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Computer simulation of hydrogen–diesel dual fuel exhaust gas emissions with experimental verification
Available online at www.sciencedirect.com
Fuel 87 (2008) 1372–1378 www.fuelfirst.com
Computer simulation of hydrogen–diesel dual fuel exhaust gas emissions with experimental verification M. Masood *, M.M. Ishrat Mechanical Engineering Department, M.J. College of Engineering and Technology, Hyderabad 34, India Received 13 July 2006; received in revised form 26 June 2007; accepted 4 July 2007 Available online 30 July 2007
Abstract The drawback of lean operation with hydrocarbon fuels is a reduced power output. Lean operation of hydrocarbon engines has additional drawbacks. Lean mixtures are hard to ignite, despite the mixture being above the low fire (point) limit of the fuel. This results in misfire, which increases un-burned hydrocarbon emissions, reduces performance and wastes fuel. Hydrogen can be used in conjunction with compact liquid fuels such as gasoline; alcohol or diesel provided each is stored separately. Mixing hydrogen with other hydrocarbon fuels reduces all of these drawbacks. Hydrogen’s low ignition energy limit and high burning speed makes the hydrogen/hydrocarbon mixture easier to ignite, reducing misfire and thereby improving emissions, performance and fuel economy. Regarding power output, hydrogen augments the mixture’s energy density at lean mixtures by increasing the hydrogen-to-carbon ratio, and thereby improves torque at wide-open throttle conditions. This paper involves the simulation program for determining the mole fraction of each of the exhaust species when the hydrogen is burnt along with diesel and the results are presented. The proportion of hydrogen in the hydrogen–diesel blend affecting the mole fraction of the exhaust species is also simulated. Experimental investigations were carried out, in hydrogen–diesel dual fuel mode, which showed a good agreement between the predicted and experimental results. The program code developed is valid for any combination of dual fuels. 2007 Elsevier Ltd. All rights reserved. Keywords: Mole fraction; Equivalence ratio; Dissociation reaction; Combustion
1. Introduction Hydrogen, with its remarkable combustion properties in the conventional internal combustion engine, appears to be proving itself as the best transportation fuel of the future if it can be produced economically. Importantly, it can be used in existing internal combustion engines, yielding unprecedented efficiencies and extremely low levels of exhaust pollution [2]. Hydrogen has, for years, been recognized for its extremely high energy potential. But because of inherent difficulties in handling hydrogen in its gaseous form, technology has, over the past two decades, emphasized the utilization of hydrogen in its liquid form. *
Corresponding author. E-mail addresses: [email protected] (M. Masood), [email protected] (M.M. Ishrat). 0016-2361/$ - see front matter 2007 Elsevier Ltd. All rights reserved. doi:10.1016/j.fuel.2007.07.001
Most research in dual fuel engine has concentrated on defining the extent of dual fueling and its effect on emissions and performance [1,2]. Natural gas in combination with diesel was tried and found to be very effective in NOx reduction but engine operation can suffer from high hydrocarbons (HC) emissions and poor performance, especially at high loads [3,4]. The auto-ignition of methane was studied experimentally to obtain ignition delay data as a function of engine cylinder pressure and temperature by Sandia National Laboratory [5,6]. Karim et al. [9,10] and Gunea et al. [11] concluded that at low outputs, much of the primary gaseous fuel remain un-burned leading to high hydrocarbon (HC) and CO emissions which is mainly due to very lean operation and a weak ignition source. At high loads a large amount of gaseous fuel admission results in uncontrolled reaction rates near the pilot spray causing rough engine operation.
M. Masood, M.M. Ishrat / Fuel 87 (2008) 1372–1378
1373
Notation p q r s x y /
number of C atoms number of H atoms number of O atoms number of N atoms percentage of diesel in diesel–hydrogen fuel blend percentage of hydrogen in diesel–hydrogen fuel blend fuel–air equivalence ratio
Experimental investigation on a LPG–diesel dual fuel engine by Poonia et al. [7,8] revealed that at low loads, the brake thermal efficiency is always lower than diesel values but is better at high loads. Also, at low outputs increasing the pilot quantity and intake temperature improves the thermal efficiency. The HC and CO emissions were found to increase in the dual fuel mode. Daisho et al. [12] showed that emission characteristics and brake thermal efficiency of a dual engine using natural gas can be improved by reciprocating exhaust gas, increased brake thermal efficiency with increase intake temperature is also reported. In the present study experiments were conducted on hydrogen–diesel dual fuel engine under constant speed, variable compression ratios and variable load conditions. The amount of primary fuel, i.e. diesel admitted was varied and hydrogen was substituted at each load. The objective was to determine in detail the performance, emissions and combustion characteristics of the engine. 