- Email: [email protected]
Combustion characteristics of an SI engine fueled with H2–CO blended fuel and diluted by CO2
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 1 4 6 3 2 e1 4 6 3 9
Available online at www.sciencedirect.com
journal homepage: www.elsevier.com/locate/he
Combustion characteristics of an SI engine fueled with H2eCO blended fuel and diluted by CO2 Lei Chen a,*, Seiichi Shiga b, Mikiya Araki b a b
School of Aerospace Engineering, Shenyang Aerospace University, Liaoning, Shenyang 110136, China Department of Mechanical Engineering, Gunma University, 1-5-1 Tenjincho, Kiryu, Gunma 376-8515, Japan
article info
abstract
Article history:
This paper presents further results of the study on fundamental combustion characteris-
Received 23 April 2012
tics of gaseous fuels simulated for a biogas produced through a biomass gasification
Received in revised form
process with a catalyzer. The main work focuses on combustion characteristics of H2eCO
7 July 2012
blended fuel and the effect of CO2 dilution on it in a spark-ignition engine under the
Accepted 11 July 2012
condition of WOT, MBT and a constant speed of 1500 rpm. Equivalence ratio were limited to
Available online 13 August 2012
lower than 0.8 in order to avoid excessive high combustion temperature to damage the engine, and lean conditions were maintained during the experiment to get acceptable
Keywords:
economy and emissions. The results show that the BMEP decreases with an increase in
H2
dilution rate. The COV of IMEP is lower than 10% under most conditions, while H2 and CO2
CO
have the opposite influence on brake thermal efficiency. CO2 dilution combustion could
CO2 dilution
induce to remarkable decreasing in NOx emission with little decrease in brake thermal
Biomass fuel
efficiency, which benefits for biomass gaseous fuel application. If 500 ppm of NOx emission
Internal combustion engine
and 26% of brake thermal efficiency could be viewed as accepted level, the accepted operation range of H2eCO mixture have been obtained. Copyright ª 2012, Hydrogen Energy Publications, LLC. Published by Elsevier Ltd. All rights reserved.
1.
Introduction
The disposal of animal waste produced from livestock feedlots causes some environmental problems, such as surface or ground water contamination and air pollution with the release of flammable gas such as methane (CH4), and toxic gases such as ammonia (NH3) and hydrogen sulfide (H2S). Therefore, it is necessary to study environmentally proper method to process the animal waste. In the countryside of many countries, there are many farms of chicken, pigs and cows. The waste produced in these farms could be transformed into gaseous, liquid and solid impurity which includes combustible and in-combustible gases, while the traditional treatment of farm waste would
pollute the environment. As a result, it is necessary to explore new technique of waste after treatment. Some researchers carried out experimental investigation on low temperature gasification [1], which is an interesting alternative from the energy point of view and the temperature are usually around 700 C or even lower. The component of biogas varies with the reactants and the reaction conditions. However, it usually includes hydrogen (H2), carbon monoxide (CO), CH4, carbon dioxide (CO2), nitrogen (N2), and some other hydrocarbons. Fig. 1 shows the effect of experimental conditions on the produced gas components obtained in a study carried out by Xiao et al. [2]. The objective of their research is to introduce the design and construction of Internally Circulating Fluidized-bed Gasifier
* Corresponding author. E-mail address: [email protected] (L. Chen). 0360-3199/$ e see front matter Copyright ª 2012, Hydrogen Energy Publications, LLC. Published by Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.ijhydene.2012.07.048
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 1 4 6 3 2 e1 4 6 3 9
Fig. 1 e Effect of experiment conditions on product gas composition with temperature between 600 C and 700 C.
(ICFG), and to investigate the feasibility of gasifying manure compost using ICFG, and to evaluate the effects of pressure balance, reaction temperature and steam ratio on the performance of the gasifier. The temperature conditions of the experiments are between 600 C and 700 C. The results show that the concentration of the components varies much with the reaction conditions. The authors have conducted a research aimed at revealing the optimum gas components for fueling internal combustion engines as stationary use. In order to examine the fundamental combustion behavior, the study is divided into three stages, combustion characteristics of individual fuel component [3], effect of diluent component [4], and combustion of fuel mixtures. In the present study, combustion characteristics of mixed gas components are examined. In recent years, effects of H2 addition have been carried out by using both conventional liquid fuels (gasoline and diesel fuel) and a gaseous fuel (CNG) as the base fuel.
