Energy Fuels 2010, 24, 1655–1667 Published on Web 02/19/2010
: DOI:10.1021/ef9015153
Numerical Investigation into the Formation of CO and Oxygenated and Nonoxygenated Hydrocarbon Emissions from Isooctane- and Ethanol-Fueled HCCI Engines Neofytos P. Komninos* and Constantine D. Rakopoulos Internal Combustion Engines Laboratory, Thermal Engineering Department, School of Mechanical Engineering, National Technical University of Athens, 9 Heroon Polytechniou St., Zografou Campus, 15780 Athens, Greece Received December 11, 2009. Revised Manuscript Received January 28, 2010
The present study focuses on the investigation of the formation of CO and unburned oxygenated and nonoxygenated hydrocarbon emissions from HCCI engines fueled with neat ethanol and neat isooctane. This is achieved with the use of a multizone model, which describes the essential features of HCCI combustion, that is, heat and mass transfer within the combustion chamber, both of which are modeled using phenomenological submodels. These mechanisms affect the formation of the main HCCI engine pollutants, namely, unburned hydrocarbons and carbon monoxide. Combustion is simulated using chemical kinetics coupled to oxidation mechanisms for isooctane and ethanol. These mechanisms also describe the decomposition of the original fuel into intermediate hydrocarbons and carbon monoxide. A validation of the model for both fuels is given for various load cases. In the numerical investigation, the formation of CO is described for the corresponding experimental cases and the essential features of the transition from CO production due to bulk quenching and to CO production due to postcombustion partial HC oxidation are shown. Additionally, the formation of HC emissions is described including both oxygenated and nonoxygenated compounds. This distinction was found to be necessary since both fuels include oxygenated species in the exhaust gases, the relative amount of which depends on load conditions and the fuel used. The fraction of oxygenated compounds to total unburned HC is high for ethanol at all loads, primarily due to the presence of ethanol, acetaldehyde and formaldehyde, in descending order of importance. The relative proportion of oxygenates in total unburned HC in the case of isooctane was found to depend on load. These findings raised questions regarding the assessment of unburned hydrocarbon emissions using conventional measuring devices, such as the FID. For this reason the relative error in the FID measurement was estimated, using the simulated HC composition results and the FID relative response of each of the species constituting the HC.
Among the various methods for the ignition and combustion rate control is the use of alternative fuels, either as neat fuels or additives.2-8 Such fuels include PRF, methane, reformed gas, hydrogen, and biofuels such as methanol, ethanol, biodiesel, diethyl ether, etc. Ethanol has been used as a fuel in HCCI engines, either as neat ethanol9,10 or as a supplement.11,12 Current research also focuses on using wet ethanol in HCCI combustion since its production is cheaper
1. Introduction Homogeneous charge compression ignition (HCCI) engines have received growing attention in the past decade. The motivation for this interest lies in the beneficial characteristics of HCCI engines, that is, mainly low NOx emissions due to the relatively poor mixtures used, absence of particulates in the exhaust gases due to the premixed nature of combustion, and the potential for high thermal efficiency due to the relatively short combustion duration.1 Despite these beneficial characteristics, HCCI combustion is not devoid at adverse effects. Specifically, it has been found that carbon monoxide (CO) and unburned hydrocarbon (HC) emissions are at high levels,1 and any gain in the thermal efficiency presupposes proper ignition timing. The latter is an issue of great importance since in HCCI engines there are, by definition, no external means of initiating combustion. Ignition depends on the oxidation chemistry of the fuel, which also governs the combustion rate. An excessive combustion rate at high loads and fueling rates can lead to abnormally high pressure rise rates (knock). These issues are the subject of current research on HCCI engines.
(2) Yao, M.; Huang, C.; Zheng, Z. Energy Fuels 2007, 21, 812–821. (3) Yap, D.; Megaritis, A.; Wyszynski, M. L. Energy Fuels 2004, 18, 1315–1323. (4) Shibata, G.; Oyama, K.; Urushihara, T.; Nakano, T. The effect of fuel properties on low and high temperature heat release and resulting performance of an HCCI engine. SAE Paper No. 2004-01-0553; 2004. (5) Maurya, R. K.; Agarwal, A. K. Proc. Inst. Mech. Eng., Part D, J. Automob. Eng. 2009, 223, 1445–1458. (6) Yao, M.; Chen, Z.; Zheng, Z.; Zhang, B.; Xing, Y. Fuel 2006, 85, 2046–2056. (7) Yap, D.; Peucheret, S. M.; Megaritis, A.; Wyszynski, M. L.; Xu, H. Int. J. Hydrogen Energy 2006, 31, 587–595. (8) Tsolakis, A.; Megaritis, A.; Yap, D. Energy 2008, 33, 462–470. (9) Christensen, M.; Einewall, P.; Johansson, B. Homogeneous charge compression ignition (HCCI) using isooctane, ethanol and natural gas - A comparison with spark-ignition operation. SAE Paper No. 972874; 1997. (10) Christensen, M.; Johansson, B.; Amneus, P. J. H.; Mauss, , F. Supercharged homogeneous charge compression ignition. SAE Paper No. 980787; 1998. (11) L€ u, X.; Ji, L.; Zu, L.; Hou, Y.; Huang, C.; Huang, Z. Combust. Flame 2007, 149, 261–270. (12) L€ u, X.; Hou, Y.; Zu, L.; Huang, Z. Fuel 2006, 85, 2622–2631.
