs. C.8 and C.6, the same NO emission trend shows at 2200 rpm and 1600 rpm; that is, the NO emission increases with increased intake oxygen concentration. However, at test engine speed of 2200 rpm the NO emission was lower than
2.2. Experiment method In the study, the engine was operated at a constant speed of 1600 rpm and 2200 rpm, 100% fuel throttle and full load condition. The fuel consumption and fuel supply rate curve were stable. During the tests, the dynamometer was adjusted to the full load at the setting speed, and real-time monitoring of the CO2 and O2 volume fractions in the intake gas through the AVL gas analyzer. Control of CO2 and O2 volume fractions in the intake gas achieved the concentration of test requirements through adjustment of the EGR and O2 throttle valves. When the engine operation was stable (which shows by speed and loads), data collection instruments were used to collect data. To compare the result, 1600 rpm and 2200 rpm of speed, full load, EGR rates were set at 0%, 21%, 32%, 48% and 50%. At each EGR rate, 21%, 22%, 23%, 24% O2 concentrations were tested. The test condition is shown in Table B.2. 3. Results and discussion 3.1. The effect of EGR and oxygen-enriched on smoke and NO emissions It can be seen from Fig. C.2, the smoke showed a rapid downward trend with the increase of O2 concentration in the intake gas. The trend was more obvious when the EGR rate is high. When the oxygen concentration increased to 24%, soot emission is much lower than that of original engine in the exhaust under each EGR rate. It can be seen from Fig. C.3, smoke emission decreased rapidly with the increase of O2 concentration. The dash line in Fig. C.3 represents the smoke emission of the unmodified engine (which is air for combustion and no EGR involved). High carbon smoke
Fig. C.1. Layout of experimental setup. 1. EGR cooler, 2. EGR temperature sensor, 3. EGR pressure sensor, 4. inter-cooler, 5. intake air buffer/mixing tank, 6. throttle valve, 7. pressure reducing valve, 8. liquid oxygen vaporizer, 9. liquid oxygen storage tank, 10. diesel, 11. control PC, 12. oil consumption meter, 13. data acquisition instrument, 14. diesel engine, 15. dynamometer, 16. compressor, 17. combustion DMS500, 18. exhaust pipe, 19. AVL gas analyzer, 20. exhaust pressure valve, 21. EGR valve, 22. AVL 415S smoke meter, 23. turbine, 24. over flow valve, 25. air flow throttle valve, 26. optical shaft encoder, 27. intake air temperature sensor, 28. intake air flow meter, 29. intake air pressure sensor, 30. AVL4000 gas analyzer and 31. oxygen sensor.
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4.0
3.2 2.8
2.0 1.8
2.5
0.5
2.4 Original Smoke emission 2.0 1.6
3.0
40
EGR Rate
3.6
Smoke / FSN
50
0% EGR 21% EGR 32% EGR 48% EGR 50% EGR
1.0
30
1.5
SPEED=2200 rpm LOAD=100% Smoke / FSN
20
SPEED=1600 r/m LOAD=100%
1.2
10 0.5
0.8 0.4 21.0
21.5
22.0
22.5
23.0
23.5
0 21.0
24.0
21.5
O2 Concentration / % Fig. C.2. Influence of O2 concentration and EGR rate on smoke emission at 1600 rpm, 100% load.
3.0 1.5
EGR Rate
40
1.1 1.8
30
20 2.6
SPEED=1600 rpm LOAD=100% Smoke / FSN
2.2
10 2.7 0 21.0
21.5
22.0
22.5
23.0
22.5
23.0
23.5
24.0
23.5
Fig. C.5. Influence of O2 concentration and EGR rate on smoke emission at 2200 rpm, 100% load (contour map).
engine speed of 1600 rpm, with the same EGR rate and intake oxygen concentration. This is because the reaction time of NO formation is reduced due to increase of engine speed. Therefore, the different combinations of EGR rate and intake oxygen concentration are needed to control NO emissions under the different engine speeds. In Fig. C.9, the dash line is the contour map of NO emission at 2200 r/min of speed and full load of the unmodified engine emission. The upper left of the figure is O2 concentration-EGR rate combination of low NO emission area, the bottom right corner of the figure is the high NO emission area.
