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Improvement of DME HCCI engine combustion by direct injection and EGR
Fuel 113 (2013) 617–624
Contents lists available at SciVerse ScienceDirect
Fuel journal homepage: www.elsevier.com/locate/fuel
Improvement of DME HCCI engine combustion by direct injection and EGR Jinyoung Jang a, Youngjae Lee a, Chongpyo Cho a, Youngmin Woo a, Choongsik Bae b,⇑ a b
Korea Institute of Energy Research, 152 Gajeong-ro Yuseong-gu, Daejon 305-343, Republic of Korea Korea Advanced Institute of Science and Technology, 373-1 Kusong-dong, Yusong-gu, Daejon 305-701, Republic of Korea
h i g h l i g h t s Evaluation of the effect of injection location and EGR on performances in DME HCCI engine. In-cylinder direct injection was better than port DME injection engine for DME HCCI engine. Combustion is changed from homogeneous combustion to mixing controlled combustion by varied in-cylinder injection timing. In-cylinder direct injection with EGR made more power than without EGR.
a r t i c l e
i n f o
Article history: Received 22 January 2013 Received in revised form 30 May 2013 Accepted 1 June 2013 Available online 17 June 2013 Keywords: HCCI (homogeneous charge compression ignition) EGR (exhaust gas recirculation) Direct injection Port injection DME (dimethyl ether)
a b s t r a c t To extend the operating range of a homogeneous charge compression ignition (HCCI) engine fuelled with DME, direct injection and exhaust gas recirculation (EGR) were used to investigate their effects on the change of combustion characteristics in this experiment. Single cylinder engine was used with port fuel injector and direct injector. EGR line was connected from exhaust pipe to intake pipe. The experimental result for direct injection was compared with that for port injection. The gross indicated mean effective pressure (IMEPgross), which represented the operation region, for direct injection was higher than that for port injection due to late combustion resulting from lower temperature combustion induced by evaporating latent heat of injected DME in the cylinder. There was optimal direct injection timing for DME HCCI engine because the combustion characteristics were changed from overall lean combustion to locally rich combustion. IMEPgross was increased by EGR. It was because EGR made tardy combustion and increased positive work during expansion stroke. Consequently, direct injection with optimal injection timing and EGR increased the IMEPgross by 22% and 55% compared to port injection case, respectively. Crown Copyright Ó 2013 Published by Elsevier Ltd. All rights reserved.
1. Introduction The homogeneous charge compression ignition (HCCI) engine is a promising concept for future automobile engines and stationary power plants. In HCCI, a premixed air/fuel mixture is inhaled as in conventional spark ignition (SI) engines, which it is auto-ignited as in conventional compression ignition (CI) engines. HCCI uses spontaneous ignition, which is governed primarily by chemical kinetics during the compression process with the premixed air/fuel mixture. Basic characteristics of the HCCI engine, such as combustion stability and no flame propagation, have been studied by previous researchers [1,2]. HCCI engines emit less nitrogen oxide (NOx) than conventional engines due to the low combustion temperature resulting from the high air excess ratio [3]. Particulate matter (PM) emitted from the HCCI engine is also less than conventional ⇑ Corresponding author. Tel.: +82 42 350 3044; fax: +82 42 350 5023. E-mail address: [email protected] (C. Bae).
CI engines due to the lean homogeneous charge mixture [3–5]. However, HCCI engine operation has some disadvantages, such as difficulty in controlling the combustion phasing, much lower maximum load than typical SI and CI engines, high pressure and noise at a low air excess ratio or high load condition, and higher carbon monoxide (CO) and hydrocarbon (HC) emissions than SI and CI engines due to the low combustion temperature [4,6]. Dimethyl ether (DME) is considered an alternative fuel for CI engine [7,8]. DME is a high cetane numbered fuel similar to diesel that does not produce soot due to its molecular structure; DME has no double C–C bonds and contains oxygen atoms which oxidize intermediate soot product. DME is easily evaporated and mixes well with the air so that the rich regions are avoided, leading to soot-free combustion. The ignition delay of DME in a CI engine is shorter than that of diesel under the same operating conditions as the evaporation rate of DME is two or three times faster than diesel [9–11]. Because of the shorter ignition delay, various control schemes to change the auto-ignition timing have been proposed in
0016-2361/$ - see front matter Crown Copyright Ó 2013 Published by Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.fuel.2013.06.001
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Nomenclature C CA5 CA50 CA95 mfuel QLHV QHVi Wi xi ðxCO2 ÞE ðxCO2 ÞI k
gc gf
carbon 5% mass fraction burned point 50% mass fraction burned point 95% mass fraction burned point dimethyl ether quantity lower heating value of the fuel lower heating value of each species indicated work done on the piston mass fraction CO2 concentration in the exhaust gas CO2 concentration in the intake gas air excess ratio combustion efficiency indicated thermal efficiency
Subscripts a air f fuel Abbreviations AC alternating current ABDC after bottom dead center
DME HCCI such as internal and/or external exhaust gas recirculation (EGR), fuel mixing, and additives [12]. EGR is considered to be an effective method for controlling the combustion phasing and reducing exhaust emissions, especially NOx. In SI engines, EGR can reduce NOx emissions without the penalties of increase in other emissions or excessive fuel consumption [12]. The effects of EGR are classified as dilution-, thermal-, and chemical effects [13,14]. The three EGR mechanisms affect the combustion rate, burn duration, combustion temperature, and exhaust emission. In HCCI, thermal and dilution effects are more important than chemical effects of EGR due to the lower combustion temperature. The retarded onset of auto-ignition with EGR for a fixed induction air temperature is due to the thermal effect of the cooler compressed gas temperatures during the compression stroke [15–19]. Oxygen (O2) dilution due to EGR retards the auto-ignition timing [17]. The heated intake charge by EGR advances the auto-ignition timing [19]. There is a potential for improving volumetric efficiency due to fuel evaporation in the cylinder resulting in increased air density with the direct injection during induction stroke. Fuel stratification by direct injection can also improve the HCCI engine combustion with variation of the injection timing [20]. Therefore, direct injection can be one of the ways to improve emissions and power in HCCI engine. In this study, the effect of fuel injection strategy and EGR on an HCCI engine combustion fueled with DME was investigated. Output performance related to power and efficiency and emission were characterized with EGR rate and for the various injection locations in port and cylinder at different injection timings. To find the way to improvement of indicated mean effective pressure, direct injection with the change of injection timing and exhaust gas recirculation were used and combustion characteristics was analyzed.