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Verhelst et al. [5] examined the combustion characteristics of a spark-ignition engine fueled with H2-gasoline blended fuel. They found an increase in thermal efficiency, especially at part load in a modified bi-fuel spark-ignition engine with compression ratio of 10.5. Ji et al. [6] applied H2 to a 4-cylinder spark-ignition gasoline engine. The H2 fraction in the total fuel (gasoline þ H2) was from 0 to 18% on mass basis, and the engine stability was improved by H2 addition. Porpatham et al. [7] conducted an experiment in a sparkignition engine fueled with simulated biogas with the main components of 2/3 CH4 and 1/3 CO2 on volume basis. H2 was admitted to the extent of 5, 10 and 15% by volume at whole open throttle (WOT). H2 addition extended lean burn limit, and improved thermal efficiency and power output. Akansu et al. [8] carried out an experimental research of the effect of H2 addition to CH4 in a spark-ignition engine with compression ratio of 10. CH4/H2 ratios of 100/0, 90/10, 80/20 and 70/30 on volume basis were tested. Equivalence ratio was varied from 0.6 to 1.2. Brake thermal efficiency and the exhaust emissions of hydrocarbon (HC), CO, CO2 were improved by H2 addition, but nitric oxides (NOx) emission was increased. Bauer and Forest [9] conducted a similar research in a single-cylinder CFR (coordinating fuel research) engine fueled with CH4eH2 blended fuel with displacement volume of 610 cc and compression ratio of 8.5. H2 volume percentage of 0, 20, 40 and 60% were employed. The partial burn limit was reduced by H2 addition. All CO2, CO and HC emissions were remarkably decreased. However, NOx was increased due to higher combustion temperature. Shudo et al. [10] examined the influence of H2 addition and CO2 dilution on CO combustion in a single-cylinder sparkignition engine with compression ratio of 13. The composition of gaseous mixture had a small influence on thermal efficiency, while NOx emission reduced with an increase in H2 fraction because the reduction of combustion temperature. Yamasaki et al. [11] investigated the combustion and engine performance of a spark ignition (SI) engine operated on simulated biogas. Two kinds of fuel conditions, constant air flow and constant fuel flow, were adopted. The simulated biogas gave 33% of indicated thermal efficiency and stable combustion over wide range equivalence ratio from 0.45 to 1.0. The combustion and exhaust emission characteristics of biogas fueled internal combustion engine with H2 addition were found in a few literatures. Shrestha and Narayanan [12] used landfill gas to fuel a CFR (cooperative fuel research) engine with variable compression ratio from 4 to 16. Effects of H2 addition up to 30% in the landfill gas were studied. Addition of even small quantities of H2 such as 3e5% induced to an increase in thermal efficiency from 33% to almost 38%. Cyclic variation also reduced due to H2 addition. Although it is shown that H2 addition improves thermal efficiency and engine stability, the problem of increase in NOx emission has not been satisfactorily resolved in these studies. Takarada, T. et al. [2] have been doing a study on so-called low-temperature gasification of biomass with a catalyst, which can give higher energy conversion efficiency from biomass to gaseous fuels. The main components of the produced gas are: H2 35e50%, CO 6e15%, CH4 2e8%, CO2 19e25%, N2 15e29% and some other impurity. Results of each
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i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 1 4 6 3 2 e1 4 6 3 9
component fuel combustion and dilution effect in a sparkignition engine were reported in previous papers [3,4]. This paper focuses on effects of H2 substitution and CO2 dilution on H2eCO blended fuel combustion. In the present study, wide range of H2 substitution ratio from 0 to 83% was covered to examine the combustion characteristics of blended fuel. The range of equivalence ratios were from the lean misfire limit of 0.3e0.8. The rich limit of 0.8 was determined to avoid excessively high combustion temperature which may cause a thermal problem in the engine.
2.
Experimental setup and procedure
A production spark-ignition engine with the main specifications given in Table 1 was utilized in the experiment without any major modifications. Fig. 2 shows the schematic diagram of the experimental setup. The engine was coupled to an eddy current type dynamometer to be operated for a performance test. A data acquisition system was used to store the cylinder pressure data from a KISTLER 6125A transducer. Four gas bottles were prepared to supply H2, CO, CH4 and CO2. These gases were introduced into the intake manifold through each pressure regulator and each gas injector for a production CNG engine (Keihin Co.). Thus, four gas injectors were installed at the intake passage approximately 400 mm upstream of the intake valve. A fine mesh made of thin stainless wires was installed at 300 mm downstream of the gas injectors and at 100 mm upstream of the intake valve as a flame arrester. Considering the length of the intake passage between the flame arrester and the gas injectors, the existence of the flame arrester and the mean Reynolds number of intake gaseous mixture, those supplied gases were pretty well-mixed prior to be sucked into the cylinder. The injection pressure for CO, CH4 and CO2 were set at the original value of 0.255 MPa, while for H2, the injection pressure was reduced to 0.105 MPa to extend the lower limit of the fuel flow rate. The exhaust gas was analyzed with a HORIBA MEXA4000 FT-IR analyzer. The engine was operated under WOT, maximum brake torque (MBT) and a constant speed of 1500 rpm.
3.
Fuel conditions
As mentioned before, the produced gas composition can vary much with the gasification conditions. The present study aims
Table 1 e Engine specifications. HONDA GX340, single-cylinder, air-cooled, 4-stroke Valve arrangement Combustion chamber geometry Stroke volume (cc) Bore Stroke (mm) Compression ratio Rated power (kW/rpm) for gasoline Throttle setting Engine speed
Single over-head valve Wedge 337 82 64 8 8.1/3600 WOT 1500 rpm
at giving useful information of the optimum gas composition in terms of engine combustion. Thus, the range of gas composition is not only determined on the basis of the previous studies of low-temperature gasification [2], but it also includes the extended range out of the previous possibility, since it must be worth to show the possible condition of the composition in the future. In the present study, CO and H2 are supplied as fuels, since these are usually major components of fuel, and CO2 is supplied as a diluent. Since WOT at 1500 rpm is maintained, the basic parameter of fueling is the equivalence ratio, f. Air and fuel flow rate are measured respectively to determine the equivalence ratio conditions. In order to express the variation of supplied gas composition, two parameters are defined: Hydrogen Substitution Ratio, H.S.R. and Fuel Dilution Ratio, F.D.R. Equations (1) and (2) show the definition of H.S.R. and F.D.R., respectively. H:S:R: ¼
fH2 fH2 þ fCO
(1)
F:D:R: ¼
CO2;mole Fuelmole þ CO2;mole
(2)
As shown in the previous study on the combustion characteristics of individual gas component [4], H2 has higher limit of f due to the occurrence of backfire. Thus, in the present study of supplying two fuel gases, CO is taken to be the base gas, and the effect of substituting the part of CO for H2 is examined at a constant f, which can be expressed by H.S.R. Regarding the effect of dilution with CO2, the volume fraction of CO2 in the fuel gas was taken to be the variable to express the extent of dilution, which can be expressed by F.D.R. The intake gas components are shown in Fig. 3.