1.1. Problem formation For any of the fuel combinations, it becomes imperative to optimize the combination depending not only upon the performance but also upon exhaust emissions. For this if the optimal performance and reduced emissions if simulated, can practically be achieved effectively. Here in this paper an effort is made to simulate the exhaust emissions in dual fuel mode. The optimal combination of fuels depending upon the exhaust can be found effectively and applied practically. This program code is even valid with slight variation to the other combination of fuels. Thermodynamics is able to predict the equilibrium state that results from burning a fuel–air mixture given only the initial conditions. Combustion is a chemical reaction between a fuel and oxygen, which is accompanied by the production of a considerable amount of heat. The composition of the exhaust gas produced is a function of temperature as well as equivalence ratio (ratio of actual fuel air ratio to theoretical fuel air ratio). A lean mixture has / < 1. A rich mixture has / > 1. The mixture is said to be stoichiometric if / = 1. Many components are present
e K T P mi yi
stoichiometric molar fuel–air ratio equilibrium constant temperature in K pressure in bar coefficient describing product composition of ith species mole fraction of i-th species
in the exhaust gas because of dissociation of some species. The heat of combustion of a fuel is defined as the heat transferred out of a system per unit mass or mole of fuel when the initial and final states are at the same temperature and pressure [13]. Based on the combustion stoichiometric theory, a computer program had been developed for blended fuels to calculate the mole fractions of the exhaust gases [14]. Thermodynamic data for elements, combustion products and many pollutants are available in a compilation published by the National Bureau of Standards, called the JANAF (Joint Army–Navy–Air Force) tables (1971). For single component fuels the data presented by Stull, Westrum and Sinke (1969) is in the same format as that of JANAF tables. A compilation by Rossini (1953) is useful for hydrocarbon fuels at temperatures as high as 1500 K. 2. Inputs to the program The fuel is to be specified in terms of the C, H, O and N atoms in the fuel. For the blend of two fuels considered i.e., diesel and hydrogen, the percentage with which they blend in the mixture also has to be specified. The other parameters that need to be specified are equivalence ratio, pressure and temperature. For the calculation of equilibrium constant, the data for constants is considered from JANAF tables. The molar-air fuel ratio is calculated from the number of carbon, hydrogen, nitrogen and oxygen atoms present in the fuel. 3. Logic of the computer program 3.1. Formation of equations The mixture is a blend of fuel of composition Cp HqOrNs and hydrogen. Considering that there are 10 constituents the combustion reaction is written as e/½x ðCp Hq Or Ns ÞÞ þ ðy H2 Þ þ ð0:21 O2 Þ þ ð0:79 N2 Þ ! t1 CO2 þ t2 H2 O þ t3 N2 þ t4 O2 þ t5 CO þ t6 H2 þ t7 H8 O þ t9 OH þ t10 NO
ð1Þ
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M. Masood, M.M. Ishrat / Fuel 87 (2008) 1372–1378
where ti are the coefficients that describe the product composition [14]. The molar fuel–air ratio is given by e ¼ 0:21 ½x=ðp þ 0:25 q 0:5 rÞ þ 2 y
ð2Þ
Convenient approximations for lean and rich combustion are / < 1 t5 ¼ 0
ð3Þ
/ > 1 t4 ¼ 0
ð4Þ
Olikara and Borman (1975) have curve fitted the equilibrium constants to JANAF table data. Their expressions are of the form log K p ¼ AlnðT =1000Þ þ ðB=T Þ þ C þ ðD T Þ þ ðE T 2 Þ ð5Þ where T is in K. The value of A, B, C, D and E are obtained from the JANAF tables [13]. The mole fractions are obtained for the products are obtained by y i ¼ mi =Rmi i ¼ 1 to 6
ð6Þ
For a lean mixture the coefficients of combustion products are obtained as 9 t1 ¼ x ðp / eÞ; > > > > t2 ¼ ðx ðq / ðe=2ÞÞÞ þ ðy ðq / ðe=2ÞÞÞ; > > > > = t3 ¼ ðx ð0:79 þ ðs / ðe=2ÞÞÞÞ þ ðy 0:79Þ; ð7Þ t4 ¼ ðx ð0:21 ð1 /ÞÞÞ þ ðy ð0:21 ð1 /ÞÞÞ; > > > > > > t5 ¼ 0; > > ; t6 ¼ y 0:42; For a rich mixture the coefficients of combustion products are obtained as t5 ¼ ðb þ ðsqrtððb bÞ ð4 a cÞÞÞÞ=ð2 aÞ
ð8aÞ
where a ¼ ðx ð1 kÞÞ þ ðy ð1 kÞÞ; b ¼ ðx ð0:42 ð/ e ð2 rÞÞ þ ðk ðð0:42 ð/ 1ÞÞþ
9 > > > =
ðp / eÞÞÞÞÞ þ ðy ð0:42 ð2 / eÞ þ ðk ð0:42 ð/ 1ÞÞÞÞÞ; > > > ; c ¼ ðx ð0:42 p / e ð/ 1Þ kÞÞ;
ð8bÞ
and k ¼ expð0:273 ð1:761=tÞ ð1:611=t2Þ þ ð0:283=t3ÞÞ; 9 t1 ¼ ðx ððp / eÞ v5ÞÞ þ ðy v5Þ; > > > > t2 ¼ ðx ð0:42 þ ð/ e ðð2 pÞ rÞÞ þ v5ÞÞ ðy ð0:42 þ v5ÞÞ; > > = t3 ¼ ðx ð0:79 þ ðs / ðe=2ÞÞÞÞ þ ðy 0:79Þ; t4 ¼ 0; t6 ¼ ðx ðð0:42 ð/ 1ÞÞ v5ÞÞ þ ðy 0:42Þ;
> > > > > > ;
ð8cÞ