ATDC BBDC BTDC CAD CI CO CO2 DME DOHC EGR EVC EVO HC HCCI IVC IVO IMEPgross NOx O2 SI PM TDC
after top dead center before bottom dead center before top dead center crank angle degree compression ignition carbon monoxide carbon dioxide dimethyl ether double overhead camshaft exhaust gas recirculation exhaust valve close exhaust valve open hydro carbon homogeneous charge compression ignition intake valve close intake valve open gross indicated mean effective pressure nitrogen oxide oxygen spark ignition particulate matter top dead center
presented in Table 1. The engine was a four-stroke, water-cooled, and single-cylinder double overhead camshaft (DOHC) engine. The engine speed and load were controlled by an alternating current (AC) dynamometer (82 kW, Unico Co.). The DME fuel injector was placed upstream, 30 cm away from the intake port, to form a homogeneous air/fuel mixture and another injector was installed in cylinder to inject the fuel into the cylinder. DME was pressurized to 5 MPa by nitrogen gas. The DME injector for port injection was inclined at 30° to minimize wall wetting and improve air/fuel mixing. A small amount (500 ppm) of lubrication enhancer (Infineum, R655) was added to the DME to protect the injector and fuel feeding system due to lower lubricity of DME. For the exhaust gas recirculation (EGR), EGR line was connected from exhaust pipe to intake pipe, as shown in Fig. 1. The in-cylinder pressure was measured by a piezoelectric pressure transducer (Kistler, 6052b). The exhaust gases were analyzed with a gas analyzer (HORIBA MEXA-1500d) to measure the HC, CO, CO2, NOx and O2 in the exhaust and CO2 in the intake. The experimental data were acquired by a data acquisition system (IO Tech, Wavebook 516 series with 12 bits of resolution, 1 ls/channel sampling rate, and an accuracy of 0.025%). The exhaust gas recirculation (EGR) rate is defined as in Eq. (1) in terms of its mole concentration. It is experimentally determined based on a comparison between the CO2 mole-concentration in the intake gas, ð xCO2 ÞI , and the CO2 concentration in the exhaust gas, ð xCO2 ÞE .
EGR rateð%Þ ¼
ðxCO2 ÞI 100 ðxCO2 ÞE
ð1Þ
2.2. Experimental conditions 2. Experiments 2.1. Experimental apparatus Fig. 1 shows a schematic diagram of the experimental apparatus used in this study. The detailed specifications of the test engine are
To compare the effect of injection location on the DME HCCI combustion, the injection location was changed between the port and in-cylinder. The air excess ratio (k) was changed from 3.1 to 4.8. In-cylinder direct injection timing was varied from 20 to 350 CAD to investigate the effect of injection timing while the injection timing of the port injection was maintained at 20 CAD.
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Fig. 1. Experimental engine set-up.
Table 1 Engine specification. Bore (mm) Stroke (mm) Compression ratio Intake/Exhaust valve duration Intake/Exhaust lift duration (CAD) Valve timing (CAD) IVO (BTDC) IVC (ABDC) EVO (BBDC) EVC (ATDC) DME injection pressure (MPa)
82 93.5 13 228/228 8.5/8.4 1 39 42 6 5
The EGR rate was changed from 0% to 60% to change the combustion phasing and improve the gross indicated mean effective pressure (IMEPgross). In-cylinder and intake/exhaust pressure data for one hundred consecutive engine cycles were used for the analysis of combustion and IMEPgross [21].
work. It is reason why IMEPgross of direct injection case is higher than that of port injection case. IMEPgross for two injection locations is shown in Fig. 3a, as a function of air excess ratio (k). In both port injection and in-cylinder direct injection, IMEPgross would be reaching their highest value as injected DME quantity was increased. The ignition timing was advanced and/or the burn duration was reduced during the compression stroke as DME quantity increased, resulting in increased negative work. After this end, the IMEPgross did not increase further. In the case of port injection, maximum pressure rate at highest IMEPgross was 0.8 MPa which was higher than criteria for load limit used by Jung et al.[22]. In the case of direct injection, IMEPgross was improved approximately by 3% compared to port injection at each air excess ratio condition due to the elongated ignition delay resulted from slow
3. Results and discussion 3.1. Comparing between port and direct injection To understand the improvement of performance by direct injection, firstly, IMEPgross and auto-ignition timing were compared with those by port injection. The injection quantity of DME was varied and the injection timing was maintained at 20 crank angle degree (CAD) after top dead center (TDC) of intake stroke. Fig. 2 shows comparing results of in-cylinder pressure and heat release rate profile for port and direct injection case. The start of the heat release with direct injection was more delayed than with the port injection while the burn duration was similar at the same operating conditions. It usually led to less negative compression
Fig. 2. In-cylinder pressure and heat release rate profile comparison between port and direct injection.