*To whom correspondence should be addressed. Telephone: þþ30 210 772 1710. E-mail:
[email protected]. (1) Yao, M.; Zheng, Z.; Liu, H. Prog. Energy. Combust. Sci. 2009, 35, 398–437. r 2010 American Chemical Society
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: DOI:10.1021/ef9015153
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and energy-saving. In this perspective, HCCI research can incorporate results obtained from another great field of research, namely, the use of biofuels in internal combustion engines. Biofuel research has been motivated mainly due to both socioeconomics and environmental effects.14-19 The production of biofuels and the gradual replacement of fossil fuels decreases the energy dependence from fossil fuels and supports the agricultural population.16,20 Moreover, biofuels have the potential to reduce certain pollutant emissions such as greenhouse gases due to their renewable character and particulates due to their oxygenated chemical structure.16 For these reasons, biofuels have been used in compression ignition (CI)16-19 and spark ignition (SI) engines21-23 as neat fuels or supplements, as well as in HCCI engines as already mentioned. Ethanol is a biofuel that has been studied under HCCI conditions. Experimental comparative studies of ethanol, natural gas, and isooctane HCCI combustion have been conducted to assess the general features of HCCI combustion when using these fuels.9,10 The focus was to assess the global characteristics of HCCI combustion relative to spark ignition operation and to determine the attainable air-fuel equivalence ratio, the preheating requirement for the proper ignition timing of the fuels and the effect of supercharging on HCCI combustion. The emphasis was on the features of HCCI combustion that are present regardless of fuel, such as the increase in HC emissions, the significant reduction of NOx emissions, the increased efficiency, and the increase of attainable indicated mean effective pressure (imep) when using supercharging. Gnanam24 focused on the effect of fuel;neat ethanol, neat isooctane, and ethanol/isooctane blends;on the produced imep, the thermal efficiency, and on NOx, HC, and CO emissions. Other studies focused on the characteristics of ethanol as neat fuel or additive in a HCCI engine with residual gas trapping.3,25 Yap et al. experimented with the initial temperature and inlet valve timing to extend the operating range of ethanol HCCI combustion with residual gas trapping.3 Xie et al. used gasoline-ethanol or gasolinemethanol blends to study the effect of mixture composition on the in-cylinder temperature, ignition timing, NOx emissions, and other parameters.25 A study, which focused on the formation of emissions from HCCI engines running on a range of fuels;among which are isooctane and ethanol;was the
one by Lemel et al. In their study, these researchers showed that formaldehyde existed in the exhaust gases of all fuels tested, especially the ones with significant low temperature heat release. These results confirmed and extended the earlier findings of Dec and Sj€ oberg, who used single-zone modeling to study low load isooctane HCCI combustion at the limiting cases of incomplete bulk gas reactions.27 The present study focuses on the investigation of the formation of CO and unburned oxygenated and nonoxygenated hydrocarbon emissions from HCCI engines fueled with neat ethanol and neat isooctane. A rudimental validation of the multizone model against experimental cases is first presented. These cases refer to neat ethanol- and neat isooctanefueled HCCI engines. The validation includes pressure traces, heat release rates, and specific HC and CO emissions. The model used for the simulation is a multizone model incorporating heat and mass transfer models for the description of the corresponding phenomena, and chemical kinetics for the simulation of combustion and the determination of species concentrations including HC and CO emissions. Subsequently, an analysis is presented regarding the formation of CO emissions as predicted by the multizone model, for some of the experimental cases. The same analysis is presented for the formation of HC emissions, in which a distinction is made between oxygenated and nonoxygenated species. The predicted composition of the unburned HC emissions;including oxygenated and nonoxygenated species;is also given for various load cases. The estimation of oxygenated unburned hydrocarbon (OHC) emissions is significant, since they cannot be easily measured in actual cases even when sophisticated measuring devices are used.26 Moreover, these oxygenated compounds induce errors in the estimation of total HC using conventional HC measuring devices such as the FID.28,29 For this reason, the simulated HC data are used to estimate the relative error in the detection of the predicted HC emissions, assuming they were actually introduced in a FID analyzer. 2. Model Description In this section a brief outline of the multizone model is provided, since its development has been presented in detail in previous reports.30-32 2.1. Zone Configuration. As shown in Figure 1, the combustion chamber is divided into three different types of zones, that is, the core zone, the outer zones, and the crevice zone. The core zone is a cylinder, each of the outer zones is a cylindrical annulus, and the crevice zone lies beneath the outmost zone. The crevice zone is a fraction of the clearance volume and represents the crevice regions that communicate directly with the combustion chamber, that is, the region above and behind the first compression ring, the head gasket crevice, etc.
(13) Mack, H. J.; Aceves, S. M.; Dibble, R. W. Energy 2009, 34, 782– 787. (14) Rakopoulos, C. D.; Antonopoulos, K. A.; Rakopoulos, D. C.; Hountalas, D. T. Energy Convers. Manage. 2008, 49, 625–643. (15) Rakopoulos, C. D.; Antonopoulos, K. A.; Rakopoulos, D. C. Energy Convers. Manage. 2007, 48, 1881–1901. (16) Rakopoulos, D. C.; Rakopoulos, C. D.; Kakaras, E. C.; Giakoumis, E. G. Energy Convers. Manage. 2008, 49, 3155–3162. (17) Rakopoulos, C. D.; Rakopoulos, D. C.; Hountalas, D. T.; Giakoumis, E. G.; Andritsakis, E. C. Fuel 2008, 87, 147–157. (18) Rakopoulos, D. C.; Rakopoulos, C. D.; Giakoumis, E. G.; Papagiannakis, R. G.; Kyritsis, D. C. Fuel 2008, 87, 1478–1491. (19) Rakopoulos, C. D.; Antonopoulos, K. A.; Rakopoulos, D. C. Energy 2007, 32, 1791–1808. (20) Yan, J.; Lin, T. Appl. Energy 2009, 86, S1–S10. (21) Koc, M.; Sekmen, Y.; Topgul, T.; Yucesu, H. S. Renew. Energy 2009, 34, 2101–2106. (22) Pang, X.; Mua, Y.; Yuan, J.; Hea, H. Atmos. Environ. 2008, 42, 1349–1358. (23) Rakopoulos, C. D.; Michos, C. N. Energy Convers. Manage. 2008, 49, 2924–2938. (24) Gnanam, G.; Sobiesiak, A.; Reader, G.; Zhang, C. An HCCI engine fueled with iso-octane and ethanol. SAE Paper No. 2006-01-3246; 2006. (25) Xie, H.; Wei, Z.; He, B.; Zhao, H. Comparison of HCCI combustion respectively fueled with gasoline, ethanol and methanol through the trapped residual gas strategy. SAE Paper No. 2006-01-0635; 2006.