The mole fractions for all the remaining species is obtained in terms of y3, y4 and y6 i.e., the mole fractions of N2, O2 and H2, respectively as 9 0:5 > y 7 ¼ C 1 ðy 6 Þ ; > > > 0:5 = y 8 ¼ C 2 ðy 4 Þ ; ð9aÞ y 9 ¼ C 3 ðy 4 Þ0:5 ðy 6 Þ0:5 ; > > > > 0:5 0:5 ; y 10 ¼ C 4 ðy 4 Þ ðy 3 Þ ; 9 C 1 ¼ K 1 =P 1=2 ; > > > C 2 ¼ K 2 =P 1=2 ; = ð9bÞ > C3 ¼ K 3; > > ; C4 ¼ K 4; where KP value is obtained from Eq. (5). With the variation in the input parameters various results and plots can be obtained. 3.2. Experimental setup The essential components for the experimental setup were chosen properly and assembled keeping in mind the objectives of the research work. The instrumentation systems were selected judiciously with a clear understanding of their working, range and limitations. The engine used in the present study was a Kirloskar AV-1, single cylinder direct injection diesel engine with the specifications given in Table 1. The engine was coupled to a DC dynamometer and all the experiments were carried out a constant speed of 1400 rpm. The engine was modified and provision was provided to vary the compression ratio from 14.5 to 24.5. Crank-angle resolved in-cylinder pressure and the diesel injection pressure were measured. A computer interfaced piezoelectric sensor, of range 145 bar was used to sense the in-cylinder pressure. Pressure signals were obtained at 1 crank-angle intervals using a digital data acquisition system. The average pressure data from 100 consecutive cycles were used for calculating combustion parameters. Special software was used to obtain combustion parameters. The transducer is connected to the computer through a logic card, which converts the pressure signals to digital signals. The Hydrogen gas was injected into the cylinder and a thermal mass flow controller controlled its flow rate. Diesel was injected through a nozzle of orifice diameter 0.15 mm, having only one orifice and its flow was controlled by governor mechanism. The maximum amount of hydrogen supplied was limited by unstable operation at low outputs and by rough engine running due to knock at high outputs. The special software stores the data of pressures and volumes corresponding to a particular crank-angle location for plotting the P–V and P–h curves. Engine exhausts emissions were measured using an advanced AVL five-gas analyzer, which is a non dispersive infrared gas analyzer. The sample to be evaluated is passed through a cold trap to condense the water vapors, which influences the functioning of the infrared analyzer. The
M. Masood, M.M. Ishrat / Fuel 87 (2008) 1372–1378 Table 1 Specification
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4-Stroke, single cylinder, compression ignition engine with variable compression ratio
Make Rated power Bore and stroke Compression ratio Cylinder capacity Dynamometer Cylinder pressure Orifice dia. Fuel Hydrogen injection
Kirloskar AV-1 3.7 KW, 1500 RPM 85 mm · 110 mm 16.5: 1, variable from 14.3 to 24.5 624. cc Electrical-AC alternator Piezo sensor, range: 145 bar 0.15 mm Diesel, hydrogen By hydrogen injector
exhaust gas analyzer is calibrated periodically using standard calibration gas. Standard Bosch smoke measuring instrument is used to measure the exhaust smoke emission from the engine. In this paper, only the gaseous exhaust emission results are presented to validate the prediction. 4. Results and discussion Figs. 1 and 2 show the change in mole fraction of CO2, forvarious percentage hydrogen substitutions at different constant equivalence ratios. It can be noted that for all equivalence ratios except for 1.2 and 1.4, the mole fraction of CO2 increases between 20% and 40% hydrogen substitution and then decreases with further increase in hydrogen percentage, however for equivalence ratios of 1.2 and 1.4, it increases continuously. As shown in Fig. 3, with higher temperature of 3000 K, the CO2 value is lower than that of 2000 K. CO2 is one of the major products of lean as well as rich combustion. Under the stoichiometric conditions the mole fractions of CO2 is at its peak and decreases when the mixture becomes either richer or leaner due to simultaneous presence of other products. As the temperature increases, the mole fraction of CO2 decreases since the dissociation increases with temperature. Figs. 4 and 5, shows the change in mole fraction of NO for various percentages of hydrogen substitutions for dif-
0.3
Phi = 0.6
0.2
phi=0.9 0.1
phi=1.2
0 0
20
40
60
80
100
% Hydrogen
Fig. 2. Effect of hydrogen substitution on CO2 at T = 2000 K.
ferent constant equivalence ratios for temperatures of 1500 K and 2000 K. With the increase in percentage of hydrogen, the NO decreases slightly. As the value of equivalence ratio is increasing, the mole fraction of NO is decreasing for both temperatures of 1500 K and 2000 K. Similar trend was also noted for 2500 K and 3000 K. As shown in Fig. 6, for 80% Hydrogen substitution, and with 3000 K, the mole fraction of NO is higher than that of other lower temperatures. NO present in the exhaust is very much less as the concentration of N atoms is very much less. The lack of dissociation of N2 molecules is the result of the strong triple covalent bond. Equilibrium concentrations of NO are rather flat and peak in the lean region falling rapidly in the rich region. In most combustion systems, NO levels are well below the equilibrium concentrations
Mole Fraction Vs Equivalence Ratio, 80% Hydrogen Mole Fraction Of CO2
Type
Mole Fraction of CO2
Mole Fraction of CO2 Vs Equivalence Ratio,T=2000k
0.4 T=2000 0.3 T=2500 0.2 T=3000 0.1 0 0
0.5
1
1.5
Equivalence Ratio
Fig. 3. Mole fraction of CO2 against the equivalence ratio at 80% hydrogen and at T = 2000 K, 2500 K, and 3000 K.