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reaction rate, which was mainly affected by lower mixture temperature. The injected fuel at the port produced a homogeneous mixture during the induction stroke without additional temperature drop in the cylinder. However, there was a temperature drop due to the latent heat of liquid DME when it was injected into the cylinder and evaporated. In Fig. 3b, auto-ignition timing of direct injection was more retarded and negative work of the direct injection decreased more than those of port injection due to the temperature drop with the direct injection of DME fuel [23]. Fig. 3b shows the CA5 (5% mass fraction burned) with respect to injection location and air excess ratio. Combustion phasing was presented with CA5, CA50 and CA95, which are 5%, 50% and 95% mass fraction burn points, respectively. The 5% burn point of the combustion is closely related to the auto-ignition timing. Since 5% burn point is given in CAD, if CA5 is earlier than 360 CAD, it occurs before TDC. Ignition timing for direct injection was later than that for port injection, in spite of same injection timing. In the case of direct injection, the advance of ignition timing was not significant by increasing DME injection quantity.
Through previous section, it was found that direct injection was better way to improve IMEPgross on DME HCCI engine operation than port injection. Therefore, in this section, the effect of DME direct injection timing on HCCI engine was investigated to seek the optimal injection timing. Fig. 4 shows the variation of IMEPgross and NOx emission as a function of injection timing and also compares the IMEPgross of port injection case. The change of IMEPgross had three regions as a function of injection timing. The IMEPgross increased in the first region approximately from 20 to 260 CAD. The IMEPgross decreased within
the second region approximately from 260 to 320 CAD. The third region was later than 320 CAD and the IMEPgross was augmented. The change of IMEPgross with injection timing was explained by combustion phasing, as shown in Fig. 5 and 6. Fig. 5 shows the in-cylinder pressure and heat release rate profiles as a function of direct injection timing. And, Fig. 6 shows the combustion phasing as a function of direct injection timing. Within the first region approximately from 20 to 260 CAD, the CA5 and CA50 was retarded due to the cooling effect of evaporated DME and late injection timing. Retarded CA50 not only decreased negative work during compression but also increased positive work during expansion. Therefore, late direct injection made IMEPgross increase. However, during the second region approximately from 260 to 320 CAD, IMEPgross decreased as the burn duration from CA5 to CA95 was shorter than the first region and the combustion process ended before the top dead center (TDC) resulting in increased negative work and decreased positive work, in spite of the retarded onset of auto-ignition. The combustion characteristic was changed from relatively leaner to locally-richer combustion due to a change of mixture homogeneity as direct injection timing was delayed. The locally-rich combustion took place by a lack of mixing time resulting from the late injection timing with direct injection. The locally richer combustion by direct injection timing could be explained using NOx emission result as a function of injection timing, as shown in Fig. 3b. NOx emission was negligible with early injection and increased with late injection near the TDC. High temperature and oxygen concentrations resulted in high NOx formation rates [24]. NOx emission was recognized when injection timing was later than 240 CAD and the peak value of NOx emission occurred near 320 CAD due to locally-rich and higher combustion temperatures. During the third region approximately from 320 to 350 CAD, the onset of auto-ignition was more retarded than TDC at lower
Fig. 3. IMEPgross and CA5 comparison between port and direct injection as a function of injection location and air excess ratio (k).
Fig. 4. IMEPgross and NOx emission variation as a function of direct injection timing.
3.2. Improving engine performance by direct-injection timing optimization
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where xi is the mass fraction of CO, H2, and unburned HC, QHVi represents the lower heating values of these species, and the subscripts f and a denote fuel and air. The indicated thermal efficiency, gf, was defined as the ratio of the work done on the piston during the compression and expansion strokes to the heat supplied by the fuel, as shown in the following equation [25].
gf ð%Þ ¼
Fig. 5. In-cylinder pressure and heat release rate profile as a function of direct injection timing.
Wi 100 mfuel Q LHV
ð3Þ
where mfuel is the DME mass, QLHV is the lower heating value of the fuel, and Wi is the indicated work done on the piston calculated using in-cylinder pressure data over the operating cycle of the engine. When direct injection timing was near 320 CAD, HC and CO emissions were minimized and NOx emission peaked. HC and CO emissions are mainly dependent on the combustion temperature and are reduced by higher combustion temperature [26]. In these injection timings, combustion temperature was somewhat higher than that in other injection timings because of locally richer combustion with higher gc. However, gf was low at these injection timings because of low IMEPgross resulted from fast combustion. Therefore, optimal HCCI injection timing, which could be determined near at 260 CAD, was the best operating condition for DME direct injection HCCI engine based on IMEP, NOx, HC, CO, gc and gf. It was because that injection timing made moderate rich and tardy combustion. In a view of acceptable operating condition, Oakley et al. [27] reported that partial misfire as a minimum acceptable level of combustion efficiency was defined as from 85 % to 90 % of combustion efficiency. In this study, most of direct injection timing conditions were in acceptable operating condition.
3.3. Improving engine performance by exhaust gas recirculation Fig. 6. Combustion phasing variation (CA5, CA50, CA95) as a function of direct injection timing.
temperature condition due to the absorption of latent heat of DME and the late injection timing, which was too late to ignite before TDC (BTDC). Therefore, the negative work from the compression stroke was reduced, and consequently the IMEPgross increased. In this region, the combustion characteristics was mixing controlled combustion like a traditional CI combustion. From the view of HCCI combustion, which uses multipoint autoignition, direct injection timing approximately at 260 CAD was the optimal injection timing based on IMEPgross and NOx emission because the third region was not HCCI region. IMEPgross with direct injection made approximately 22% improvement compared to port injection case. Hydrocarbon (HC), carbon monoxide (CO), combustion efficiency (gc) and thermal efficiency (gf) are presented in Fig. 7 as a function of direct injection timing. Hydrocarbon (HC), carbon monoxide (CO), combustion efficiency (gc) and indicated thermal efficiency (gf) are presented in Fig. 7 as a function of air excess ratio for two injection locations. The calculation of combustion efficiency, gc, was based on the exhaust gas which contained combustible elements, such as CO, hydrogen (H2), and unburned HC, as shown in Eq. (2) [24]. In this calculation, unburned HC was assumed as DME.