(26) Lemel, M.; Hultqvist, A.; Vressner, A.; Nordgren, H.; Persson, H.; Johansson, B. Quantification of the formaldehyde emissions from different HCCI engines running on a range of fuels. SAE Paper No. 2005-01-3724; 2005. (27) Dec, J. E.; Sj€ oberg, M. A parametric study of HCCI combustion - the sources of emissions at low loads and the effects of GDI fuel injection. SAE Paper No. 2003-01-0752; 2003. (28) Schofeld, K. Prog. Energ. Combust. Sci. 2008, 34, 330–350. (29) Cheng, W.K .; Summers, T.; Collings, N. Prog. Energ. Combust. Sci. 1998, 24, 89–124. (30) Komninos, N. P.; Hountalas, D. T. Energy Convers. Manage. 2008, 49, 2530–2537. (31) Komninos, N. P. Appl. Energy 2009, 86, 2141–2151. (32) Komninos, N. P.; Hountalas, D. T.; Rakopoulos, D. C. Energy Convers. Manage. 2007, 48, 2934–2941.
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DT Drn
Komninos and Rakopoulos
= rn ¼0
T2 -Tw , in all other cases t2 =2
ð3bÞ
The heat transfer submodel also includes heat transfer between zones. This is incorporated into the multizone model by considering heat transfer via a mechanism similar to conduction, that is, the heat flux between neighboring zones is based on their temperature difference and mean distance, so that: DT : ð4Þ q ¼ -ktot Drn The amount of heat exchanged at the interface of two neighboring zones is calculated by multiplying the heat flux by the interface area and the time step used. The net heat exchanged for each zone Qi is then the algebraic sum of the heat exchanged with its outer and inner zone. For the determination of the total conductivity in eq 4, the approach of Yang and Martin34 is followed: ð5Þ ktot ¼ kl þ kt
Figure 1. Schematic of the multizone model zones configuration.
The volume of the remaining zones varies throughout the engine cycle and it is calculated according to their thickness. As seen in Figure 1, the thickness of each zone is the distance between its inner and its outer boundary. Since the core zone has only an outer boundary, its thickness is defined as half its height. As the piston head moves during compression and expansion, the distance of the bottom boundary of each zone from the piston head is held constant. Therefore, the thickness of all outer zones, which is considered the same in the x- and y-direction, is constant during the cycle. On the contrary, the thickness of the core zone is altered since its height decreases during compression and increases during expansion. At top dead center (TDC) position, the sum of all zone thicknesses must equal half the cylinder clearance height: z X Smin ð1Þ ti jTDC ¼ 2 i ¼2
The ratio of turbulent to laminar conductivity is calculated using the following formula: kt Prl μt ¼ ð6Þ kl Prt μl The formula presented above presupposes swirl dominated flows and it is used in the absence of other data. The viscosity ratio of eq 6 is calculated from the formula: μt ¼ Krn þ ½1 -expð -2RKrn þ Þ ð7Þ μl where the dimensionless normal distance from the wall rnþ, which is effectively a Reynolds number, is given as: Z u rn F drn ð8Þ rn þ ¼ μw 0
2.2. Heat Transfer. Heat is transferred from the outmost zone and the crevices to the combustion chamber wall as well as between neighboring zones. The crevices are considered to rapidly assume the wall temperature and preserve this value for the remaining part of the closed cycle. For the estimation of the wall heat flux q_ w, the no-slip boundary condition33 is applied to the gas in contact with the wall. According to the no-slip condition, the fluid in contact with the combustion chamber wall assumes the velocity of the boundary (wall) and thus it is considered stationary. Consequently, heat is transferred through this thin fluid layer only via conduction. Therefore, the wall heat flux is estimated by: DT : ðwall boundary conditionÞ ð2Þ qw ¼ -kw Drn
where κ = 0.41 is the von Karman constant, R = 0.06, and rn is the normal distance from the wall. In the multizone model, rn denotes the normal distance of the combustion chamber wall from the interface between neighboring zones. This is found by adding all thicknesses of the zones located between the wall and the interface of interest and is used for the determination of the conductivity at the interface via eqs 5-8. The characteristic velocity u* is considered to be proportional to the engine speed. This proportionality constant was taken equal to 0.06, for the purposes of the present study, and it was chosen in a way to match mainly the compression and peak combustion pressures. The Yang and Martin approach has been used in the past by other authors35 for the estimation of heat losses in HCCI engines via CFD modeling, and for the estimation of wall heat transfer in premixed charge engine combustion.36 More recently, the Yang and Martin formulas for turbulent conductivity and the adaptation to HCCI multizone modeling, as presented herein and in previous studies by the authors,31
rn ¼0
where kw is the conductivity of the charge at the cylinder wall temperature, rn is the normal distance from the wall, and DT=Drn jrn ¼0 the temperature gradient at the wall (i.e., at rn = 0). The temperature gradient at the wall is approximated numerically by the relations: -1 DT T2 -Tw 2 T3 -T2 = , when Drn t2 =2 ðt2 þ t3 Þ=2
(34) Yang, J.; Martin, J. K. Trans. ASME, J. Heat Transfer 1989, 111, 619–624. (35) Kong, S.; Ayoub, N.; Reitz, R. D. Modeling combustion in compression ignition homogeneous charge engines. SAE Paper No. 920512; 1992. (36) Reitz, R. D. Assessment of wall heat transfer models for premixed-charge engine combustion computations. SAE Paper No. 910267; 1991.
rn ¼0
Tw < T2 < T3 and Tw < T
ð3aÞ
(33) Mills, A. F. Heat and Mass Transfer, 1st ed; Irwin: Chicago, 1995.