50 2.6
22.0
O2 Concentration / %
3.2. The effects of EGR and enriched oxygen on in-cylinder pressure and heat release rate
24.0
O2 Concentration / % Fig. C.3. Influence of O2 concentration and EGR rate on smoke emission at 1600 rpm, 100% load (contour map).
Regarding the O2 concentration and EGR rate distribution, we selectively chose the 2200 rpm and 1600 rpm of speed and different oxygen concentrations and EGR rates to draw the in-cylinder pressures and calculated heat release curves to compare and analyze.
2500
0% EGR
SPEED=2200 rpm LOAD=100%
3.5 3.0
Smoke / FSN
SPEED=1600 rpm LOAD=100%
32% EGR
0% EGR 21% EGR 32% EGR 48% EGR 50% EGR
2.5
48% EGR 50% EGR
1500
NO / ppm
4.0
21% EGR
2000
2.0 1.5
1000
500
Original NO emission
1.0 0.5 0.0
Original Smoke emission
21.0
21.5
22.0
0
22.5
23.0
23.5
24.0
O2 Concentration / % Fig. C.4. Influence of O2 concentration and EGR rate on smoke emission at 2200 rpm, 100% load.
21.0
21.5
22.0
22.5
23.0
23.5
24.0
O2 Concentration / % Fig. C.6. Influence of O2 concentration and EGR rate on NO emission at 1600 rpm, 100% load.
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50
12.0M
240
10.0M
200
1275
EGR rate
30
20
Cylinder pressure / Pa
SPEED=1600 rpm LOAD=100% NO emission / ppm 790
1275 994
1838
8.0M
O2:21% EGR:0% O2:21% EGR:48%
6.0M 4.0M
SPEED=1600 rpm LOAD=100%
0 21.0
21.5
22.0
22.5
O2:22% EGR:32%
2.0M
23.5
40
O2:23% EGR:48%
2119
23.0
120 80
10 1556
160
Heat release / (J/deg)
431
40
0.0 -10
-5
0
10
15
20
0 25
Crank angle / º CA
24.0
O2 Concentration / %
5
Fig. C.10. Influence of O2 concentration and EGR rate on cylinder pressure and heat release rate at 1600 rpm, 100% load.
Fig. C.7. Influence of O2 Concentration and EGR rate on NO.
2000
0% EGR 21% EGR 32% EGR 48% EGR 50% EGR
1800 1600
NO / ppm
1400
SPEED=2200 rpm LOAD=100%
1200 1000 800 600 400
Original NO emission
200 21.0
21.5
22.0
22.5
23.0
23.5
24.0
O2 Concentration / % Fig. C.8. Influence of O2 concentration and EGR rate on NO emission at 2200 rpm, 100% load.
Four test conditions, 21% O2 and 0% EGR rate, 21% O2 and 48% EGR rate, 22% O2 and 32% EGR rate, 23% O2 and 48% EGR rate, were chosen to be compared and analyzed. The results are shown in Fig. C.10. As Fig. C.10 shows, when the EGR rate is 48%, the in-cylinder pressure is lower and the ignition delay time is longer due to more CO2 getting into each cylinder involved in combustion. In the case of 22% O2 and 32% EGR rate, there is no significant difference shown in the in-cylinder pressure and heat release curve compared with the unmodified engine, which indicated that although the oxygen concentration increased 1%, nearly the same in-cylinder pressure and heat release can be reached by using 32% EGR. Four conditions are chosen to be compared and analyzed: 21% O2 and 0% EGR, 22% O2 and 21% EGR, 23% O2 and 32% EGR, 24% O2 and 50% EGR. The in-cylinder pressure and heat release of these four conditions are shown in Fig. C.11. As Fig. C.11 shows, in-cylinder pressure of 22% O2 and 20% EGR is very close to that of the unmodified engine, and the heat release delayed a little. In-cylinder pressure in the case of 23% O2 and 32% EGR decreased noticeably and ignition time was further delayed. In the case of 24% O2 and 50% EGR, ignition time was delayed by 4° CA, the maximum pressure was reduced by 2 MPa, and the combustion deteriorated. 3.3. The effects of EGR and enriched oxygen on brake specific fuel consumption and power Fig. C.12 shows the effects of oxygen concentration and EGR rate on brake specific fuel consumption. The effect of oxygen con-
50 276
12.0M
30
Cylinder pressure / Pa
SPEED=2200 rpm LOAD=100% NO / ppm
20
10
513 775
0 21.0
985 21.5
22.0
1221 22.5
SPEED=2200 rpm LOAD=100%
280
10.0M
240 200
8.0M
160 6.0M 4.0M
O2 :21% EGR:0%
120
O2 :22% EGR:21%
80
2.0M
O2 :23% EGR:32%
1458 23.0
23.5
24.0
O2 Concentration / % Fig. C.9. Influence of O2 concentration and EGR rate on NO emission at 2200 rpm, 100% load (contour map).