0
X xi Q HV i
1
C i gc ð%Þ ¼ B @1 A 100 _ f =ðm _ aþm _ f Q LHV ½m
ð2Þ
In this section, the possibility and the effect of EGR on improving IMEP, emissions and efficiencies were investigated. Fig. 8 shows the change of IMEPgross with optimal direct injection timing as a function of EGR rate. EGR rate was limited by restriction of EGR system. Fig. 9 shows the combustion phasing with optimal direct injection as a function of EGR rate. CA5 and CA50 were retarded and combustion duration became longer as the EGR rate increased due to the increased thermal capacity and decreased oxygen (O2) concentration [17]. EGR could improve the IMEPgross due to tardy combustion. IMEPgross with EGR and direct injection made 18% and 53% improvement, respectively, compared to port injection case and direct injection case without EGR. Hydrocarbon (HC), carbon monoxide (CO), combustion efficiency (gc) and thermal efficiency (gf) are presented in Fig. 10 as a function of EGR rate. The HC and CO emissions increased due to lower combustion temperature resulting from increased EGR rate. Higher EGR rate made combustion temperature decrease. Though gc also decreased with the higher EGR rate due to the lower combustion temperature, gc under this experimental condition was in acceptable range based on the minimum acceptable level of combustion efficiency, which is 90% [27]. The late combustion due to EGR improved the thermal efficiency. It was because of increasing IMEPgross by tardy combustion, as shown in Fig. 9. In a gasoline engine, EGR could improve thermal efficiency because of reduced pumping loss. However, in a diesel engine, EGR could deteriorate the thermal efficiency because of lower combustion efficiency in some case of operating condition.
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Fig. 7. HC, CO, combustion efficiency (gc) and thermal efficiency (gf) variation as a function of direct injection timing.
Fig. 8. The change of IMEPgross with optimal direct injection timing as a function of EGR rate.
Fig. 9. Combustion phasing variation for direct injection case as a function of EGR rate.
4. Conclusions In this study, operating range extension of DME homogeneous charge compression ignition (HCCI) engine, in a view of the gross indicated mean effective pressure (IMEPgross), affected by direct injection and exhaust gas recirculation (EGR) was investigated and combustion characteristics was analyzed with regard to injection location, direct injection timing and EGR rate were changed. The following conclusions were drawn from this study. (1) Direct injection with optimal injection timing and EGR increased the IMEPgross 18% and 53% respectively compared to port injection cases respectively.
(2) Direct injection and EGR increased IMEPgross because those made tardy combustion at lower temperature condition induced by latent heat absorption during evaporation by direct injection, and thermal and dilution effect of EGR. There was optimal direct injection timing for the best IMEPgross in DME HCCI engine and it was determined by IMEPgross and emission results. (3) Direct injection made combustion characteristics change from HCCI to traditional compression ignition (CI) by varing direct injection timing. During the combustion characteristics change from fully homogeneous combustion to mixing controlled like combustion, HCCI combustion with locally
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Fig. 10. HC, CO, combustion efficiency (gc) and thermal efficiency (gf) variation as a function of EGR rate.
rich combustion took place and it was not only to increase IMEPgross but also to decrease hydrocarbon (HC) and carbon monoxide (CO) due to lack of mixing time.
Acknowledgments The authors would like to thank the Korea Institute of Energy Technology Evaluation and Planning (KETEP) for their financial support of this work. References [1] Onishi S, Jo S, Shoda K, Jo P, Kato S. Active thermo-atmosphere combustion (ATAC) – a new combustion process for internal combustion engines; 1979. SAE Paper No. 790501. [2] Noguchi M, Tanaka Y, Tanaka T, Takeuchi Y. A study on gasoline engine combustion by observation of intermediate reaction products during combustion; 1979. SAE Paper No. 790840. [3] Epping K, Aceves S, Bechtold R, Dec J. The potential of HCCI combustion for high efficiency and low emissions; 2002. SAE Paper No. 2002-01-1923. [4] Zhao F, Asumus T, Assanis D, Dec J, Eng J, Najt P. Homogeneous charge compression ignition (HCCI) engines: key research and development issues; PT-94. Warrendale, (PA): Society of Automotive Engineers; 2003. [5] Kitamura T, Ito T, Senda J, Fujimoto H. Extraction of the suppression effects of oxygenated fuels on soot formation using a detailed chemical kinetic model. JSAE Rev 2001;22:139–45. [6] Kaiser E, Yang J, Culp T, Xu N, Maricq M. Homogeneous charge compression ignition engine-out emissions – does flame propagation occur in homogeneous charge compression ignition? Int J Engine Res 2002;3(4):185–95. [7] Semelsberger T, Borup R, Greene H. Dimethyl ether (DME) as an alternative fuel. J Power Sources 2006;156:497–511. [8] Acroumanis C, Bae C, Crookes R, Kinoshita E. The potential of di-methyl ether (DME) as an alternative fuel for compression–ignition engines: a review. Fuel 2008;87:1014–30.