4.
Results and discussions
4.1.
Effect of H2 substitution without dilution
Combustion experiments of pure H2eCO blended fuel are carried out to examine the effect of H2 substitution. Fig. 4 shows the effect of H2 substitution on BMEP, brake thermal efficiency and main combustion duration under different equivalence ratio conditions. In this paper, the initial combustion duration is defined as the crank angle from 0 to 10 percent mass fraction of the cylinder mass has burned, and the main combustion duration is defined as the crank angle from 10 to 90 percent mass fraction of the cylinder mass has burned [13]. Except for the highest f ¼ 0.8, the H.S.R. has almost no effect on BMEP. As shown in the previous study of individual fuel combustion, CO-air mixture has higher calorific value (LHVmixture base CO: 83.56 MJ/kmol (mixture f ¼ 1), H2: 71.40 MJ/ kmol (mixture f ¼ 1)), higher combustion temperature and then higher NOx comparing with the H2-air mixture. Since H2 combustion gives the same change as CO combustion in terms of molar numbers, an increase in H.S.R. should have induced to a decrease in the BMEP due to the reduction of input heat. However, in fact BMEP hardly changes, or even slightly increases according to Fig. 4(a). This indicates that H2 fuel has
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 1 4 6 3 2 e1 4 6 3 9
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Fig. 2 e Experimental setup of SI combustion.
better characteristics than CO. It must be ascribed for the faster burn of H2 as shown in the main combustion duration in Fig. 4(c). At f ¼ 0.8, the main combustion duration decreases remarkably. Thus it is shown that H2 substitution for CO is effective to improve the thermal efficiency, mainly because of the faster burn for H2 fuel. Moreover, H2 addition could improve combustion efficiency of H2eCO blended fuel, which benefits for brake thermal efficiency. As shown in Fig. 4(b), except for the leanest condition, the brake thermal efficiency can reach 28% over wide range of H.S.R. With increasing H.S.R., brake thermal efficiency increases at any f. That is mainly because the input heat decreases with substituting more H2 at a constant f and at almost constant or even more output of BMEP. Fig. 5 shows NOx and CO emissions. When f is lower than 0.4, NOx emission is almost 0. When f is 0.6, NOx emission decreases with the increase in H.S.R. This supports higher heating value and combustion temperature for CO than H2 discussed above, since more fraction of H2 decreases the combustion temperature due to a decrease in input heat. Fig. 6 shows the exhaust temperature, which could partly indicate combustion temperature. As shown in the figure, there is an apparent decrease in exhaust temperature with H.S.R.
Fig. 3 e Intake gas components without dilution.
increasing at each f, especially at f ¼ 0.6. It indicates that H2 addition leads to lower combustion temperature than pure CO at the same f due to less input heat, and could explain the variation of NOx emission when f ¼ 0.6 in Fig. 5. When f increases to 0.8, according to Fig. 6 it could affirm that the combustion temperature is much higher than other conditions, as a result the NOx emission observed in Fig. 5 is beyond the measurable range of the analyzer. This also verifies higher combustion temperature for CO combustion than H2. There is a clear decreasing tendency of CO emission against H.S.R. at each f. This must be reasonable in terms of lowering the amount of CO in the unburned fuel. However, it also shows that H.S.R. is not a unique factor determining the CO emission, since there is an increase in CO emission with decreasing f. It would also be reasonable in terms of the increase in the quenching distance with the decrease in f.
4.2. Effect of F.D.R. (CO2 dilution) on the combustion of H2eCO mixture When the experiments of H2eCO mixture were finished, CO2 is introduced into the mixture to examine the effect of dilution. Suitable F.D.R. conditions for each f are determined according to the results of H2eCO blending combustion. Fig. 7 shows the effect of F.D.R. on the engine performance parameters. With the increase in F.D.R. at a constant f, the amount of input heat decreases. It unavoidably reduces the BMEP, which is shown as dashed lines in the figure of BMEP. At higher f and relatively lower F.D.R., the whole parameters shown in Fig. 7 vary consistently with the variation of input heat. However, when f decreases and/or F.D.R. increases, the BMEP reduction becomes more than expected from the reduction of input heat, which could not be explained only by the input heat variation. This may not be related to the change of engine stability and the lengthening the combustion duration by dilution as shown in these figures. Fig. 8 shows the correlation between COVIMEP (coefficient of variation of IMEP) and BMEP. The correlation appears two
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i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 1 4 6 3 2 e1 4 6 3 9
Fig. 5 e Variation of exhaust emissions with H.S.R. (a) The correlation between NOx emission and H.S.R. (b) The correlation between CO emission and H.S.R.
Fig. 4 e Variation of basic performance parameters with H.S.R. (a) The correlation between BMEP and H.S.R. (b) The correlation between brake thermal efficiency and H.S.R. (c) The correlation between main combustion duration and H.S.R.
ranges. In range 1, linear correlation between BMEP and COVIMEP could be obtained, while in range 2 it could be seen that there is a remarkable increase in COVIMEP with little increasing of BMEP. The different correlation between BMEP and COVIMEP could not be explained only by the input heat variation. In reference to Fig. 8, the F.D.R. of data in range 1 are less than 0.6, while in range 2 F.D.R. equals 0.7. As discussed in former work, the reason could be viewed as cycle variation
Fig. 6 e The correlation between exhaust temperature and H.S.R.