Mole Fraction Of NO Vs% Hydrogen , T = 1500
Mole Fraction of CO2 Vs % Hydrogen
0.03
0.3
Phi =0.3
0.25
Phi =0.5
0.2
Phi = 0.8
0.15
Phi = 1.0
0.1
Phi = 1.2
0.05
Phi = 1.4
0
Mole Fraction NO
Mole fraction of CO2
0.35
0.025 0.02
Phi =0.3
0.015
Phi =0.5
0.01
Phi = 0.8
0.005 0
0
20
40
60
80
100
% Hydrogen
Fig. 1. Effect of hydrogen substitution on CO2 at T = 1500 K.
0
50
100
% Hydrogen
Fig. 4. Effect of hydrogen substitution on NO at T = 1500 K.
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M. Masood, M.M. Ishrat / Fuel 87 (2008) 1372–1378 Mole Fraction of N2 Vs% H2, T = 1500 K
Mole Fraction of NO Vs Equivalence Ratio
0.03 20% H2
0.02 40% H2
0.01 60% H2
Mole fraction of N2
Mole Fraction Of NO
0% H2
0 0
0.5
1
1.5
0.9 0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1 0
Fig. 5. Effect of equivalence ratio on NO T = 1500 K.
Phi =0.5 Phi = 0.8 Phi = 1.0 Phi = 1.2 Phi = 1.4
0
80% H2
Equivalence Ratio
Phi =0.3
T=2000
0.02
T=2500
0.015 0.01
T=3000
0.005 0 0
0.5
1
100
Mole Fraction of H2o Vs % Hydrogen, T=1500 k Mole fraction of H2o
Mole Fraction Of NO
0.03
40 60 80 % Hydrogen
Fig. 8. Effect of hydrogen substitution on N2 at T = 1500 K.
Mole Fraction of NO Vs Equivalence Ratio , 80% Hydrogen
0.025
20
1.5
0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1 0
Phi =0.3 Phi =0.5 Phi = 0.8 Phi = 1.0 Phi = 1.2 Phi = 1.4 0
Equivalence Ratio
20
40
60
80
100
% Hydrogen
Fig. 6. Effect of temperature on NOx. Fig. 9. Effect of hydrogen substitution on H2O at T = 1500 K.
because of relatively slow formation reaction. As the temperature increases, the mole fraction of NO increases due to increase in dissociation of N2 [13]. Fig. 7 shows the results for CO against the percentage hydrogen substitution. It can be noted that as the hydrogen percentage is increasing the mole fraction of CO increases, and for more than 1 equivalence ratio, the increase is very sharp. However its value is lower with higher temperatures. CO is the minor species in the lean combustion and as the mixture gets richer the mole fraction of CO increases due to incomplete combustion of carbon. Further, as the temperature increases CO2 dissociates to form CO and hence mole fraction of CO increases as temperature increases (see Fig. 7). The decreasing behavior of N2 was observed in Fig. 8, and the trend was the same as that was observed in case
Mole Fraction Vs Equivalence Ratio , 80% Hydrogen
Mole Fraction for CO
T=2000 0.08
of NO. As shown in Fig. 9, as the percentage hydrogen increases, the mole fraction of H2O also increases; it has highest value at equivalence ratio of 1. If the equivalence ratio is further increased, mole fraction of H2O falls down considerably. As the mole fraction of H2O increases with hydrogen substitution, it brings down the peak combustion temperature, and hence resultantly the values of NO at higher percentages of hydrogen substitution decrease considerably. Figs. 10–15 show the comparison of predicted and experimental values of CO2, NOx, and CO, respectively with variation of hydrogen substitution and also against the equivalence ratios. The predicted values are in good agreement with the experimental results. This validates the code and the other predictions emerging from the same code.
CO2 Vs % H2 0.3 0.25 0.2
T=2500
0.06 0.04
T=3000
0.02
Experiment Simulated
0.15 0.1 0.05
0 -0.02 0
0.5
1
1.5
Equivalence Ratio
Fig. 7. Effect of temperature on CO.
0 0
20
40 60 80 % Hydrogen
100
Fig. 10. Comparison of simulated and experimental values of CO2.
M. Masood, M.M. Ishrat / Fuel 87 (2008) 1372–1378 CO2 Vs Phi (Equivalence ratio)
0.25 Simulated experimental
0.1
Mole fraction of CO
Mole fraction of CO2
0.3
0.15
CO vs Phi
0.35
0.35
0.2
1377
0.3 0.25 0.2
Experiment Simulated
0.15 0.1 0.05 0
0.05
0
0 0
0.5
1
1.5
Equivalence Vs Phi
0.2 0.4 0.6 0.8 1 1.2 1.4 Phi (Equivalence ratio)
Fig. 15. Comparison of simulated and experimental values of CO.