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has been adopted by other researches in their HCCI multizone models.37-39 2.3. Mass Transfer. Mass is transferred between zones to maintain the pressure uniform inside the combustion chamber. The mass transferred between zones is calculated using the ideal gas equation of state along with the conservation of mass within the combustion chamber. The assumption of uniform pressure and the use of the ideal gas equation lead to the equations: mi ð9Þ Pcyl Vi ¼ Ru Ti , i ¼ 1, z Mi
Table 1. NOx Formation Reactions
Solving eq 9 for the mass of each zone and summing up, we end up with: z z X X Pcyl Vi Mi mi ¼ mcyl ¼ ð10Þ Ru Ti i ¼1 i ¼1
taken into account in the model during convergence at each CA step. The mass transfer procedure presented herein has been adopted by other researches for the simulation of mass transfer in a relatively new combustion concept, namely the porous medium engine.40 2.4. Combustion and Pollutants Formation. Combustion is described using the appropriate set of chemical reactions. For isooctane (i-C8H18) the oxidation mechanism used is that of Golovitchev, created at the Chalmers University, Sweden.41 This mechanism consists of 84 species and 412 reactions, including NOx formation reactions. For ethanol (C2H5OH), a reduced mechanism was used, namely that of R€ ohl and Peters, created at the Institute of Combustion Technology of Aachen University, Germany.42 This reduced mechanism includes 38 species and 228 reactions, and it is ohl and Peters’ ethanol meavailable online.43 Since the R€ chanism does not provide for NOx reactions, it has been augmented by the NOx formation reactions included in the isooctane mechanism. These NOx reactions and the relevant species are presented in Table 1. As regards the computational procedure, the rate of production (or destruction) of each species in each zone is calculated and the set of differential equations obtained is solved by using the Chemkin44,45 libraries, in order to determine the variation of mixture composition for each zone i: ω_ j Mj dYj ð15Þ ¼ , j ¼ 1, j and i ¼ 1, z dt F
Equation 10 relates the cylinder pressure to the total charge mass, which remains constant at any instant since blow-by is not taken into account. Therefore, the mean cylinder pressure at each crank angle (CA) step is calculated from: mcyl Ru ð11Þ Pcyl ¼ z P Vi Mi i ¼1 Ti Once the uniform pressure is calculated, the mass of each zone is determined using the ideal gas equation of state: Vi Mi , i ¼ 1, z ð12Þ mi ¼ Pcyl Ru T i The mass change of each zone during the crank angle (CA) step is the difference between the zone mass calculated by eq 12 and the zone mass at the previous CA step: CA , i ¼ 1, z Δmi ¼ mi -mprevious i
ð13Þ
This mass change of each zone, which is thermodynamically calculated by eqs 12 and 13, equals the net mass flow from its neighboring zones. Taking into account the configuration of the zones, mass flows only between neighboring zones: flow Δmi ¼ mflow i -1 f i -mi f i þ 1
N þ NO = N2 þ O N þ O2 = NO þ O N þ OH = NO þ H NO þ HO2 = NO2 þ OH NO2 þ O = NO þ O2 NO2 þ H = NO þ OH NO þ O þ M = NO2 þ M N þ CO2 = NO þ CO N2O þ O = NO þ NO N2O þ O = N2 þ O2 N2O þ H = N2 þ OH N2O þ M = N2 þ O þ M N2O þ OH = N2 þ HO2
1 2 3 4 5 6 7 8 9 10 11 12 13
ð14Þ
i
i
· j is the where Yj is the mass fraction of species j in zone i, ω molar rate of production of species j, Mj is the molar mass of species j, F is the density of zone i, and z is the number of zones. Since mass transfer is taken into account in the present multizone model, the final amount of each chemical species in each zone;including CO, HC, and NOx;is determined by its net formation rate and by the net amount of the species transferred to the zone via mass transfer. The exchange of species between zones and, especially, the transfer of combustion products in and out of the crevices play an important role in the formation of CO and HC emissions.
The mass flow between zones can be calculated by successively applying eq 14 for all zones starting from zone 1 (crevice), for which the mass change Δm1 is equal to mflow 1f2, that is, the mass flowing through the interface of zones 1 and 2. The latter holds true, since blow-by is not considered in the model. The transfer of species is calculated based on the assumption that the mass flowing from a zone to its neighboring one has the chemical composition of the zone from which it originates. This mixing owed to mass transfer affects the final chemical composition of the zones and the average zone molar mass Mi used in eqs 11 and 12. Both of these effects are
(40) Liu, H.; Xie, M.; Wu, D. Appl. Therm. Eng. 2009, 29, 3189–3197. (41) http://www.tfd.chalmers.se/∼valeri/MECH.html, 2003. (42) R€ ohl, O.; Peters, N. A Reduced Mechanism for Ethanol Oxidation. 4th European Combustion Meeting, Vienna University of Technology, Vienna, Austria, April 14-17, 2009. (43) http://www.itv.rwth-aachen.de/index.php?id=16&L=5, 2009. (44) Kee, R. J.; Rupley, F. M.; Miller, J. A. CHEMKIN II: A FORTRAN chemical kinetics package for the analysis of gas-phase chemical kinetics. Report No. SAND89-8009B; Sandia National Laboratories: Livermore, CA, 1989. (45) http://www.reactiondesign.com, 2009.
(37) Jia, M.; Xie, M.; Peng, Z. A comparative study of multi-zone combustion models for HCCI engines. SAE Paper No. 2008-01-0064; 2008. (38) Kongsereeparp, P.; Checkel, M. D. Investigating the effects of reformed fuel blending in a methane- or n-heptane-HCCI engine using a multi-zone model. SAE Paper No. 2007-01-0205; 2007. (39) Kongsereeparp, P.; Checkel, M. D. Novel method of setting initial conditions for multi-zone HCCI combustion modeling. SAE Paper No. 2007-01-0674; 2007.