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 1 4 6 3 2 e1 4 6 3 9
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Fig. 7 e Variation of basic performance parameters with F.D.R. (a) The correlation between BMEP and F.D.R. (b) The correlation between COVIMEP and F.D.R. (c) The correlation between ht and F.D.R. (d) The correlation between main combustion duration and F.D.R.
caused by partial burn. The correlation between COVIMEP and initial combustion duration is shown in Fig. 9. There are also two ranges: one shows the linear correlation while the other shows the remarkably increasing of COVIMEP with little initial combustion duration variation. In range 1, the initial combustion duration increases linearly with COVIMEP, which indicates that cycle variations linearly increase. While in
Fig. 8 e Correlation between BMEP and COVIMEP.
range 2, COVIMEP remarkably increases without greatly initial combustion duration change under F.D.R. limitation of each condition. As been discussed, excess F.D.R. induces to partial burn, which causes especial high cycle variation. This leads to remarkable reduction of BMEP and brake thermal efficiency, as well as greatly increased COVIMEP. And also, in a certain
Fig. 9 e Correlation between initial combustion duration and COVIMEP.
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In CO emission figure, it also shows that the increase in CO emission at low f such as f ¼ 0.3 and 0.4 is much faster than that of f ¼ 0.8. And also, CO emission at some conditions such as f ¼ 0.3, H.S.R. ¼ 67% and f ¼ 0.4, H.S.R. ¼ 25% is beyond 15,000 ppm. It indicates that high F.D.R. at low f or H.S.R. conditions will lead to poor combustion efficiency. As a result, although high F.D.R. could lead to desirable NOx emission, it should be avoided at low f or H.S.R. conditions. In terms of NOx emission, when f is less than 0.6 and F.D.R. beyond 0.2 almost no NOx emission can be observed. Combining the acceptable level of deterioration of thermal efficiency, if the thermal efficiency of 26% and NOx emission of 500 ppm would be set as the acceptable level, the realistic and favorable gas component could be obtained as shown in Table 2.
5.
Fig. 10 e Variation of exhaust emissions with F.D.R. (a) NOx emission. (b) CO emission. range with high f, high H.S.R. and low F.D.R., partial burn could not be observed, COVIMEP of all condition are lower than 5%. The result suggested that high H.S.R. and/or low F.D.R. conditions are necessary for engine stability. Fig. 10 shows the variation of NOx and CO emissions with F.D.R. As shown in Fig. 10, NOx decreases and CO increases with the decrease in f and the increase in F.D.R. The reduction of NOx is due to the reduction of combustion temperature, and the increase in CO is due to the increase in the quenching distance.
Table 2 e Favorable gas components at f [ 0.6
Conclusions
Although input heat decreases with increasing H.S.R, BMEP increases because of faster burn of H2 than CO. This also improves brake thermal efficiency over whole range of f. Higher H.S.R. induces to less NOx and CO emissions for less input heat and less input CO, respectively. Increasing F.D.R. could remarkably reduce NOx emission for the reduction of combustion temperature, while CO emissions gets worse for increased quenching distance. With increasing F.D.R., BMEP and brake thermal efficiency decrease for partial burn. However, in a certain range with high f, high H.S.R. and low F.D.R., engine stability was high enough to give any unfavorable effect on combustion under most operation fuel conditions. If the thermal efficiency of 26% would be set as the acceptable level, large operation range could be obtained by utilizing H2 substitution and CO2 dilution, which are benefit for improve combustion.
Acknowledgments The authors would like to express their gratitude for the financial support from Gunma Industry Support Organization, and acknowledge Prof. TAKARADA’s group for their help with the experiment.
references
f
H.S.R.
F.D.R.
he
NOx ppm
0.6
16.7% 16.7% 33.3% 33.3% 50.0% 50.0% 50.0% 66.7% 66.7% 83.3% 83.3% 83.3% 83.3%
0.1 0.2 0.1 0.2 0.1 0.2 0.3 0.1 0.2 0.1 0.2 0.3 0.4
26% 27% 27% 27% 27% 27% 26% 27% 26% 28% 27% 27% 27%
436 209 137 104 95 63 12 123 0 81 46 13 0
[1] Sweeten JM, Annamalai K, Thien B, McDonald LA. Co-firing of coal and cattle feedlot biomass (FB) fuels. Part I. Feedlot biomass (cattle manure) fuel quality and characteristics. International Journal of Fuel 2003;82:1167e82. [2] Xiao XB, Le DD, Morishita K, Zhang SY, Li LY, Takarada T. Multi-stage biomass gasification in Internally Circulating Fluidized-bed Gasifier (ICFG): test operation of animal-wastederived biomass and parametric investigation at low temperature. Fuel Processing Technology, in press. [3] Shiina R, Araki M, Chen L, Nakamura H, Shiga S, Obokata T. Performance and emissions of a spark-ignition engine fuelled with lower calorific gases (basic combustion characteristics of individual gas component). Transactions of
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 1 4 6 3 2 e1 4 6 3 9
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the Japan Society of Mechanical Engineers 2008;74(748): 2684e9. Chen L, Shiina R, Araki M, Nakamura H, Shiga S, Obokata T. Effect of CO2 dilution on the combustion characteristics of a spark-Ignition engine fuelled with individual component of bio-gas, Transactions of the Japan Society of Mechanical Engineers, in press. Verhelst S, Maesschalck P, Rombaut N, Sierens R. Efficiency comparison between hydrogen and gasoline, on a bi-fuel hydrogen/gasoline engine. International Journal of Hydrogen Energy 2009;34:2504e10. Ji CW, Wang SF. Effect of hydrogen addition on the idle performance of a spark ignited gasoline engine at stoichiometric condition. International Journal of Hydrogen Energy 2009;34:3546e56. Porpatham E, Ramesh A, Nagalingam B. Effect of hydrogen addition on the performance of a biogas fuelled spark ignition engine. International Journal of Hydrogen Energy 2007;32:2057e65. Akansu SO, Kahraman N, Ceper B. Experimental study on a spark ignition engine fuelled by methane-hydrogen mixtures. International Journal of Hydrogen Energy 2007;32: 4279e84. Bauer CG, Forest TW. Effect of hydrogen addition on the performance of methane-fueled vehicles. Part I: effect on S.I. engine performance. International Journal of Hydrogen Energy 2001;26:55e70.