Fig. 11. Comparison of simulated and experimental results at different equivalence ratios.
5. Conclusions NOx Vs % H2 0.016 0.014 0.012 0.01 Simulated Experiment
0.008 0.006 0.004
The code developed gave reasonably good results which are in good agreement with the experimental values. However, there exist many areas which are unaddressed by the code. At low and high percentages of hydrogen and during transition between diesel and hydrogen the model predictions are not very clear, this eventually shows the limitation of the model and opens the doors for further investigation.
0.002 0 0
20 40 60 80 100 % Hydrogen substitution
Fig. 12. Comparison of simulated and experimental values of NOx.
NOx vs Phi Mole fraction of CO2
0.016 0.014 0.012 0.01 Simulated Experiment
0.008 0.006 0.004 0.002
• For all equivalence ratios except for 1.2 and 1.4, the mole fraction of CO2 increases between 20% and 40% hydrogen substitution and then decreases with further increase in hydrogen percentage. However for equivalence ratios of 1.2 and 1.4, it increases continuously. • As the value of equivalence ratio increases, the mole fraction of NOx decreases for both temperatures of 1500 K and 2000 K. • As the hydrogen percentage increases the mole fraction of CO increases, and for more than 1 equivalence ratio, the increase is sharp. • The mole fraction of H2O increases with hydrogen substitution; this decreased the peak combustion temperatures.
0 0
0.5
1
1.5
References
Equivalence ratio (phi)
Fig. 13. Comparison of simulated and experimental values of NOx With the variations of equivalence ratios.
CO Vs % Hydrogen 0.35 0.3 0.25 0.2
Experiment Simulated
0.15 0.1 0.05 0 0
20
40
60
80
100
% Hydrogen substitution
Fig. 14. Comparison of simulated and experimental values of CO.
[1] Karim GA, Liu T, Jones W. Exhaust emissions from dual fuel engines at light loads. Presented at SAE Paper. SAE 932822; 1993. [2] Yashuhiro Kou T, Yuki Shigeki N, Ryoji K, Takeshi S. Controlling combustion and exhaust emissions in a direct-injection diesel engine dual fueled with natural gas. Presented at SAE Paper. SAE 952436; 1995. [3] Abd-Alla GH, Soliman HA, Badr OA, Abd-Rabbo MF. Energ Convers Manage 2000;41:559–72. [4] Singh S. The effect of fuel injection timing and pilot quantity on the pollutant emissions from a pilot ignited natural gas engine. Masters thesis, The University Of Alabama; 2002. [5] Fraser RA, Siebers DL, Edwards CF. Auto-ignition of methane and natural gas in a simulated diesel environment. Presented at SAE Paper. SAE 2003-01-0755. [6] Naber JD, Siebers DL, Caton JA, Westbrook CK, Di jjulio SS. Natural gas auto-ignition under diesel condition: experiments and chemical kinetics modeling. Presented at SAE Paper. SAE 942034; 1994. [7] Poonia PM, Ramesh A, Gaur RR. Experimental investigation of the factors effecting the performance of a LPG–diesel dual fuel engine.
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SAE Paper No. 99-01-1123, SAE transactions, Journal of Fuels and Lubricants; 1999. [8] Poonia PM, Ramesh A, Gaur RR. The effect of air temperature and pilot fuel quantity on the combustion characteristics of a LPG–diesel dual fuel engine. SAE Paper No. 982455; 1998. [9] Karim GA. The dual fuel engine of the compression ignition type – prospects, problems and solutions – a review. SAE Paper No. 831073; 1983. [10] Karim GA, Jones W, Rain RR. An experimentation of ignition delay period in dual fuel engines.
[11] Gunea C, Razavi MRM, Karim GA. The effect of pilot fuel quantity on dual fuel engine ignition delay. SAE Paper No. 982453; 1998. [12] Daisho Y, et.al. Controlling combustion and exhaust emissions in a direct injection diesel engine dual-fueled with natural gas. SAE Paper No. 952436; 1995. [13] Deshpande NV et.al, Exhaust gas simulation. IE (I) Journal–MC; 2004. [14] Ferguson CR. Internal combustion engines – applied thermo sciences. John Wiley and Sons; 1986.