0.0 -10
O2 :24% EGR:50% -5
0
5
10
15
20
Heat release / (J/deg)
EGR rate
40
40 0 25
Crank angle / º CA Fig. C.11. Influence of O2 concentration and EGR rate on cylinder pressure and heat release rate at 2200 rpm, 100% load.
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BSFC / ( g. (kw. h) -1)
200
160
120
SPEED=1600 rpm LOAD=100%
80
0% EGR 21% EGR 32% EGR 48% EGR
40
50% EGR 0
21
22
23
24
O2 Concentration / % Fig. C.12. Influence of O2 concentration and EGR rate on brake specific fuel consumption at 1600 rpm, 100% load.
52 50
Power / KW
0% EGR 21% EGR
48
32% EGR 48% EGR
46
centration on brake specific fuel consumption is not obvious, however, the specific fuel consumption increase significantly when EGR is more than 32%. Fig. C.13 shows the influence of oxygen concentration and EGR rate on power, similar with brake specific fuel consumption, the intake oxygen concentration has little influence on power, and power reduces with increasing of EGR. When EGR rate reached 32%, the power reduces a little (<3%), when EGR rate reached 50%, the decline reached about 8 kW (>15%) comparing with 0% EGR. As Fig. C.14 shows, at 1600 rpm of engine speed, full load, the power of original engine is 50.7 kW. The upper limit of power loss was set as 5% (48 kW), which was shown as a dash line at the contour map of EGR, intake oxygen concentration and power output. The area under dash line in the figure is the area of power loss less than 5% and above the dash line area the power loss is more than 5%. It can be observed that the power loss under different oxygen concentration can be controlled within 5% when EGR rate was controlled within 40%. Fig. C.15 shows the brake specific fuel consumption at 2200 rpm. Similar to 1600 rpm, the brake specific fuel consumption under the same EGR rate and different intake oxygen concentration has little difference, and the brake specific fuel consumption increases significantly when the EGR rate is greater than 32%. Fig. C.16 shows the power output at different oxygen concentration and EGR rates at 2200 rpm. The power output is reduced 5% when the EGR rate is 48%, and the power loss increases to 6 kW (>9%) when 50% EGR. The dash line 61 kW in Fig. C.17 was the upper limit of 5% power loss when engine at 2200 rpm, full load. Power loss can be controlled within 5% under various oxygen concentrations with 45% EGR.
50% EGR
3.4. The NO–Smoke emission area at different intake oxygen concentration and EGR rates
44
SPEED=1600 r/m LOAD=100%
42 40 21.0
21.5
22.0
22.5
23.0
23.5
24.0
O2 Concentration / % Fig. C.13. Influence of O2 concentration and EGR rate on power at 1600 rpm, 100% load.