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2.5. The First Law of Thermodynamics Applied to the Multizone Model. Each zone of the multizone model is considered as an open thermodynamic system exchanging heat, mass, species, and enthalpy based on the configuration shown in Figure 1. All interactions between zones occur through the interface of neighboring zones. Heat transfer to the combustion chamber wall is transferred only through the crevice zone (zone 1), which is considered to assume the wall temperature, and the outmost zone (zone 2), which is in contact with the wall. The volume of each zone changes due to the piston motion and can be directly calculated based on the zone thickness and the configuration shown in Figure 1. This volume change of each zone induces mass transfer between neighboring zones to maintain uniform pressure within the combustion chamber. Moreover, mass exchange can also be the result of combustion in a zone or its neighboring ones. Since blow-by is neglected, the conservation of mass principle requires that at any instant the total mass within the combustion chamber remains unaltered. Treating each zone as an open thermodynamic system, the First Law of Thermodynamics takes the following form: dUi ¼ Qi -Pcyl dVi þ
J X j ¼1
flow, in
nj
hj -
J X
flow, out
nj
Table 2. Engine Main Design Data compression ratio bore (mm) stroke (mm) connecting rod (mm) EVO (deg bBDC) EVC (deg bTDC) IVO (deg aTDC) IVC (deg aBDC)
21:1 120.65 140 260 39 10 5 13
Table 3. Engine Operating Conditions fuel engine speed (rpm) fuel-air equivalence ratio φ inlet conditions air inlet temperature Tin (°C) residual gas fraction (est.) (%)
C2H5OH or i-C8H18 994 varied naturally aspirated 120 (preheated air) 3.5
Table 4. FID Relative Response for Selected Hydrocarbons
hj , i ¼ 1, z
j ¼1
ð16Þ
species
carbon atoms in molecule
C2H5OH CH2O CH3CHO CH4 C2H4 i-C8H18 i-C4H8 C3H6 C4H6
2 1 2 1 2 8 4 3 4
a
FID relative response (RR) 1.59a 0a 1 (est) 1b 1.99a 7.98b 4 (est) 2.88b 4a
From ref 28. b From ref 29.
its stoichiometric value:
where dUi is the change in the internal energy of the gas in zone i, nj are the kmol of species j transferred in or out of zone i, and hj is the corresponding molar specific enthalpy of species j. Equation 16 essentially states that the change in the internal energy of the gas in each zone depends upon the net heat gained Qi by zone i, the work produced or consumed Pcyl dVi, and the enthalpy inflow and outflow from the neighboring zones. The enthalpy flow between zones represents energy transfer due to bulk motion and corresponds to the convective heat transfer term. As soon as the mass flowing from each zone to the neighboring zones is determined, the species transfer and the enthalpy transfer term(s) in eq 16 are evaluated easily, assuming that the mass flowing between neighboring zones has the thermodynamic properties (i.e., temperature, specific enthalpy, and chemical species composition) of the zone from which it originates.
φ ¼
mfuel =mair ðmfuel =mair Þst
ð17Þ
Sixteen (16) zones were used for the multizone model simulation, including the crevice zone (zone 1). The crevice volume is considered constant throughout the engine cycle and estimated to be about 1.5 cm3. The temperature and composition at IVC event are considered uniform throughout the cylinder. 5. Results and Discussion In the following subsections, a validation of the multizone model against experimental cases is first presented in terms of pressure traces, heat release rates, and specific HC and CO emissions. Subsequently, an analysis is presented regarding the formation of CO emissions as predicted by the multizone model, for some of the experimental cases. The same analysis is presented for the formation of HC emissions, in which a distinction is made between oxygenated and nonoxygenated species. The predicted composition of the unburned HC emissions;including oxygenated and nonoxygenated species;is also given for various load cases. Finally, these simulated HC data are used to estimate the relative error in the detection of predicted HC emissions, if they were actually introduced in a FID analyzer. 5.1. Pressure Traces and Heat Release Rates. In Figures 2 and 3 the experimental and simulated pressure traces and heat release rates are shown for isooctane and ethanol, respectively. For the ethanol simulation results to match the ignition timing, the temperature at inlet valve closing Tivc had to be increased somewhat. This could be attributed to fluctuations of Tivc relative to the inlet temperature Tin, which was measured prior to the inlet valve port. Heat transfer phenomena in the inlet duct could contribute to this
3. Engine Specifications - Operating and Modeling Conditions For the validation of the multizone model, experimental cases are compared to simulation predictions for ethanol and isooctane. The experimental data used for reference were provided by Lund Institute of Technology from a relevant experimental investigation. The details of this investigation have been published in ref 9, and involve the use of ethanol and isooctane as a HCCI fuel under various equivalence ratios. The engine considered was a Volvo TD100 diesel engine, modified to operate in HCCI mode. This was achieved by preheating the air stream and injecting the fuel into the air stream prior to its entrance to the engine cylinder. The main engine design data used in the experiments are given in Table 2, while the corresponding operating conditions are shown in Table 3. The term φ included in Table 4 is defined as the fuel-air equivalence ratio, that is, the actual fuel-air ratio divided by 1659
Energy Fuels 2010, 24, 1655–1667
: DOI:10.1021/ef9015153
Komninos and Rakopoulos
Figure 2. Experimental and simulated pressure traces and heat release rates for isooctane:(a) φ = 0.1131, (b) φ = 0.1504, (c) φ = 0.1866, (d) φ = 0.2075, (e) φ = 0.2770.