[9] Teng H, McCandless J, Schneyer J. Compression ignition delay (physical + chemical) of dimethyl ether – an alternative fuel for compression-ignition engines. SAE Trans J Fuels Lubricants 2003;112(4):377–89. SAE Paper No. 2003-01-0759. [10] Zhang G, Liu H, Xia X, Yang Q. Study on the injection process of a directinjection diesel engine fuelled with dimethyl ether. Proc Ins Mech Engine Part D: J Auto Eng 2004;218:1341–7. [11] Wakai K, Nishida K, Yoshizaki T, Hiroyasu H. Spray and ignition characteristics of dimethyl ether injected by a DI diesel injector. In: Fourth Intl. Conf. COMODIA. 1998; 98: 537–542. [12] Tabata M, Yamamoto T, Fukube T. Improving NOx and fuel economy for mixture injected si engine with EGR. SAE Trans J Engine 1995;104(3):1221–30. SAE Paper No. 950684. [13] Ladommatos N, Abdelhallm SM, Zhao H, Hu Z. The dilution, chemical, and thermal effects of exhaust gas recirculation on diesel engine emissions – Part 1: Effect of reducing inlet charge oxygen; 1996. SAE Paper No. 961165. [14] Ladommatos N, Abdelhallm SM, Zhao H, Hu Z. The dilution, chemical, and thermal effects of exhaust gas recirculation on diesel engine emissions – Part 2: Effect of carbon dioxide; 1996. SAE Paper No. 961167. [15] Christensen M, Johansson B. Influence of mixture quality on homogeneous charge compression ignition. SAE Trans J Fuels Lubicants 1998;107(4):948–60. SAE Paper No. 982454. [16] Mitchell DL, Pinson JA, Litzinger TA. The effects of simulated EGR via intake air dilution on combustion in an optically accessible DI diesel engine. SAE Trans J Engine 1993;102(3):2313–31. SAE Paper No. 932798. [17] Sahashi W, Azetsu A, Oikawa C. Effects of N2 and CO2 mixing on ignition and combustion in a homogeneous charge compression ignition engine operated on dimethyl ether. Int J Engine Res 2005;6:423–31. [18] Nakano M, Mandokoro Y, Kubo S, Yamazaki S. Effects of exhaust gas recirculation in homogeneous charge compression ignition engines. Int J Engine Res 2000;1:269–79. [19] Qu S, Deng K, Shi L, Cui Y. Effect of direct in-cylinder CO2 injection on HCCI combustion and emission characteristics. Int J Auto Technol 2009;10:529–35. [20] Kanda T, Hakozaki T, Uchimoto T, Hatano J, Kitayama N, Sono H. PCCI operation with fuel injection timing set close to TDC; 2006. SAE Paper No. 2006-01-0920. [21] Douglas R, Kee R, Carberry B. Analysis of in-cylinder pressure data in twostroke engines; 1997. SAE Paper No. 972792. [22] Jung S, Ishida M, Ueki H, Sakaguchi D. Ignition characteristics of methanol and natural-gas in a HCCI engine assisted by DME; 2007. SAE Paper No. 2007-011863.
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[23] Milovanovic N, Chen R. A Review of experimental and simulation studies on controlled auto-ignition combustion; 2001. SAE Paper No. 2001-01-1890. [24] Heywood JB. Internal combustion engine fundamentals. McGraw-Hill; 1988. [25] Yao M, Zheng Z, Zhang B, Chen Z. The effect of PRF fuel octane number on HCCI operation; 2004. SAE Paper No. 2004-01-2992.
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Contents lists available at SciVerse ScienceDirect
Fuel journal homepage: www.elsevier.com/locate/fuel
Improvement of DME HCCI engine combustion by direct injection and EGR Jinyoung Jang a, Youngjae Lee a, Chongpyo Cho a, Youngmin Woo a, Choongsik Bae b,⇑ a b
Korea Institute of Energy Research, 152 Gajeong-ro Yuseong-gu, Daejon 305-343, Republic of Korea Korea Advanced Institute of Science and Technology, 373-1 Kusong-dong, Yusong-gu, Daejon 305-701, Republic of Korea
h i g h l i g h t s Evaluation of the effect of injection location and EGR on performances in DME HCCI engine. In-cylinder direct injection was better than port DME injection engine for DME HCCI engine. Combustion is changed from homogeneous combustion to mixing controlled combustion by varied in-cylinder injection timing. In-cylinder direct injection with EGR made more power than without EGR.
a r t i c l e
i n f o
Article history: Received 22 January 2013 Received in revised form 30 May 2013 Accepted 1 June 2013 Available online 17 June 2013 Keywords: HCCI (homogeneous charge compression ignition) EGR (exhaust gas recirculation) Direct injection Port injection DME (dimethyl ether)
a b s t r a c t To extend the operating range of a homogeneous charge compression ignition (HCCI) engine fuelled with DME, direct injection and exhaust gas recirculation (EGR) were used to investigate their effects on the change of combustion characteristics in this experiment. Single cylinder engine was used with port fuel injector and direct injector. EGR line was connected from exhaust pipe to intake pipe. The experimental result for direct injection was compared with that for port injection. The gross indicated mean effective pressure (IMEPgross), which represented the operation region, for direct injection was higher than that for port injection due to late combustion resulting from lower temperature combustion induced by evaporating latent heat of injected DME in the cylinder. There was optimal direct injection timing for DME HCCI engine because the combustion characteristics were changed from overall lean combustion to locally rich combustion. IMEPgross was increased by EGR. It was because EGR made tardy combustion and increased positive work during expansion stroke. Consequently, direct injection with optimal injection timing and EGR increased the IMEPgross by 22% and 55% compared to port injection case, respectively. Crown Copyright Ó 2013 Published by Elsevier Ltd. All rights reserved.
1. Introduction The homogeneous charge compression ignition (HCCI) engine is a promising concept for future automobile engines and stationary power plants. In HCCI, a premixed air/fuel mixture is inhaled as in conventional spark ignition (SI) engines, which it is auto-ignited as in conventional compression ignition (CI) engines. HCCI uses spontaneous ignition, which is governed primarily by chemical kinetics during the compression process with the premixed air/fuel mixture. Basic characteristics of the HCCI engine, such as combustion stability and no flame propagation, have been studied by previous researchers [1,2]. HCCI engines emit less nitrogen oxide (NOx) than conventional engines due to the low combustion temperature resulting from the high air excess ratio [3]. Particulate matter (PM) emitted from the HCCI engine is also less than conventional ⇑ Corresponding author. Tel.: +82 42 350 3044; fax: +82 42 350 5023. E-mail address: [email protected] (C. Bae).