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[10] Shudo T, Nakajima Y, Tsuga K. Combustion characteristics of H2-CO-CO2 mixture in an IC engine; 2001. SAE paper, 200101-0252. [11] Yamasaki Y, Tomatsu G, Nagata Y, Kaneko S. Development of high thermal efficiency and small size gas engine system using biomass gas fuel; 2007. SAE paper, 2007-01-2042. [12] Shrestha B, Narayanan G. Landfill gas with hydrogen addition e a fuel for SI engines. International Journal of Fuel 2008;87:3616e26. [13] Heywood JB. Internal combustion engine fundamentals. 1st ed.; 1988.
Nomenclatures BMEP: Brake mean effective pressure, MPa CO2, molar: Molar number of CO2 COV of IMEP: Coefficient of variation of IMEP ECU: Electronic control unit F.D.R.: Fuel dilution ratio Fuel molar: Molar number of fuels H.S.R.: Hydrogen substitution ratio LHV: Lower heating value, kJ/kg NOx: Nitric oxides concentration, ppm f: Equivalence ratio fCO: Equivalence ratio of CO fH2: Equivalence ratio of H2
Available online at www.sciencedirect.com
journal homepage: www.elsevier.com/locate/he
Combustion characteristics of an SI engine fueled with H2eCO blended fuel and diluted by CO2 Lei Chen a,*, Seiichi Shiga b, Mikiya Araki b a b
School of Aerospace Engineering, Shenyang Aerospace University, Liaoning, Shenyang 110136, China Department of Mechanical Engineering, Gunma University, 1-5-1 Tenjincho, Kiryu, Gunma 376-8515, Japan
article info
abstract
Article history:
This paper presents further results of the study on fundamental combustion characteris-
Received 23 April 2012
tics of gaseous fuels simulated for a biogas produced through a biomass gasification
Received in revised form
process with a catalyzer. The main work focuses on combustion characteristics of H2eCO
7 July 2012
blended fuel and the effect of CO2 dilution on it in a spark-ignition engine under the
Accepted 11 July 2012
condition of WOT, MBT and a constant speed of 1500 rpm. Equivalence ratio were limited to
Available online 13 August 2012
lower than 0.8 in order to avoid excessive high combustion temperature to damage the engine, and lean conditions were maintained during the experiment to get acceptable
Keywords:
economy and emissions. The results show that the BMEP decreases with an increase in
H2
dilution rate. The COV of IMEP is lower than 10% under most conditions, while H2 and CO2
CO
have the opposite influence on brake thermal efficiency. CO2 dilution combustion could
CO2 dilution
induce to remarkable decreasing in NOx emission with little decrease in brake thermal
Biomass fuel
efficiency, which benefits for biomass gaseous fuel application. If 500 ppm of NOx emission
Internal combustion engine
and 26% of brake thermal efficiency could be viewed as accepted level, the accepted operation range of H2eCO mixture have been obtained. Copyright ª 2012, Hydrogen Energy Publications, LLC. Published by Elsevier Ltd. All rights reserved.
1.
Introduction
The disposal of animal waste produced from livestock feedlots causes some environmental problems, such as surface or ground water contamination and air pollution with the release of flammable gas such as methane (CH4), and toxic gases such as ammonia (NH3) and hydrogen sulfide (H2S). Therefore, it is necessary to study environmentally proper method to process the animal waste. In the countryside of many countries, there are many farms of chicken, pigs and cows. The waste produced in these farms could be transformed into gaseous, liquid and solid impurity which includes combustible and in-combustible gases, while the traditional treatment of farm waste would
pollute the environment. As a result, it is necessary to explore new technique of waste after treatment. Some researchers carried out experimental investigation on low temperature gasification [1], which is an interesting alternative from the energy point of view and the temperature are usually around 700 C or even lower. The component of biogas varies with the reactants and the reaction conditions. However, it usually includes hydrogen (H2), carbon monoxide (CO), CH4, carbon dioxide (CO2), nitrogen (N2), and some other hydrocarbons. Fig. 1 shows the effect of experimental conditions on the produced gas components obtained in a study carried out by Xiao et al. [2]. The objective of their research is to introduce the design and construction of Internally Circulating Fluidized-bed Gasifier
* Corresponding author. E-mail address: [email protected] (L. Chen). 0360-3199/$ e see front matter Copyright ª 2012, Hydrogen Energy Publications, LLC. Published by Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.ijhydene.2012.07.048
i n t e r n a t i o n a l j o u r n a l o f h y d r o g e n e n e r g y 3 7 ( 2 0 1 2 ) 1 4 6 3 2 e1 4 6 3 9
Fig. 1 e Effect of experiment conditions on product gas composition with temperature between 600 C and 700 C.
(ICFG), and to investigate the feasibility of gasifying manure compost using ICFG, and to evaluate the effects of pressure balance, reaction temperature and steam ratio on the performance of the gasifier. The temperature conditions of the experiments are between 600 C and 700 C. The results show that the concentration of the components varies much with the reaction conditions. The authors have conducted a research aimed at revealing the optimum gas components for fueling internal combustion engines as stationary use. In order to examine the fundamental combustion behavior, the study is divided into three stages, combustion characteristics of individual fuel component [3], effect of diluent component [4], and combustion of fuel mixtures. In the present study, combustion characteristics of mixed gas components are examined. In recent years, effects of H2 addition have been carried out by using both conventional liquid fuels (gasoline and diesel fuel) and a gaseous fuel (CNG) as the base fuel.