Fuel 87 (2008) 1372–1378 www.fuelfirst.com
Computer simulation of hydrogen–diesel dual fuel exhaust gas emissions with experimental verification M. Masood *, M.M. Ishrat Mechanical Engineering Department, M.J. College of Engineering and Technology, Hyderabad 34, India Received 13 July 2006; received in revised form 26 June 2007; accepted 4 July 2007 Available online 30 July 2007
Abstract The drawback of lean operation with hydrocarbon fuels is a reduced power output. Lean operation of hydrocarbon engines has additional drawbacks. Lean mixtures are hard to ignite, despite the mixture being above the low fire (point) limit of the fuel. This results in misfire, which increases un-burned hydrocarbon emissions, reduces performance and wastes fuel. Hydrogen can be used in conjunction with compact liquid fuels such as gasoline; alcohol or diesel provided each is stored separately. Mixing hydrogen with other hydrocarbon fuels reduces all of these drawbacks. Hydrogen’s low ignition energy limit and high burning speed makes the hydrogen/hydrocarbon mixture easier to ignite, reducing misfire and thereby improving emissions, performance and fuel economy. Regarding power output, hydrogen augments the mixture’s energy density at lean mixtures by increasing the hydrogen-to-carbon ratio, and thereby improves torque at wide-open throttle conditions. This paper involves the simulation program for determining the mole fraction of each of the exhaust species when the hydrogen is burnt along with diesel and the results are presented. The proportion of hydrogen in the hydrogen–diesel blend affecting the mole fraction of the exhaust species is also simulated. Experimental investigations were carried out, in hydrogen–diesel dual fuel mode, which showed a good agreement between the predicted and experimental results. The program code developed is valid for any combination of dual fuels. 2007 Elsevier Ltd. All rights reserved. Keywords: Mole fraction; Equivalence ratio; Dissociation reaction; Combustion
1. Introduction Hydrogen, with its remarkable combustion properties in the conventional internal combustion engine, appears to be proving itself as the best transportation fuel of the future if it can be produced economically. Importantly, it can be used in existing internal combustion engines, yielding unprecedented efficiencies and extremely low levels of exhaust pollution [2]. Hydrogen has, for years, been recognized for its extremely high energy potential. But because of inherent difficulties in handling hydrogen in its gaseous form, technology has, over the past two decades, emphasized the utilization of hydrogen in its liquid form. *
Corresponding author. E-mail addresses: [email protected] (M. Masood), [email protected] (M.M. Ishrat). 0016-2361/$ - see front matter 2007 Elsevier Ltd. All rights reserved. doi:10.1016/j.fuel.2007.07.001
Most research in dual fuel engine has concentrated on defining the extent of dual fueling and its effect on emissions and performance [1,2]. Natural gas in combination with diesel was tried and found to be very effective in NOx reduction but engine operation can suffer from high hydrocarbons (HC) emissions and poor performance, especially at high loads [3,4]. The auto-ignition of methane was studied experimentally to obtain ignition delay data as a function of engine cylinder pressure and temperature by Sandia National Laboratory [5,6]. Karim et al. [9,10] and Gunea et al. [11] concluded that at low outputs, much of the primary gaseous fuel remain un-burned leading to high hydrocarbon (HC) and CO emissions which is mainly due to very lean operation and a weak ignition source. At high loads a large amount of gaseous fuel admission results in uncontrolled reaction rates near the pilot spray causing rough engine operation.
M. Masood, M.M. Ishrat / Fuel 87 (2008) 1372–1378
1373
Notation p q r s x y /
number of C atoms number of H atoms number of O atoms number of N atoms percentage of diesel in diesel–hydrogen fuel blend percentage of hydrogen in diesel–hydrogen fuel blend fuel–air equivalence ratio
Experimental investigation on a LPG–diesel dual fuel engine by Poonia et al. [7,8] revealed that at low loads, the brake thermal efficiency is always lower than diesel values but is better at high loads. Also, at low outputs increasing the pilot quantity and intake temperature improves the thermal efficiency. The HC and CO emissions were found to increase in the dual fuel mode. Daisho et al. [12] showed that emission characteristics and brake thermal efficiency of a dual engine using natural gas can be improved by reciprocating exhaust gas, increased brake thermal efficiency with increase intake temperature is also reported. In the present study experiments were conducted on hydrogen–diesel dual fuel engine under constant speed, variable compression ratios and variable load conditions. The amount of primary fuel, i.e. diesel admitted was varied and hydrogen was substituted at each load. The objective was to determine in detail the performance, emissions and combustion characteristics of the engine. 