CI engines due to the lean homogeneous charge mixture [3–5]. However, HCCI engine operation has some disadvantages, such as difficulty in controlling the combustion phasing, much lower maximum load than typical SI and CI engines, high pressure and noise at a low air excess ratio or high load condition, and higher carbon monoxide (CO) and hydrocarbon (HC) emissions than SI and CI engines due to the low combustion temperature [4,6]. Dimethyl ether (DME) is considered an alternative fuel for CI engine [7,8]. DME is a high cetane numbered fuel similar to diesel that does not produce soot due to its molecular structure; DME has no double C–C bonds and contains oxygen atoms which oxidize intermediate soot product. DME is easily evaporated and mixes well with the air so that the rich regions are avoided, leading to soot-free combustion. The ignition delay of DME in a CI engine is shorter than that of diesel under the same operating conditions as the evaporation rate of DME is two or three times faster than diesel [9–11]. Because of the shorter ignition delay, various control schemes to change the auto-ignition timing have been proposed in
0016-2361/$ - see front matter Crown Copyright Ó 2013 Published by Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.fuel.2013.06.001
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Nomenclature C CA5 CA50 CA95 mfuel QLHV QHVi Wi xi ðxCO2 ÞE ðxCO2 ÞI k
gc gf
carbon 5% mass fraction burned point 50% mass fraction burned point 95% mass fraction burned point dimethyl ether quantity lower heating value of the fuel lower heating value of each species indicated work done on the piston mass fraction CO2 concentration in the exhaust gas CO2 concentration in the intake gas air excess ratio combustion efficiency indicated thermal efficiency
Subscripts a air f fuel Abbreviations AC alternating current ABDC after bottom dead center
DME HCCI such as internal and/or external exhaust gas recirculation (EGR), fuel mixing, and additives [12]. EGR is considered to be an effective method for controlling the combustion phasing and reducing exhaust emissions, especially NOx. In SI engines, EGR can reduce NOx emissions without the penalties of increase in other emissions or excessive fuel consumption [12]. The effects of EGR are classified as dilution-, thermal-, and chemical effects [13,14]. The three EGR mechanisms affect the combustion rate, burn duration, combustion temperature, and exhaust emission. In HCCI, thermal and dilution effects are more important than chemical effects of EGR due to the lower combustion temperature. The retarded onset of auto-ignition with EGR for a fixed induction air temperature is due to the thermal effect of the cooler compressed gas temperatures during the compression stroke [15–19]. Oxygen (O2) dilution due to EGR retards the auto-ignition timing [17]. The heated intake charge by EGR advances the auto-ignition timing [19]. There is a potential for improving volumetric efficiency due to fuel evaporation in the cylinder resulting in increased air density with the direct injection during induction stroke. Fuel stratification by direct injection can also improve the HCCI engine combustion with variation of the injection timing [20]. Therefore, direct injection can be one of the ways to improve emissions and power in HCCI engine. In this study, the effect of fuel injection strategy and EGR on an HCCI engine combustion fueled with DME was investigated. Output performance related to power and efficiency and emission were characterized with EGR rate and for the various injection locations in port and cylinder at different injection timings. To find the way to improvement of indicated mean effective pressure, direct injection with the change of injection timing and exhaust gas recirculation were used and combustion characteristics was analyzed.
ATDC BBDC BTDC CAD CI CO CO2 DME DOHC EGR EVC EVO HC HCCI IVC IVO IMEPgross NOx O2 SI PM TDC
after top dead center before bottom dead center before top dead center crank angle degree compression ignition carbon monoxide carbon dioxide dimethyl ether double overhead camshaft exhaust gas recirculation exhaust valve close exhaust valve open hydro carbon homogeneous charge compression ignition intake valve close intake valve open gross indicated mean effective pressure nitrogen oxide oxygen spark ignition particulate matter top dead center
presented in Table 1. The engine was a four-stroke, water-cooled, and single-cylinder double overhead camshaft (DOHC) engine. The engine speed and load were controlled by an alternating current (AC) dynamometer (82 kW, Unico Co.). The DME fuel injector was placed upstream, 30 cm away from the intake port, to form a homogeneous air/fuel mixture and another injector was installed in cylinder to inject the fuel into the cylinder. DME was pressurized to 5 MPa by nitrogen gas. The DME injector for port injection was inclined at 30° to minimize wall wetting and improve air/fuel mixing. A small amount (500 ppm) of lubrication enhancer (Infineum, R655) was added to the DME to protect the injector and fuel feeding system due to lower lubricity of DME. For the exhaust gas recirculation (EGR), EGR line was connected from exhaust pipe to intake pipe, as shown in Fig. 1. The in-cylinder pressure was measured by a piezoelectric pressure transducer (Kistler, 6052b). The exhaust gases were analyzed with a gas analyzer (HORIBA MEXA-1500d) to measure the HC, CO, CO2, NOx and O2 in the exhaust and CO2 in the intake. The experimental data were acquired by a data acquisition system (IO Tech, Wavebook 516 series with 12 bits of resolution, 1 ls/channel sampling rate, and an accuracy of 0.025%). The exhaust gas recirculation (EGR) rate is defined as in Eq. (1) in terms of its mole concentration. It is experimentally determined based on a comparison between the CO2 mole-concentration in the intake gas, ð xCO2 ÞI , and the CO2 concentration in the exhaust gas, ð xCO2 ÞE .
EGR rateð%Þ ¼
ðxCO2 ÞI 100 ðxCO2 ÞE
ð1Þ
2.2. Experimental conditions 2. Experiments 2.1. Experimental apparatus Fig. 1 shows a schematic diagram of the experimental apparatus used in this study. The detailed specifications of the test engine are
To compare the effect of injection location on the DME HCCI combustion, the injection location was changed between the port and in-cylinder. The air excess ratio (k) was changed from 3.1 to 4.8. In-cylinder direct injection timing was varied from 20 to 350 CAD to investigate the effect of injection timing while the injection timing of the port injection was maintained at 20 CAD.