discrepancy. For the case of isooctane a lower temperature than Tin was used. Although this lower Tivc requirement for isooctane relative to ethanol, it is in accordance to the higher octane number of ethanol (ON 107) and it is generally met in studies9,10,24 that the inlet temperature requirement for the two fuels is roughly the same or slightly higher for isooctane. This difference between simulated and experimental initial temperature requirement can be attributed partly to the aforementioned uncertainty of the actual Tivc at experimental conditions and, more importantly, to the specific oxidation mechanisms used for the simulation of combustion. An extensive study on isooctane oxidation mechanisms and the relevant differences of ignition delay predictions can be found in ref 46. The simulation pressure traces show adequate agreement with the experimental ones. Heat release rates are also in good agreement, despite the overestimation of the peak heat release rates. This overestimation could be attributed partly to the unavoidable smoothing of the experimental heat release rate and also to the different models used for the estimation of heat losses during combustion. The first peak in the simulated HRR for isooctane is not observed in the experimental HRR. This could be attributed to the chemical mechanism used for the description of isooctane oxidation. Moreover, some heat may have actually been released in the
experiments prior to the main heat release without inducing a sharp peak at the experimental HRR, due to its lower rate and extended duration. The data shown cover both relatively high and low load cases, as it is evident by the peak heat release rates. 5.2. CO Emissions. In Figure 4 the experimental and simulated specific CO emissions are presented. The specific emissions refer to emissions expressed in g/kWh and are evaluated over the closed part of the engine cycle, that is, from IVC to EVO. Therefore, the experimental values presented in ref 9, which were net values, were converted to values corresponding to the closed cycle. Since the work produced during the closed part of the cycle is greater than the net work produced for the cases studied, the experimental values are lower than the net values presented in ref 9. This difference between specific net values and specific IVC to EVO values is small at high loads and more pronounced at low loads, in which the net work is quite lower than the gross work. The CO emissions for ethanol at the highest load were too low to be extracted with accuracy from the original diagrams as presented in ref 9; hence, they are omitted. From Figure 4 it can be seen that the trend of decreasing CO emissions with increasing load is captured by the multizone model. However, the absolute values show some discrepancy with an underestimation for the isooctane case. For the ethanol case, CO emissions are underestimated at low load and overestimated at medium load. The prediction of CO
(46) Jia, M.; Xie, M. Fuel 2006, 85, 2593–2604.
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In Figures 5 and 6 the in-cylinder CO and zone CO production rates are shown for isooctane and ethanol fuel, respectively, at three load conditions for each case. For the isooctane case (Figure 5) the low CO oxidation rate at the lowest load (Figure 5a) is obvious: after the CO peak during main combustion, CO is reduced slightly and stabilizes early in the expansion stroke. In this load case some CO is produced in the outer zones (zones 1-7) and concurrently consumed in the inner and hotter zones (zones 8-16), as it is shown in Figure 5b. Figures 5c-e show that as load increases the CO oxidation rate increases, since the negative in-cylinder CO slope becomes steeper at the end of main combustion. At the highest load a positive CO-CA slope is observed after the main combustion (∼10-20° aTDC). This is indicative of partial HC oxidation after main combustion, which leads to CO production. This assumption is verified in Figure 5f, where a net positive CO production rate is observed in the outer zones (zones 2-5). CO emissions formation from ethanol is presented in Figure 6. The main features of CO formation observed for isooctane are also present for ethanol fuel. The bulk gas quenching is evident in Figures 6a and b, although it is less severe than the isooctane case, probably due to the higher fuel-air equivalence ratio of ethanol. Figures 6c and e correspond to the higher load cases with a positive CO-CA slope after the main combustion event due to HC partial oxidation. This positive slope is not observed at the low load case (Figure 6a). The CO production is more pronounced at the higher load cases due to the higher gas temperature, which increases the CO production rate in the outer zones (Figures 6d and f). 5.3. HC Emissions. In Figure 7 the specific total (unburned) hydrocarbons (THC) emissions is shown. The agreement is less satisfactory relatively to the CO emissions. However, it must be noted that simulated HC emissions refer to in-cylinder emissions and include both oxygenated and nonoxygenated compounds. A portion of the HC located within the combustion chamber will not escape toward the exhaust valve, while the one that does escape from the combustion chamber may be partially oxidized in the exhaust manifold. Therefore, in-cylinder HC emissions at EVO are expected to be higher than the exhaust values as shown in the simulated results of Figure 7. The prediction of the work produced during the closed cycle could also contribute to any discrepancy present, since these are specific HC emissions. Another observation is that the overestimation of HC emissions is significantly higher for the ethanol case relative to isooctane case. This could be due to the fact that simulated THC emissions refer to both oxygenated and nonoxygenated HC. In the ethanol case oxygenated HC (OHC) emissions are expected to constitute a greater proportion of the total hydrocarbon emissions (THC). In order to investigate this assumption, the in-cylinder THC and OHC emissions are given in Figures 8 and 9 for isooctane and ethanol fuel respectively. The main species constituting the THC, that is, in quantities greater than 5 10-10 kmol, are also presented. It can be seen in Figures 8a, c, and e that for the case of isooctane the proportion of OHC into THC is higher at low load (about 51% by vol. or 33% by mass) and decreases at high load (about 36% by vol. or 19% by mass). Observing Figures 8b, d and f it is concluded that this trend is due to formaldehyde (CH2O), which is higher at low loads. The main species constituting the THC are
Figure 3. Experimental and simulated pressure traces and heat release rates for ethanol: (a) φ = 0.1544, (b) φ = 0.2083, (c) φ = 0.2641.