Concerning the smoke emission, NO emission, brake specific fuel consumption, and power at 1600 rpm and 2200 rpm, the oxygen concentration from 21% to 24% and the EGR rate from 0% to 50% can be matched to realize lower Smoke–NO emissions than the unmodified engine with loss of brake specific fuel consumption and power no more than 5%. The contour map of Smoke emission and NO emission at 1600 rpm and 2200 rpm, with the 5% power loss line, results in a NO–Smoke emission and power loss line contour map as shown
50
240
45 46
200
40
BSFC / ( g. (kw. h) -1)
48 49
EGR rate
50 30
51
20
SPEED=1600 rpm LOAD=100% Power / kW
10
0 21.0
160
120
SPEED=2200 rpm LOAD=100%
0% EGR 21% EGR
80
32% EGR 48% EGR
40
50% EGR
0 21.5
22.0
22.5
23.0
23.5
24.0
O2 Concentration / % Fig. C.14. Influence of O2 concentration and EGR rate on power of 1600 rpm, 100% load (contour map).
21
22
23
24
O2 Concentration / % Fig. C.15. Influence of O2 concentration and EGR rate on brake specific fuel consumption at 2200 rpm, 100% load.
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64
Power / kW
60 56 0% EGR
SPEED=2200 rpm LOAD=100%
52
21% EGR 32% EGR 48% EGR
48
50% EGR
44 40 21.0
21.5
22.0
22.5
23.0
23.5
24.0
O2 Concentration / % Fig. C.16. Influence of O2 concentration and EGR rate on power at 2200 rpm, 100% load.
50
4. The construction of numerical models and the analysis of simulation results
61 62
EGR Rate
40
63
4.1. The combined model containing detailed PAHs and NOx formation mechanisms
30 64 20
10
0 21.0
SPEED=2200 rpm LOAD=100% Power / kW
65
21.5
22.0
22.5
23.0
23.5
24.0
O2 Concentration / % Fig. C.17. Influence of O2 concentration and EGR rate on power of 2200 rpm, 100% load (contour map).
(a) 1600rpm
in Fig. C.18. The white area in the figure is lower NO–Smoke emission area. Lower NO–Smoke emission than the unmodified engine with no significant power loss in this area can be attained by matching EGR and O2 in this area. In Fig. C.18a, the white area represents 1600 rpm, full load, 30–40% EGR rate, 21.5–22.5% O2. Matching the EGR rate and oxygen concentration in this area can achieve lower NO–Smoke emission than the unmodified engine. Similarly, the white area in Fig. C.18b is the lower NO–Smoke emission area of 2200 rpm, full load, 20–45% EGR rate, 22–24% O2. The matches of EGR rate and oxygen concentration can achieve lower NO–Smoke emission. When comparing Fig. C.18a and b, the size and shape of lower NO–Smoke emission areas are different in the engine’s various work conditions and produces different experiment results. The changes trends of combustion and emissions with oxygen concentration and EGR at 1600 rpm were similar with the case of those at 2200 rpm. In the next section, the simulation researches were conducted to study the effects of oxygen-enriched combustion and EGR on NO formation and PAH growth experience at 1600 rpm and full load.
The composition of diesel fuel is complex and difficult to accurately simulate. n-Heptane is usually used to simulate the combustion and emission characteristics, because its cetane number is close to diesel fuel [13]. The detailed chemical model developed by the Lawrence Livermore National Laboratory (LLNL) contains 631 species and 2827 reactions, and was validated by shock tube and rapid compression machine experiments, and was proved to be reliable [14]. This research used a reduced skeleton model (SKLE model) based on the LLNL detailed model, and contains 41 species and 63 reactions [15]. The new SKLE + PAHs detail model contains the detailed PAHs mechanism proposed by Wang containing 99 species and 527 reactions [16]. The detailed PAHs mechanism of Wang was validated by premixed laminar flame [17,18].
(b) 2200rpm
Fig. C.18. NO–Smoke emission zone distribution at 1600 rpm and 2200 rpm, 100% load. (1) Red Zone: Power Loss zone, (2) Green Zone: High Smoke Emission Zone, (3) Blue Zone: High Smoke and NO Emission Zone, (4) Cyan Zone: High NO Emission Zone and (5) White Zone: Low Smoke and NO Emission Zone. (For interpretation of the references to colour in this figure legend, the reader is referred to the web version of this article.)
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Fig. C.19. The growth of PAHs with HACA reactions.
Fuel Injection Rate
0.0012
0.0008
0.0004
0.0000 0
4
8
12
16
20
Injection Duration / deg Fig. C.20. Spray curve and combustion chamber 1/6 mesh.