1.1. Problem formation For any of the fuel combinations, it becomes imperative to optimize the combination depending not only upon the performance but also upon exhaust emissions. For this if the optimal performance and reduced emissions if simulated, can practically be achieved effectively. Here in this paper an effort is made to simulate the exhaust emissions in dual fuel mode. The optimal combination of fuels depending upon the exhaust can be found effectively and applied practically. This program code is even valid with slight variation to the other combination of fuels. Thermodynamics is able to predict the equilibrium state that results from burning a fuel–air mixture given only the initial conditions. Combustion is a chemical reaction between a fuel and oxygen, which is accompanied by the production of a considerable amount of heat. The composition of the exhaust gas produced is a function of temperature as well as equivalence ratio (ratio of actual fuel air ratio to theoretical fuel air ratio). A lean mixture has / < 1. A rich mixture has / > 1. The mixture is said to be stoichiometric if / = 1. Many components are present
e K T P mi yi
stoichiometric molar fuel–air ratio equilibrium constant temperature in K pressure in bar coefficient describing product composition of ith species mole fraction of i-th species
in the exhaust gas because of dissociation of some species. The heat of combustion of a fuel is defined as the heat transferred out of a system per unit mass or mole of fuel when the initial and final states are at the same temperature and pressure [13]. Based on the combustion stoichiometric theory, a computer program had been developed for blended fuels to calculate the mole fractions of the exhaust gases [14]. Thermodynamic data for elements, combustion products and many pollutants are available in a compilation published by the National Bureau of Standards, called the JANAF (Joint Army–Navy–Air Force) tables (1971). For single component fuels the data presented by Stull, Westrum and Sinke (1969) is in the same format as that of JANAF tables. A compilation by Rossini (1953) is useful for hydrocarbon fuels at temperatures as high as 1500 K. 2. Inputs to the program The fuel is to be specified in terms of the C, H, O and N atoms in the fuel. For the blend of two fuels considered i.e., diesel and hydrogen, the percentage with which they blend in the mixture also has to be specified. The other parameters that need to be specified are equivalence ratio, pressure and temperature. For the calculation of equilibrium constant, the data for constants is considered from JANAF tables. The molar-air fuel ratio is calculated from the number of carbon, hydrogen, nitrogen and oxygen atoms present in the fuel. 3. Logic of the computer program 3.1. Formation of equations The mixture is a blend of fuel of composition Cp HqOrNs and hydrogen. Considering that there are 10 constituents the combustion reaction is written as e/½x ðCp Hq Or Ns ÞÞ þ ðy H2 Þ þ ð0:21 O2 Þ þ ð0:79 N2 Þ ! t1 CO2 þ t2 H2 O þ t3 N2 þ t4 O2 þ t5 CO þ t6 H2 þ t7 H8 O þ t9 OH þ t10 NO
ð1Þ
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M. Masood, M.M. Ishrat / Fuel 87 (2008) 1372–1378
where ti are the coefficients that describe the product composition [14]. The molar fuel–air ratio is given by e ¼ 0:21 ½x=ðp þ 0:25 q 0:5 rÞ þ 2 y
ð2Þ
Convenient approximations for lean and rich combustion are / < 1 t5 ¼ 0
ð3Þ
/ > 1 t4 ¼ 0
ð4Þ
Olikara and Borman (1975) have curve fitted the equilibrium constants to JANAF table data. Their expressions are of the form log K p ¼ AlnðT =1000Þ þ ðB=T Þ þ C þ ðD T Þ þ ðE T 2 Þ ð5Þ where T is in K. The value of A, B, C, D and E are obtained from the JANAF tables [13]. The mole fractions are obtained for the products are obtained by y i ¼ mi =Rmi i ¼ 1 to 6
ð6Þ
For a lean mixture the coefficients of combustion products are obtained as 9 t1 ¼ x ðp / eÞ; > > > > t2 ¼ ðx ðq / ðe=2ÞÞÞ þ ðy ðq / ðe=2ÞÞÞ; > > > > = t3 ¼ ðx ð0:79 þ ðs / ðe=2ÞÞÞÞ þ ðy 0:79Þ; ð7Þ t4 ¼ ðx ð0:21 ð1 /ÞÞÞ þ ðy ð0:21 ð1 /ÞÞÞ; > > > > > > t5 ¼ 0; > > ; t6 ¼ y 0:42; For a rich mixture the coefficients of combustion products are obtained as t5 ¼ ðb þ ðsqrtððb bÞ ð4 a cÞÞÞÞ=ð2 aÞ
ð8aÞ
where a ¼ ðx ð1 kÞÞ þ ðy ð1 kÞÞ; b ¼ ðx ð0:42 ð/ e ð2 rÞÞ þ ðk ðð0:42 ð/ 1ÞÞþ
9 > > > =
ðp / eÞÞÞÞÞ þ ðy ð0:42 ð2 / eÞ þ ðk ð0:42 ð/ 1ÞÞÞÞÞ; > > > ; c ¼ ðx ð0:42 p / e ð/ 1Þ kÞÞ;
ð8bÞ
and k ¼ expð0:273 ð1:761=tÞ ð1:611=t2Þ þ ð0:283=t3ÞÞ; 9 t1 ¼ ðx ððp / eÞ v5ÞÞ þ ðy v5Þ; > > > > t2 ¼ ðx ð0:42 þ ð/ e ðð2 pÞ rÞÞ þ v5ÞÞ ðy ð0:42 þ v5ÞÞ; > > = t3 ¼ ðx ð0:79 þ ðs / ðe=2ÞÞÞÞ þ ðy 0:79Þ; t4 ¼ 0; t6 ¼ ðx ðð0:42 ð/ 1ÞÞ v5ÞÞ þ ðy 0:42Þ;
> > > > > > ;
ð8cÞ