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Verhelst et al. [5] examined the combustion characteristics of a spark-ignition engine fueled with H2-gasoline blended fuel. They found an increase in thermal efficiency, especially at part load in a modified bi-fuel spark-ignition engine with compression ratio of 10.5. Ji et al. [6] applied H2 to a 4-cylinder spark-ignition gasoline engine. The H2 fraction in the total fuel (gasoline þ H2) was from 0 to 18% on mass basis, and the engine stability was improved by H2 addition. Porpatham et al. [7] conducted an experiment in a sparkignition engine fueled with simulated biogas with the main components of 2/3 CH4 and 1/3 CO2 on volume basis. H2 was admitted to the extent of 5, 10 and 15% by volume at whole open throttle (WOT). H2 addition extended lean burn limit, and improved thermal efficiency and power output. Akansu et al. [8] carried out an experimental research of the effect of H2 addition to CH4 in a spark-ignition engine with compression ratio of 10. CH4/H2 ratios of 100/0, 90/10, 80/20 and 70/30 on volume basis were tested. Equivalence ratio was varied from 0.6 to 1.2. Brake thermal efficiency and the exhaust emissions of hydrocarbon (HC), CO, CO2 were improved by H2 addition, but nitric oxides (NOx) emission was increased. Bauer and Forest [9] conducted a similar research in a single-cylinder CFR (coordinating fuel research) engine fueled with CH4eH2 blended fuel with displacement volume of 610 cc and compression ratio of 8.5. H2 volume percentage of 0, 20, 40 and 60% were employed. The partial burn limit was reduced by H2 addition. All CO2, CO and HC emissions were remarkably decreased. However, NOx was increased due to higher combustion temperature. Shudo et al. [10] examined the influence of H2 addition and CO2 dilution on CO combustion in a single-cylinder sparkignition engine with compression ratio of 13. The composition of gaseous mixture had a small influence on thermal efficiency, while NOx emission reduced with an increase in H2 fraction because the reduction of combustion temperature. Yamasaki et al. [11] investigated the combustion and engine performance of a spark ignition (SI) engine operated on simulated biogas. Two kinds of fuel conditions, constant air flow and constant fuel flow, were adopted. The simulated biogas gave 33% of indicated thermal efficiency and stable combustion over wide range equivalence ratio from 0.45 to 1.0. The combustion and exhaust emission characteristics of biogas fueled internal combustion engine with H2 addition were found in a few literatures. Shrestha and Narayanan [12] used landfill gas to fuel a CFR (cooperative fuel research) engine with variable compression ratio from 4 to 16. Effects of H2 addition up to 30% in the landfill gas were studied. Addition of even small quantities of H2 such as 3e5% induced to an increase in thermal efficiency from 33% to almost 38%. Cyclic variation also reduced due to H2 addition. Although it is shown that H2 addition improves thermal efficiency and engine stability, the problem of increase in NOx emission has not been satisfactorily resolved in these studies. Takarada, T. et al. [2] have been doing a study on so-called low-temperature gasification of biomass with a catalyst, which can give higher energy conversion efficiency from biomass to gaseous fuels. The main components of the produced gas are: H2 35e50%, CO 6e15%, CH4 2e8%, CO2 19e25%, N2 15e29% and some other impurity. Results of each
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component fuel combustion and dilution effect in a sparkignition engine were reported in previous papers [3,4]. This paper focuses on effects of H2 substitution and CO2 dilution on H2eCO blended fuel combustion. In the present study, wide range of H2 substitution ratio from 0 to 83% was covered to examine the combustion characteristics of blended fuel. The range of equivalence ratios were from the lean misfire limit of 0.3e0.8. The rich limit of 0.8 was determined to avoid excessively high combustion temperature which may cause a thermal problem in the engine.
2.
Experimental setup and procedure
A production spark-ignition engine with the main specifications given in Table 1 was utilized in the experiment without any major modifications. Fig. 2 shows the schematic diagram of the experimental setup. The engine was coupled to an eddy current type dynamometer to be operated for a performance test. A data acquisition system was used to store the cylinder pressure data from a KISTLER 6125A transducer. Four gas bottles were prepared to supply H2, CO, CH4 and CO2. These gases were introduced into the intake manifold through each pressure regulator and each gas injector for a production CNG engine (Keihin Co.). Thus, four gas injectors were installed at the intake passage approximately 400 mm upstream of the intake valve. A fine mesh made of thin stainless wires was installed at 300 mm downstream of the gas injectors and at 100 mm upstream of the intake valve as a flame arrester. Considering the length of the intake passage between the flame arrester and the gas injectors, the existence of the flame arrester and the mean Reynolds number of intake gaseous mixture, those supplied gases were pretty well-mixed prior to be sucked into the cylinder. The injection pressure for CO, CH4 and CO2 were set at the original value of 0.255 MPa, while for H2, the injection pressure was reduced to 0.105 MPa to extend the lower limit of the fuel flow rate. The exhaust gas was analyzed with a HORIBA MEXA4000 FT-IR analyzer. The engine was operated under WOT, maximum brake torque (MBT) and a constant speed of 1500 rpm.
3.