The NOx formation mechanism in this research is the Zeldovich NOx mechanism validated by Dr. Golovitchev from Chalmers technology university, and contains 4 species and 9 reactions [19]. The acetylene formation and oxidation reactions and the C2Hx ? C6Hx formation and oxidation reactions added with benzene and phenyl formation reactions to form a complete description of the benzene rings formation. The reactions related to the formation of benzene and benzene radical are list in Table B.3. Once the formation of benzene ring, naphthalene, phenanthrene, pyrene can form through hydrogen abstraction C2H2 addition reactions (HACA) [17]. The molecular structures of benzene, naphthalene, phenanthrene, pyrene are shown in Table B.4. The formation paths from benzene to pyrene are shown in Fig. C.19. Through the combination of ethylene reaction in the SKLE model, acetylene and propargyl precursor reactions, ethylene decomposition follow-up reactions, and the formation and oxidation of PAHs in Wang’s model, so n-heptane SKLE model combines well with Wang’s PAHs model. When the NOx formation model joined the combined model, the new SKLE + PAHs model contains 101 components, 329 elementary reactions, reflecting PAHs and NOx changes in the n-heptane combustion process, and can be applied to coupling calculation with the CFD model. 4.2. The establishment of the CFD model and set up of initial condition In this research, ESE Diesel in FIRE software was used to establish the combustion chamber model. Construction parameters of
the test engine were set in the general engine parameter and piston movement specification. Combustion chamber type was chosen and modified according to the geometric data of the piston in Sketcher, as were the fuel injector position, the number of injector holes, the spray angle and other parameters. The 2D grids of the combustion chamber were created in Mesh, and then generated 3D solid models based on the 2D grid. Due to the long calculation time of the full combustion chamber grid, the 1/6 model shown at Fig. C.22 was established according to the injector holes in the nozzle to save the computing time. In the initial condition set of the coupling calculation, only time scale between intake valve closing (IVC) to exhaust valve opening (EVO) was calculated. The top dead center (TDC) was set as 0° CA, and the calculated crank angle range of 132–115° CA, total 247° CA, which was determined by IVC and EVO timing. The fuel injection time was set as 12.5° CA before TDC. The fuel supply curve shown in Fig. C.20 was maintained during the calculation. The fuel supply quantity was measured by a fuel consumption meter and the fuel each cycle supplied of the 1/6 model can be calculated by Eq. (A.2). The initial turbulent kinetic energy was calculated according to the engine speed and the intake air flow. In AVL Fire’s solver steering file, the k–epsilon model was selected as the gas turbulent model, the turbulent diffusion model was activated in spray module, and the Dukowicz fuel evaporation model and the Wave fuel broken model should be involved in numerical simulation at the same time.
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2500
25.0M
Mean pressure / Pa
20.0M
Mean temperature / K
Detail model SKLE model SKLE+PAHs model
15.0M
10.0M
Detail model SKLE model SKLE+PAHs model
2000
1500
1000
5.0M 500 0.0 -100
-50
0
50
-100
100
-50
Crack angle / º CA
0
50
100
Crack angle / º CA
(a) pressure
(b) temperature
12.0M
SPEED=1600 rpm LOAD=100%
10.0M
O2 :21% EGR:0%
0.005
NO mass fraction
Cylinder pressure / Pa
Fig. C.21. The simulation result comparison with detailed model, SKLE model and SKLE + PAHs model.
Experimental Calculated
8.0M 6.0M
SPEED=1600 rpm LOAD=100%
0.004
0.003
0.002
Calculated O2 :21% EGR:0%
4.0M
O2 :22% EGR:32%
0.001
O2 :24% EGR:0%
2.0M 0.000 0
0.0 -100
-50
0
50
10
100
20
30
40
Crank angle / º CA
Crank angle / º CA Fig. C.22. Experimental and calculated cylinder pressure on 1600 rpm, 100% load.
Fig. C.24. Influence of O2 concentration and EGR rate on NO mass fraction at 1600 rpm, 100% load.