Figure 4. Experimental and simulated CO emissions results from isooctane-fueled (i-C8H18) and ethanol-fueled (C2H5OH) HCCI engine.
emissions is very sensitive to various in-cylinder conditions, such as the established temperature field, the initial temperature estimation, etc. All these factors could contribute to the observed discrepancies. 1661
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Figure 5. CO simulation results for isooctane; left: in-cylinder CO emissions, right: zone CO production rate.
formaldehyde, the fuel itself (isooctane), propene (C3H6), ethene (C2H4), and i-C4H8. The presence of formaldehyde in isooctane HCCI combustion has also been verified experimentally in ref 26. In Figure 9 the same analysis is applied to ethanol fuel for the load cases examined. Specifically, in Figures 9a, c, and d the model predictions show that most of the THC consist of oxygenated species at all loads. For the cases studied, OHC correspond to more than 90% by vol (95% by mass) of the THC. In Figures 9b, d, and f it is shown that the main oxygenated species are the fuel itself (ethanol) and, to a lesser extent, acetaldehyde (CH3CHO) and formaldehyde (CH2O). These oxygenated species were also found in the exhaust gases of conventional engines using ethanol-gasoline blends (up to 10% ethanol).47,48 Nonoxygenated species include methane (CH4) and ethene (C2H4). These results, that is, the proportion of OHC into THC and the relative quantities of the HC species, may vary, depending on the specific oxidation models used for the simulation of the combustion process. However, the main result drawn that ethanol, as an oxygenated fuel, produces
more oxygenated compounds than nonoxygenated fuels such as isooctane is well established.49 5.4. FID Error Estimation. The aforementioned findings have implications beyond the mere allocation of the THC in the results presented: if the simulated results shown are assumed to correspond to actual experimental cases, it is not certain how much of the oxygenated compounds will be detected by the usual HC measuring devices, such as the FID. The accuracy of the FID analyzer in detecting carbon atoms from organic compounds is not ensured in the case of oxygen-containing compounds. Although this might not lead to a substantial error in the isooctane cases examined, since less oxygenated compounds are produced, in the case of ethanol (or any other oxygenated fuel) the results may be misleading and alternative methods may be necessary to assess the amount of oxygenated compounds. Some of these oxygenated species are toxic substances or irritants and also contribute to the formation of the photochemical smog.47,48 In order to estimate the expected error in measuring the simulated HC emissions with a FID analyzer, the focus must shift to the detection procedure. Generally, the response of a compound introduced in the FID analyzer is determined
(47) Poulopoulos, S. G.; Samaras, D. P.; Philippopoulos, C. J. Atmos. Environ. 2001, 35, 4399–4406. (48) Pang, X.; Mu, Y.; Yuan, J.; He, H. Atmos. Environ. 2008, 42, 1349–1358.
(49) Heywood, J. B. Internal Combustion Engine Fundamentals; McGraw Hill: New York, 1988.
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Figure 6. CO simulation results for ethanol; left: in-cylinder CO emissions, right: zone CO production rate.
FID, [C] is the concentration of the compound (kmol/m3), and V_ is the sample volume flow rate through the detector (m3/s). Therefore, the response function of the compound is essentially the electric charge obtained from the oxidation of 1 kmol of the substance in the FID. For aliphatic hydrocarbons the response function is proportional to the carbon atoms in the molecule.29 The relative response (RR) of the organic compound is defined as the response function of the compound relative to the response of an alkane, for example, n-heptane, by defining a value for the latter equal to the number of carbon atoms in the molecule, that is, seven for n-heptane.29 A relative response of a compound less than the actual carbon atoms in the molecule of the compound implies that the FID response is lower than the one obtained by an alkane molecule of equal carbon atoms. This would lead to a FID measurement, indicating that the carbon atoms passed through the detector are less than the actual ones. The relative response of the FID analyzer for most nonoxygenated hydrocarbons (alkanes, alkenes, etc.) is usually equal to the carbon atoms in the molecule. However, for oxygenated hydrocarbons the relative response depends heavily on the chemical type of the OHC.28,29 Ketones, aldehydes, and ethers lower the apparent carbon atom count by 1. Therefore, it is highly unlikely that any formaldehyde
Figure 7. Experimental and simulated HC emissions results from isooctane-fueled (i-C8H18) and ethanol-fueled (C2H5OH) HCCI engine.
experimentally via eq 18:29 R ¼
i ½CV_
ð18Þ
Where R is the response function of the compound, i is the current obtained from the oxidation of the compound in the 1663
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Figure 8. HC simulation results for isooctane; left: in-cylinder total (THC) and oxygenated (OHC) hydrocarbon emissions, right: dominant HC species.
relative error in the kmol with respect to the FID measurement is given by eq 19: P P cm nm - RRm nm m P m RerrFID ¼ ð19Þ RRm nm
present in the exhaust will be detected by the FID analyzer. Moreover, any acetaldehyde present will be underestimated. For primary alcohols, such as methanol and ethanol, the noncontributing carbon is about 0.5. The relative responses of the compounds included in Figures 8 and 9 are given in Table 4. These relative responses were taken from refs 28 and 29, except for the ones of acetaldehyde and i-C4H8. For acetaldehyde the relative response is assumed to follow the trend of the other aldehydes,28 that is, to lower the apparent carbon count by 1. For i-C4H8, which is a nonoxygenated species, the relative response was set equal to the actual carbon content of the molecule as in the case of 1-butene28 and other alkenes. Using these relative responses, the deviation between the actual carbon atoms and the FID-measured carbon atoms can be assessed. This is achieved by multiplying the kmol of each compound by the number of the actual carbon atoms of its molecule, to obtain the total carbon contribution of the compound in kmol. Adding all the contributions from all HC compounds shown in Figure 6, the actual kmoles of carbon in the HC can be found. The corresponding measured kmol of carbon by the FID analyzer can be found by repeating the same procedure using the relative response of the FID instead of the actual carbon atoms in the molecule. The
m
Where RerrFID is the relative error with respect to the FID measurement, cm the carbon atoms per molecule of species m, nm the kmol of species m in the exhaust gases, and RRm is the relative FID response for species m. Index m refers to species which constitute the unburned HC in each load and fuel case, neglecting species with quantities lower than 5 10-10 kmol. Since the kmol of carbon are directly proportional to the carbon atoms, eq 19 also provides the relative error in carbon atoms detection. In Figure 10 the expected RerrFID versus load is shown for both fuels. The relative error is around 28% for the case of ethanol at all loads, owing to the great amount of ethanol in THC. This means that the measured FID response should be enhanced by 28% to provide the actual carbon in THC. For isooctane the relative error depends on load and it is governed primarily by the relative amount of formaldehyde in THC. As already mentioned, the relative error estimated herein depends on the relative proportion of oxygenates species, which is affected 1664
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Figure 9. HC simulation results for ethanol; left: in-cylinder total (THC) and oxygenated (OHC) hydrocarbon emissions, right: dominant HC species.