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Fig. 1. Experimental engine set-up.
Table 1 Engine specification. Bore (mm) Stroke (mm) Compression ratio Intake/Exhaust valve duration Intake/Exhaust lift duration (CAD) Valve timing (CAD) IVO (BTDC) IVC (ABDC) EVO (BBDC) EVC (ATDC) DME injection pressure (MPa)
82 93.5 13 228/228 8.5/8.4 1 39 42 6 5
The EGR rate was changed from 0% to 60% to change the combustion phasing and improve the gross indicated mean effective pressure (IMEPgross). In-cylinder and intake/exhaust pressure data for one hundred consecutive engine cycles were used for the analysis of combustion and IMEPgross [21].
work. It is reason why IMEPgross of direct injection case is higher than that of port injection case. IMEPgross for two injection locations is shown in Fig. 3a, as a function of air excess ratio (k). In both port injection and in-cylinder direct injection, IMEPgross would be reaching their highest value as injected DME quantity was increased. The ignition timing was advanced and/or the burn duration was reduced during the compression stroke as DME quantity increased, resulting in increased negative work. After this end, the IMEPgross did not increase further. In the case of port injection, maximum pressure rate at highest IMEPgross was 0.8 MPa which was higher than criteria for load limit used by Jung et al.[22]. In the case of direct injection, IMEPgross was improved approximately by 3% compared to port injection at each air excess ratio condition due to the elongated ignition delay resulted from slow
3. Results and discussion 3.1. Comparing between port and direct injection To understand the improvement of performance by direct injection, firstly, IMEPgross and auto-ignition timing were compared with those by port injection. The injection quantity of DME was varied and the injection timing was maintained at 20 crank angle degree (CAD) after top dead center (TDC) of intake stroke. Fig. 2 shows comparing results of in-cylinder pressure and heat release rate profile for port and direct injection case. The start of the heat release with direct injection was more delayed than with the port injection while the burn duration was similar at the same operating conditions. It usually led to less negative compression
Fig. 2. In-cylinder pressure and heat release rate profile comparison between port and direct injection.
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reaction rate, which was mainly affected by lower mixture temperature. The injected fuel at the port produced a homogeneous mixture during the induction stroke without additional temperature drop in the cylinder. However, there was a temperature drop due to the latent heat of liquid DME when it was injected into the cylinder and evaporated. In Fig. 3b, auto-ignition timing of direct injection was more retarded and negative work of the direct injection decreased more than those of port injection due to the temperature drop with the direct injection of DME fuel [23]. Fig. 3b shows the CA5 (5% mass fraction burned) with respect to injection location and air excess ratio. Combustion phasing was presented with CA5, CA50 and CA95, which are 5%, 50% and 95% mass fraction burn points, respectively. The 5% burn point of the combustion is closely related to the auto-ignition timing. Since 5% burn point is given in CAD, if CA5 is earlier than 360 CAD, it occurs before TDC. Ignition timing for direct injection was later than that for port injection, in spite of same injection timing. In the case of direct injection, the advance of ignition timing was not significant by increasing DME injection quantity.
Through previous section, it was found that direct injection was better way to improve IMEPgross on DME HCCI engine operation than port injection. Therefore, in this section, the effect of DME direct injection timing on HCCI engine was investigated to seek the optimal injection timing. Fig. 4 shows the variation of IMEPgross and NOx emission as a function of injection timing and also compares the IMEPgross of port injection case. The change of IMEPgross had three regions as a function of injection timing. The IMEPgross increased in the first region approximately from 20 to 260 CAD. The IMEPgross decreased within
the second region approximately from 260 to 320 CAD. The third region was later than 320 CAD and the IMEPgross was augmented. The change of IMEPgross with injection timing was explained by combustion phasing, as shown in Fig. 5 and 6. Fig. 5 shows the in-cylinder pressure and heat release rate profiles as a function of direct injection timing. And, Fig. 6 shows the combustion phasing as a function of direct injection timing. Within the first region approximately from 20 to 260 CAD, the CA5 and CA50 was retarded due to the cooling effect of evaporated DME and late injection timing. Retarded CA50 not only decreased negative work during compression but also increased positive work during expansion. Therefore, late direct injection made IMEPgross increase. However, during the second region approximately from 260 to 320 CAD, IMEPgross decreased as the burn duration from CA5 to CA95 was shorter than the first region and the combustion process ended before the top dead center (TDC) resulting in increased negative work and decreased positive work, in spite of the retarded onset of auto-ignition. The combustion characteristic was changed from relatively leaner to locally-richer combustion due to a change of mixture homogeneity as direct injection timing was delayed. The locally-rich combustion took place by a lack of mixing time resulting from the late injection timing with direct injection. The locally richer combustion by direct injection timing could be explained using NOx emission result as a function of injection timing, as shown in Fig. 3b. NOx emission was negligible with early injection and increased with late injection near the TDC. High temperature and oxygen concentrations resulted in high NOx formation rates [24]. NOx emission was recognized when injection timing was later than 240 CAD and the peak value of NOx emission occurred near 320 CAD due to locally-rich and higher combustion temperatures. During the third region approximately from 320 to 350 CAD, the onset of auto-ignition was more retarded than TDC at lower
Fig. 3. IMEPgross and CA5 comparison between port and direct injection as a function of injection location and air excess ratio (k).
Fig. 4. IMEPgross and NOx emission variation as a function of direct injection timing.
3.2. Improving engine performance by direct-injection timing optimization
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where xi is the mass fraction of CO, H2, and unburned HC, QHVi represents the lower heating values of these species, and the subscripts f and a denote fuel and air. The indicated thermal efficiency, gf, was defined as the ratio of the work done on the piston during the compression and expansion strokes to the heat supplied by the fuel, as shown in the following equation [25].
gf ð%Þ ¼
Fig. 5. In-cylinder pressure and heat release rate profile as a function of direct injection timing.