Fuel conditions
As mentioned before, the produced gas composition can vary much with the gasification conditions. The present study aims
Table 1 e Engine specifications. HONDA GX340, single-cylinder, air-cooled, 4-stroke Valve arrangement Combustion chamber geometry Stroke volume (cc) Bore Stroke (mm) Compression ratio Rated power (kW/rpm) for gasoline Throttle setting Engine speed
Single over-head valve Wedge 337 82 64 8 8.1/3600 WOT 1500 rpm
at giving useful information of the optimum gas composition in terms of engine combustion. Thus, the range of gas composition is not only determined on the basis of the previous studies of low-temperature gasification [2], but it also includes the extended range out of the previous possibility, since it must be worth to show the possible condition of the composition in the future. In the present study, CO and H2 are supplied as fuels, since these are usually major components of fuel, and CO2 is supplied as a diluent. Since WOT at 1500 rpm is maintained, the basic parameter of fueling is the equivalence ratio, f. Air and fuel flow rate are measured respectively to determine the equivalence ratio conditions. In order to express the variation of supplied gas composition, two parameters are defined: Hydrogen Substitution Ratio, H.S.R. and Fuel Dilution Ratio, F.D.R. Equations (1) and (2) show the definition of H.S.R. and F.D.R., respectively. H:S:R: ¼
fH2 fH2 þ fCO
(1)
F:D:R: ¼
CO2;mole Fuelmole þ CO2;mole
(2)
As shown in the previous study on the combustion characteristics of individual gas component [4], H2 has higher limit of f due to the occurrence of backfire. Thus, in the present study of supplying two fuel gases, CO is taken to be the base gas, and the effect of substituting the part of CO for H2 is examined at a constant f, which can be expressed by H.S.R. Regarding the effect of dilution with CO2, the volume fraction of CO2 in the fuel gas was taken to be the variable to express the extent of dilution, which can be expressed by F.D.R. The intake gas components are shown in Fig. 3.
4.
Results and discussions
4.1.
Effect of H2 substitution without dilution
Combustion experiments of pure H2eCO blended fuel are carried out to examine the effect of H2 substitution. Fig. 4 shows the effect of H2 substitution on BMEP, brake thermal efficiency and main combustion duration under different equivalence ratio conditions. In this paper, the initial combustion duration is defined as the crank angle from 0 to 10 percent mass fraction of the cylinder mass has burned, and the main combustion duration is defined as the crank angle from 10 to 90 percent mass fraction of the cylinder mass has burned [13]. Except for the highest f ¼ 0.8, the H.S.R. has almost no effect on BMEP. As shown in the previous study of individual fuel combustion, CO-air mixture has higher calorific value (LHVmixture base CO: 83.56 MJ/kmol (mixture f ¼ 1), H2: 71.40 MJ/ kmol (mixture f ¼ 1)), higher combustion temperature and then higher NOx comparing with the H2-air mixture. Since H2 combustion gives the same change as CO combustion in terms of molar numbers, an increase in H.S.R. should have induced to a decrease in the BMEP due to the reduction of input heat. However, in fact BMEP hardly changes, or even slightly increases according to Fig. 4(a). This indicates that H2 fuel has
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Fig. 2 e Experimental setup of SI combustion.
better characteristics than CO. It must be ascribed for the faster burn of H2 as shown in the main combustion duration in Fig. 4(c). At f ¼ 0.8, the main combustion duration decreases remarkably. Thus it is shown that H2 substitution for CO is effective to improve the thermal efficiency, mainly because of the faster burn for H2 fuel. Moreover, H2 addition could improve combustion efficiency of H2eCO blended fuel, which benefits for brake thermal efficiency. As shown in Fig. 4(b), except for the leanest condition, the brake thermal efficiency can reach 28% over wide range of H.S.R. With increasing H.S.R., brake thermal efficiency increases at any f. That is mainly because the input heat decreases with substituting more H2 at a constant f and at almost constant or even more output of BMEP. Fig. 5 shows NOx and CO emissions. When f is lower than 0.4, NOx emission is almost 0. When f is 0.6, NOx emission decreases with the increase in H.S.R. This supports higher heating value and combustion temperature for CO than H2 discussed above, since more fraction of H2 decreases the combustion temperature due to a decrease in input heat. Fig. 6 shows the exhaust temperature, which could partly indicate combustion temperature. As shown in the figure, there is an apparent decrease in exhaust temperature with H.S.R.
Fig. 3 e Intake gas components without dilution.
increasing at each f, especially at f ¼ 0.6. It indicates that H2 addition leads to lower combustion temperature than pure CO at the same f due to less input heat, and could explain the variation of NOx emission when f ¼ 0.6 in Fig. 5. When f increases to 0.8, according to Fig. 6 it could affirm that the combustion temperature is much higher than other conditions, as a result the NOx emission observed in Fig. 5 is beyond the measurable range of the analyzer. This also verifies higher combustion temperature for CO combustion than H2. There is a clear decreasing tendency of CO emission against H.S.R. at each f. This must be reasonable in terms of lowering the amount of CO in the unburned fuel. However, it also shows that H.S.R. is not a unique factor determining the CO emission, since there is an increase in CO emission with decreasing f. It would also be reasonable in terms of the increase in the quenching distance with the decrease in f.
4.2. Effect of F.D.R. (CO2 dilution) on the combustion of H2eCO mixture When the experiments of H2eCO mixture were finished, CO2 is introduced into the mixture to examine the effect of dilution. Suitable F.D.R. conditions for each f are determined according to the results of H2eCO blending combustion. Fig. 7 shows the effect of F.D.R. on the engine performance parameters. With the increase in F.D.R. at a constant f, the amount of input heat decreases. It unavoidably reduces the BMEP, which is shown as dashed lines in the figure of BMEP. At higher f and relatively lower F.D.R., the whole parameters shown in Fig. 7 vary consistently with the variation of input heat. However, when f decreases and/or F.D.R. increases, the BMEP reduction becomes more than expected from the reduction of input heat, which could not be explained only by the input heat variation. This may not be related to the change of engine stability and the lengthening the combustion duration by dilution as shown in these figures. Fig. 8 shows the correlation between COVIMEP (coefficient of variation of IMEP) and BMEP. The correlation appears two
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Fig. 5 e Variation of exhaust emissions with H.S.R. (a) The correlation between NOx emission and H.S.R. (b) The correlation between CO emission and H.S.R.