conducted; the appropriate molar mass assigned to each carbon atom is relatively low if this carbon atom was exclusively attached to hydrogen atoms but was relatively high if it was attached to an oxygen atom (as is the case with alcohols). Alternative devices for the measurement of oxygenated compounds and CFD or multizone modeling coupled with detailed oxidation mechanisms can help identify the main OHC compounds expected at the exhaust. The former can also lead to a quantification of the OHC from HCCI engines. Conclusions (1) The validation of the multizone model shows adequate agreement with the experimental results, provided that the initial temperature is adjusted for each fuel to reproduce the experimental ignition timing. This can be attributed to heat transfer phenomena in the inlet system. (2) Although the experimental inlet temperature requirement was the same for the two fuels, in the simulation a higher temperature was needed for ethanol relative to the corresponding initial temperature of isooctane. This can be attributed to the specific oxidation mechanisms used. (3) Both fuels showed a decrease in CO consumption rate at the low load included in this study, which was characteristic of
Figure 10. Estimated FID relative error in a hypothetical measurement of the simulated HC emissions.
by the specific chemical oxidation mechanism used for the combustion simulation. From the aforementioned, it can be concluded that the measured THC (on a ppm basis) will be generally underestimated, and that the accurate assessment of actual THC (and OHC) emissions requires the a priori knowledge of the composition of the unburned HC that are being measured, as was also pointed out for isooctane fuel in ref 27. This is especially true if an assessment of HC in mass terms is to be 1665
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Y = mass fraction (-) z = number of zones
bulk gas quenching. For the higher load cases, the fuels demonstrated postcombustion CO production during expansion due to partial HC oxidation resulting from mass transfer from the colder to the hotter zones. (4) The vast majority of total unburned hydrocarbons for the ethanol fuel were oxygenated compounds, primarily ethanol, acetaldehyde, and formaldehyde. For isooctane, the relative proportion of oxygenated HC to total HC was lower and increased at the low load case (bulk quenching) due to the presence of formaldehyde that was the main oxygenated compound. (5) The existence of oxygenated compounds at high levels in the exhaust gases for both fuels, and especially for ethanol at all loads examined, led to a theoretical assessment of the response of the FID analyzer for the simulated cases. The investigation showed that for the case of ethanol a significant underestimation of the total HC can result. The same holds true for isooctane at the limiting case of bulk quenching, due to the high formaldehyde concentration at the exhaust gases. At the higher load the relative proportion of formaldehyde was lower and the relative error in the determination of total HC decreased accordingly. (6) These results regarding the FID relative error in measuring HC emissions depend on the specific oxidation mechanism used, which determines the species existing in the exhaust gases. However, they are expected to be qualitatively correct, especially concerning the relative proportion of oxygenated HC emissions from oxygenated and nonoxygenated fuels. The results also point to the need for further investigation and for more accurate experimental assessment of oxygenated and nonoxygenated HC emissions using different devices.
Greek Symbols R = constant κ = von Karman constant φ = fuel-air equivalence ratio (-) μ = dynamic viscosity (kg/m s2) F = density (kg/m3) ω· = molar rate of production (kmol/m3 s) Subscripts cyl = cylinder in = inlet i = any zone j = any chemical species l = laminar m = any hydrocarbon species min = minimum n = normal net = net value (including all four strokes) oxid = oxidation pr = produced st = stoichiometric t = turbulent tot = total w = wall Superscripts * = characteristic value þ = dimensionless value Dimensionless numbers
Nomenclature
Pr = Prandtl number
B = cylinder bore (m) c = carbon atoms in a molecule h = molar specific enthalpy (J/kmol) h~ = mass specific enthalpy per kg of fuel (J/kg) i = electric current (A) J = number of chemical species k = thermal conductivity (W/m K) m = mass (kg) M = molar mass (kg/kmol) M = average zone molar mass (kg/kmol) nj = kmol of species j (kmol) P = pressure (N/m2) q_ = heat flux (W/m2) Qi = net heat gained by zone i (J) r = distance (m) R = FID response function of a compound (A/(kmol/s)) RR = relative FID response to a compound Ru = universal gas constant 8314 J/kmol K S = cylinder height (m) T = temperature (K) T = mean gas temperature (K) t = thickness (m), or time (s) u = velocity (m/s) U = internal energy (J) V = volume (m3) V_ = sample volume rate through the FID (m3/s)
Abbreviations a = after b = before BDC = bottom dead center CA = crank angle CO = carbon monoxide CFD = computational fluid dynamics CI = compression ignition CR = compression ratio EGR = exhaust gas recirculation EVC = exhaust valve closing EVO = exhaust valve opening FID = flame ionization detector HC = (unburned) hydrocarbons HCCI = homogeneous charge compression ignition HRR = heat release rate (J/deg) imep = indicated mean effective pressure (bar) IVC = inlet valve closing IVO = inlet valve opening NOx = nitrogen oxides OHC = oxygenated (unburned) hydrocarbons ON = octane number ppm = parts per million 1666
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PRF = primary reference fuel Rerr = relative error SI = spark ignition TDC = top dead center THC = total (unburned) hydrocarbons
Acknowledgment. We thank Professor Bengt Johansson and Ph.D. student Vittorio Manente from the Lund Institute of Technology, Sweden, for providing freely the experimental data used in this paper, for their willingness and for their useful comments.
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