Wi 100 mfuel Q LHV
ð3Þ
where mfuel is the DME mass, QLHV is the lower heating value of the fuel, and Wi is the indicated work done on the piston calculated using in-cylinder pressure data over the operating cycle of the engine. When direct injection timing was near 320 CAD, HC and CO emissions were minimized and NOx emission peaked. HC and CO emissions are mainly dependent on the combustion temperature and are reduced by higher combustion temperature [26]. In these injection timings, combustion temperature was somewhat higher than that in other injection timings because of locally richer combustion with higher gc. However, gf was low at these injection timings because of low IMEPgross resulted from fast combustion. Therefore, optimal HCCI injection timing, which could be determined near at 260 CAD, was the best operating condition for DME direct injection HCCI engine based on IMEP, NOx, HC, CO, gc and gf. It was because that injection timing made moderate rich and tardy combustion. In a view of acceptable operating condition, Oakley et al. [27] reported that partial misfire as a minimum acceptable level of combustion efficiency was defined as from 85 % to 90 % of combustion efficiency. In this study, most of direct injection timing conditions were in acceptable operating condition.
3.3. Improving engine performance by exhaust gas recirculation Fig. 6. Combustion phasing variation (CA5, CA50, CA95) as a function of direct injection timing.
temperature condition due to the absorption of latent heat of DME and the late injection timing, which was too late to ignite before TDC (BTDC). Therefore, the negative work from the compression stroke was reduced, and consequently the IMEPgross increased. In this region, the combustion characteristics was mixing controlled combustion like a traditional CI combustion. From the view of HCCI combustion, which uses multipoint autoignition, direct injection timing approximately at 260 CAD was the optimal injection timing based on IMEPgross and NOx emission because the third region was not HCCI region. IMEPgross with direct injection made approximately 22% improvement compared to port injection case. Hydrocarbon (HC), carbon monoxide (CO), combustion efficiency (gc) and thermal efficiency (gf) are presented in Fig. 7 as a function of direct injection timing. Hydrocarbon (HC), carbon monoxide (CO), combustion efficiency (gc) and indicated thermal efficiency (gf) are presented in Fig. 7 as a function of air excess ratio for two injection locations. The calculation of combustion efficiency, gc, was based on the exhaust gas which contained combustible elements, such as CO, hydrogen (H2), and unburned HC, as shown in Eq. (2) [24]. In this calculation, unburned HC was assumed as DME.
0
X xi Q HV i
1
C i gc ð%Þ ¼ B @1 A 100 _ f =ðm _ aþm _ f Q LHV ½m
ð2Þ
In this section, the possibility and the effect of EGR on improving IMEP, emissions and efficiencies were investigated. Fig. 8 shows the change of IMEPgross with optimal direct injection timing as a function of EGR rate. EGR rate was limited by restriction of EGR system. Fig. 9 shows the combustion phasing with optimal direct injection as a function of EGR rate. CA5 and CA50 were retarded and combustion duration became longer as the EGR rate increased due to the increased thermal capacity and decreased oxygen (O2) concentration [17]. EGR could improve the IMEPgross due to tardy combustion. IMEPgross with EGR and direct injection made 18% and 53% improvement, respectively, compared to port injection case and direct injection case without EGR. Hydrocarbon (HC), carbon monoxide (CO), combustion efficiency (gc) and thermal efficiency (gf) are presented in Fig. 10 as a function of EGR rate. The HC and CO emissions increased due to lower combustion temperature resulting from increased EGR rate. Higher EGR rate made combustion temperature decrease. Though gc also decreased with the higher EGR rate due to the lower combustion temperature, gc under this experimental condition was in acceptable range based on the minimum acceptable level of combustion efficiency, which is 90% [27]. The late combustion due to EGR improved the thermal efficiency. It was because of increasing IMEPgross by tardy combustion, as shown in Fig. 9. In a gasoline engine, EGR could improve thermal efficiency because of reduced pumping loss. However, in a diesel engine, EGR could deteriorate the thermal efficiency because of lower combustion efficiency in some case of operating condition.
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Fig. 7. HC, CO, combustion efficiency (gc) and thermal efficiency (gf) variation as a function of direct injection timing.
Fig. 8. The change of IMEPgross with optimal direct injection timing as a function of EGR rate.
Fig. 9. Combustion phasing variation for direct injection case as a function of EGR rate.
4. Conclusions In this study, operating range extension of DME homogeneous charge compression ignition (HCCI) engine, in a view of the gross indicated mean effective pressure (IMEPgross), affected by direct injection and exhaust gas recirculation (EGR) was investigated and combustion characteristics was analyzed with regard to injection location, direct injection timing and EGR rate were changed. The following conclusions were drawn from this study. (1) Direct injection with optimal injection timing and EGR increased the IMEPgross 18% and 53% respectively compared to port injection cases respectively.
(2) Direct injection and EGR increased IMEPgross because those made tardy combustion at lower temperature condition induced by latent heat absorption during evaporation by direct injection, and thermal and dilution effect of EGR. There was optimal direct injection timing for the best IMEPgross in DME HCCI engine and it was determined by IMEPgross and emission results. (3) Direct injection made combustion characteristics change from HCCI to traditional compression ignition (CI) by varing direct injection timing. During the combustion characteristics change from fully homogeneous combustion to mixing controlled like combustion, HCCI combustion with locally
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Fig. 10. HC, CO, combustion efficiency (gc) and thermal efficiency (gf) variation as a function of EGR rate.
rich combustion took place and it was not only to increase IMEPgross but also to decrease hydrocarbon (HC) and carbon monoxide (CO) due to lack of mixing time.
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