12.0M
1800
10.0M
1600 8.0M 1400 Calculated 1200
6.0M
O2 :21% EGR:0% O2 :22% EGR:32% O2 :24% EGR:0%
1000
4.0M -20
-10
0
10
20
30
40
Crank angle / º CA
Size spectral density / (dN/ dl ogDp/ cc)
2000
5x108
SPEED=1600 rpm LOAD=100%
Cylinder pressure / Pa
Cylinder temperature / K
2200
SPEED=1600 rpm LOAD=100%
4x108
3x10
Experimental O2 :21% EGR: 0%
8
O2 :22% EGR: 0% O2 :23% EGR: 0%
2x108
O2 :24% EGR: 0%
1x108
0 10
Fig. C.23. Influence of O2 concentration and EGR rate on pressure and temperature at 1600 rpm, 100%.
100
1000
Dp / nm Fig. C.25. Influence of O2 concentration on particle size distribution (experimental).
4.3. Validation of the chemical kinetic model and the coupled CFD calculation With the chemical kinetic analysis software CHEMKIN, developed by Sandia National Laboratory, a calculated result comparison
was carried out on three models: the specifically LLNL n-heptane detail model, reduced SKLE model, and SKLE + PAHs detail model, to verify the accuracy of the new built SKLE + PAHs detail model
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4.0x10-6
SPEED=1600 rpm LOAD=100%
1.2x10
SPEED=1600 rpm LOAD=100%
1.0x10-4
Calculated O2:21% EGR:0% O2:22% EGR:32% O2:24% EGR:0%
8.0x10-5 6.0x10-5
Calculated O2 :21% EGR:0% O2 :22% EGR:32% O2 :24% EGR:0%
3.0x10-6
A2 Mass fraction
A1 Mass fraction
-4
4.0x10-5
2.0x10-6
1.0x10-6
-5
2.0x10
0.0 0
5
10
15
25
30
0.0
35
-5
5
10
15
20
25
Crank angle / º CA
(a) A1 Mass fraction
(b) A2 Mass fraction
35
SPEED=1600 rpm LOAD=100% 2.0x10-6
Calculated O2 :21% EGR:0% O2 :22% EGR:32% O2 :24% EGR:0%
6.0x10-6
30
2.5x10-6
SPEED=1600 rpm LOAD=100%
8.0x10
4.0x10-6
2.0x10-6
Calculated O2 :21% EGR:0% O2 :22% EGR:32% O2 :24% EGR:0%
1.5x10-6
1.0x10-6
5.0x10-7
0.0 -5
0
Crank angle / º CA
-6
A3 Mass fraction
20
A4 Mass fraction
-5
0
5
10
15
20
25
30
35
0.0 -5
0
5
10
15
20
25
Crank angle / º CA
Crank angle / º CA
(c) A3 Mass fraction
(d) A4 Mass fraction
30
35
Fig. C.26. Influence of different O2 concentration on mean PAHs mass fraction.
in the research. A zero-dimension single region IC engine quick compressor model was applied in the chemistry kinetic calculation, with the boundary and initial conditions the same as the measured value of the diesel engine at 2000 rpm and full load. As can be seen from the chemical kinetic calculation result in Fig. C.21, difference exists in the SKLE model and detail model at low-temperature combustion (<800 K) and high-temperature (>1000 K), but can fulfill the calculation demands as a whole. With the SKLE + PAHs model and SKLE model added, the peak values of the pressure and temperature curve have a similar drop, but the shapes of the curve agree with the detail model totally thus the combustion state can be reflected in the cylinder. The comparison between experimental results at full load, 1600 rpm and the simulation results of the CFD model coupled with the chemical kinetic model are shown in Fig. C.22. The simulation results agree well with the experimental results indicating that the calculated result using the SKLE + PAHs detail model and coupled with CFD model in FIRE software can fulfill the real measured simulation demand.