Fig. 4 e Variation of basic performance parameters with H.S.R. (a) The correlation between BMEP and H.S.R. (b) The correlation between brake thermal efficiency and H.S.R. (c) The correlation between main combustion duration and H.S.R.
ranges. In range 1, linear correlation between BMEP and COVIMEP could be obtained, while in range 2 it could be seen that there is a remarkable increase in COVIMEP with little increasing of BMEP. The different correlation between BMEP and COVIMEP could not be explained only by the input heat variation. In reference to Fig. 8, the F.D.R. of data in range 1 are less than 0.6, while in range 2 F.D.R. equals 0.7. As discussed in former work, the reason could be viewed as cycle variation
Fig. 6 e The correlation between exhaust temperature and H.S.R.
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Fig. 7 e Variation of basic performance parameters with F.D.R. (a) The correlation between BMEP and F.D.R. (b) The correlation between COVIMEP and F.D.R. (c) The correlation between ht and F.D.R. (d) The correlation between main combustion duration and F.D.R.
caused by partial burn. The correlation between COVIMEP and initial combustion duration is shown in Fig. 9. There are also two ranges: one shows the linear correlation while the other shows the remarkably increasing of COVIMEP with little initial combustion duration variation. In range 1, the initial combustion duration increases linearly with COVIMEP, which indicates that cycle variations linearly increase. While in
Fig. 8 e Correlation between BMEP and COVIMEP.
range 2, COVIMEP remarkably increases without greatly initial combustion duration change under F.D.R. limitation of each condition. As been discussed, excess F.D.R. induces to partial burn, which causes especial high cycle variation. This leads to remarkable reduction of BMEP and brake thermal efficiency, as well as greatly increased COVIMEP. And also, in a certain
Fig. 9 e Correlation between initial combustion duration and COVIMEP.
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In CO emission figure, it also shows that the increase in CO emission at low f such as f ¼ 0.3 and 0.4 is much faster than that of f ¼ 0.8. And also, CO emission at some conditions such as f ¼ 0.3, H.S.R. ¼ 67% and f ¼ 0.4, H.S.R. ¼ 25% is beyond 15,000 ppm. It indicates that high F.D.R. at low f or H.S.R. conditions will lead to poor combustion efficiency. As a result, although high F.D.R. could lead to desirable NOx emission, it should be avoided at low f or H.S.R. conditions. In terms of NOx emission, when f is less than 0.6 and F.D.R. beyond 0.2 almost no NOx emission can be observed. Combining the acceptable level of deterioration of thermal efficiency, if the thermal efficiency of 26% and NOx emission of 500 ppm would be set as the acceptable level, the realistic and favorable gas component could be obtained as shown in Table 2.
5.
Fig. 10 e Variation of exhaust emissions with F.D.R. (a) NOx emission. (b) CO emission. range with high f, high H.S.R. and low F.D.R., partial burn could not be observed, COVIMEP of all condition are lower than 5%. The result suggested that high H.S.R. and/or low F.D.R. conditions are necessary for engine stability. Fig. 10 shows the variation of NOx and CO emissions with F.D.R. As shown in Fig. 10, NOx decreases and CO increases with the decrease in f and the increase in F.D.R. The reduction of NOx is due to the reduction of combustion temperature, and the increase in CO is due to the increase in the quenching distance.
Table 2 e Favorable gas components at f [ 0.6
Conclusions
Although input heat decreases with increasing H.S.R, BMEP increases because of faster burn of H2 than CO. This also improves brake thermal efficiency over whole range of f. Higher H.S.R. induces to less NOx and CO emissions for less input heat and less input CO, respectively. Increasing F.D.R. could remarkably reduce NOx emission for the reduction of combustion temperature, while CO emissions gets worse for increased quenching distance. With increasing F.D.R., BMEP and brake thermal efficiency decrease for partial burn. However, in a certain range with high f, high H.S.R. and low F.D.R., engine stability was high enough to give any unfavorable effect on combustion under most operation fuel conditions. If the thermal efficiency of 26% would be set as the acceptable level, large operation range could be obtained by utilizing H2 substitution and CO2 dilution, which are benefit for improve combustion.
Acknowledgments The authors would like to express their gratitude for the financial support from Gunma Industry Support Organization, and acknowledge Prof. TAKARADA’s group for their help with the experiment.
references
f
H.S.R.
F.D.R.
he
NOx ppm
0.6
16.7% 16.7% 33.3% 33.3% 50.0% 50.0% 50.0% 66.7% 66.7% 83.3% 83.3% 83.3% 83.3%
0.1 0.2 0.1 0.2 0.1 0.2 0.3 0.1 0.2 0.1 0.2 0.3 0.4
26% 27% 27% 27% 27% 27% 26% 27% 26% 28% 27% 27% 27%
436 209 137 104 95 63 12 123 0 81 46 13 0
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Nomenclatures BMEP: Brake mean effective pressure, MPa CO2, molar: Molar number of CO2 COV of IMEP: Coefficient of variation of IMEP ECU: Electronic control unit F.D.R.: Fuel dilution ratio Fuel molar: Molar number of fuels H.S.R.: Hydrogen substitution ratio LHV: Lower heating value, kJ/kg NOx: Nitric oxides concentration, ppm f: Equivalence ratio fCO: Equivalence ratio of CO fH2: Equivalence ratio of H2