It can be seen from Fig. C.23, the calculated in-cylinder pressure and temperature in the case of 22% O2 and 32% EGR rate are lower than those at the 21% O2 and 0% EGR rate. The in-cylinder pressure and temperature are higher than those at the 24% O2 and 0% EGR rate. Comparing Figs. C.23 and C.10, it can be seen that the calculated in-cylinder pressure agrees well with the experiment results, indicating that the model simulation can reflect the temperature and in-cylinder pressure changes when O2 and CO2 concentration change. As shown in Fig. C.24, the mass fraction of NO at 22% O2, 32% EGR is lower than that at the 21% O2, 0% EGR rate. The NO emission of 22% O2 can be lower than the unmodified engine when 32% EGR is used, however, the NO mass fraction in the case of 24% O2, with no EGR involved, is almost double that of the NO mass fraction at 21% O2, 0% EGR. The simulation result of NO emission is identical with the experimental results in Fig. C.6.
4.4. Simulation results
PAHs are the main precursor of soot, they play a key role in the amount of soot [20], soot particles aggregation and growth [21,22], therefore the probabilistic prediction of soot formation by PAHs via the corresponding relationship between generation amount of PAHs and soot is possible [23,24]. So numerical simulation of PAHs combined with soot experimental measurements were used to
Three conditions were chosen to simulate: 21% O2 and 0% EGR, 22% O2 and 32% EGR, 24% O2 and 0% EGR, with 1600 rpm of speed and full load. The calculation results were compared and analyzed. The main parameters of simulation are shown in Table B.5.
4.5. The effect of intake oxygen concentration on particle size distribution in exhaust gas
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study the effect of intake oxygen concentration on soot particle size distribution in the exhaust. Fig. C.25 shows the effects of intake oxygen concentration on particle number and size distribution in exhaust gas. Particle size distribution ranges from 5 to 1000 nm. When intake oxygen concentration is 21%, most particles concentrate at about 100 nm. The large particle number reduces and small particles increase at increased intake oxygen concentration. When oxygen concentration increases to 24%, most particle diameters reduce to about 15 nm. The results indicated that oxygen-enriched combustion has significant suppression effect on diameter growth of particles. Fig. C.26 shows the mass fractions of A1–A4 (benzene–pyrene) change with the crank angle, at 22% of O2 concentration and 32% of EGR rate. The formation and oxidation of PAHs were delayed with the extension of ignition delay time. The PAHs formation at 24% O2 and no EGR are obviously lower than at 21% O2, no EGR and 22% O2, 32% EGR. These results indicated that with an increase of O2 concentration in the chamber, the formation of PAHs was significantly suppressed by the oxygen-enriched environment. 5. Conclusion (1) Using a high EGR rate can achieve low NO emission when oxygen-enriched combustion is applied; however, an excessively high EGR rate will lead to combustion deterioration, plus increase of brake specific fuel consumption, and higher smoke emission even at an oxygen-enriched combustion. (2) The proper combination of oxygen concentration and EGR rate can achieve low NO–Smoke emission. The size and shape of the low NO–Smoke emission area are different with different engine speeds and loads as shown in the specific experiment results. (3) A combined n-heptane model containing a detailed PAHs mechanism was developed in this research. The calculation results using the combined model, coupled with the CFD model predict well the NO and smoke emissions, and are suitable for qualitative research of oxygen-enriched combustion of diesel engines. (4) The oxygen-enriched combustion of diesel engines can reduce smoke opacity effectively and suppress the diameter growth of soot particles, while reducing the number of large soot particles; however, this leads to the increase of small particles with diameters of about 15 nm. (5) The numerical simulation results indicated that oxygenenriched combustion can suppress HACA reactions and reduce the formation of large molecule PAHs (naphthalene, phenanthrene, pyrene), while also reducing the mass fraction of benzenes at the same time.
Acknowledgements This work was supported by the National Natural Science Foundation of China (No. 50976077). The authors also grateful to Department of Education, Yunnan Province, China (KKJA201202025, KKJA201256011) and Kunming University of Science and Technology, Yunnan Province, China (KKSY201202143, KKSY201256142). Appendix A
EGR% ¼ ðV InCO2 =V ExCO2 Þ 100%
ðA:1Þ
M cycle ¼ Hfuel =ð60 ðR=2Þ Ncylinder Nnozzle Þ
ðA:2Þ
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