The mole fractions for all the remaining species is obtained in terms of y3, y4 and y6 i.e., the mole fractions of N2, O2 and H2, respectively as 9 0:5 > y 7 ¼ C 1 ðy 6 Þ ; > > > 0:5 = y 8 ¼ C 2 ðy 4 Þ ; ð9aÞ y 9 ¼ C 3 ðy 4 Þ0:5 ðy 6 Þ0:5 ; > > > > 0:5 0:5 ; y 10 ¼ C 4 ðy 4 Þ ðy 3 Þ ; 9 C 1 ¼ K 1 =P 1=2 ; > > > C 2 ¼ K 2 =P 1=2 ; = ð9bÞ > C3 ¼ K 3; > > ; C4 ¼ K 4; where KP value is obtained from Eq. (5). With the variation in the input parameters various results and plots can be obtained. 3.2. Experimental setup The essential components for the experimental setup were chosen properly and assembled keeping in mind the objectives of the research work. The instrumentation systems were selected judiciously with a clear understanding of their working, range and limitations. The engine used in the present study was a Kirloskar AV-1, single cylinder direct injection diesel engine with the specifications given in Table 1. The engine was coupled to a DC dynamometer and all the experiments were carried out a constant speed of 1400 rpm. The engine was modified and provision was provided to vary the compression ratio from 14.5 to 24.5. Crank-angle resolved in-cylinder pressure and the diesel injection pressure were measured. A computer interfaced piezoelectric sensor, of range 145 bar was used to sense the in-cylinder pressure. Pressure signals were obtained at 1 crank-angle intervals using a digital data acquisition system. The average pressure data from 100 consecutive cycles were used for calculating combustion parameters. Special software was used to obtain combustion parameters. The transducer is connected to the computer through a logic card, which converts the pressure signals to digital signals. The Hydrogen gas was injected into the cylinder and a thermal mass flow controller controlled its flow rate. Diesel was injected through a nozzle of orifice diameter 0.15 mm, having only one orifice and its flow was controlled by governor mechanism. The maximum amount of hydrogen supplied was limited by unstable operation at low outputs and by rough engine running due to knock at high outputs. The special software stores the data of pressures and volumes corresponding to a particular crank-angle location for plotting the P–V and P–h curves. Engine exhausts emissions were measured using an advanced AVL five-gas analyzer, which is a non dispersive infrared gas analyzer. The sample to be evaluated is passed through a cold trap to condense the water vapors, which influences the functioning of the infrared analyzer. The
M. Masood, M.M. Ishrat / Fuel 87 (2008) 1372–1378 Table 1 Specification
1375
4-Stroke, single cylinder, compression ignition engine with variable compression ratio
Make Rated power Bore and stroke Compression ratio Cylinder capacity Dynamometer Cylinder pressure Orifice dia. Fuel Hydrogen injection
Kirloskar AV-1 3.7 KW, 1500 RPM 85 mm · 110 mm 16.5: 1, variable from 14.3 to 24.5 624. cc Electrical-AC alternator Piezo sensor, range: 145 bar 0.15 mm Diesel, hydrogen By hydrogen injector
exhaust gas analyzer is calibrated periodically using standard calibration gas. Standard Bosch smoke measuring instrument is used to measure the exhaust smoke emission from the engine. In this paper, only the gaseous exhaust emission results are presented to validate the prediction. 4. Results and discussion Figs. 1 and 2 show the change in mole fraction of CO2, forvarious percentage hydrogen substitutions at different constant equivalence ratios. It can be noted that for all equivalence ratios except for 1.2 and 1.4, the mole fraction of CO2 increases between 20% and 40% hydrogen substitution and then decreases with further increase in hydrogen percentage, however for equivalence ratios of 1.2 and 1.4, it increases continuously. As shown in Fig. 3, with higher temperature of 3000 K, the CO2 value is lower than that of 2000 K. CO2 is one of the major products of lean as well as rich combustion. Under the stoichiometric conditions the mole fractions of CO2 is at its peak and decreases when the mixture becomes either richer or leaner due to simultaneous presence of other products. As the temperature increases, the mole fraction of CO2 decreases since the dissociation increases with temperature. Figs. 4 and 5, shows the change in mole fraction of NO for various percentages of hydrogen substitutions for dif-
0.3
Phi = 0.6
0.2
phi=0.9 0.1
phi=1.2
0 0
20
40
60
80
100
% Hydrogen
Fig. 2. Effect of hydrogen substitution on CO2 at T = 2000 K.
ferent constant equivalence ratios for temperatures of 1500 K and 2000 K. With the increase in percentage of hydrogen, the NO decreases slightly. As the value of equivalence ratio is increasing, the mole fraction of NO is decreasing for both temperatures of 1500 K and 2000 K. Similar trend was also noted for 2500 K and 3000 K. As shown in Fig. 6, for 80% Hydrogen substitution, and with 3000 K, the mole fraction of NO is higher than that of other lower temperatures. NO present in the exhaust is very much less as the concentration of N atoms is very much less. The lack of dissociation of N2 molecules is the result of the strong triple covalent bond. Equilibrium concentrations of NO are rather flat and peak in the lean region falling rapidly in the rich region. In most combustion systems, NO levels are well below the equilibrium concentrations
Mole Fraction Vs Equivalence Ratio, 80% Hydrogen Mole Fraction Of CO2
Type
Mole Fraction of CO2
Mole Fraction of CO2 Vs Equivalence Ratio,T=2000k
0.4 T=2000 0.3 T=2500 0.2 T=3000 0.1 0 0
0.5
1
1.5
Equivalence Ratio
Fig. 3. Mole fraction of CO2 against the equivalence ratio at 80% hydrogen and at T = 2000 K, 2500 K, and 3000 K.
Mole Fraction Of NO Vs% Hydrogen , T = 1500
Mole Fraction of CO2 Vs % Hydrogen
0.03
0.3
Phi =0.3
0.25
Phi =0.5
0.2
Phi = 0.8
0.15
Phi = 1.0
0.1
Phi = 1.2
0.05
Phi = 1.4
0
Mole Fraction NO
Mole fraction of CO2
0.35
0.025 0.02
Phi =0.3
0.015
Phi =0.5
0.01
Phi = 0.8
0.005 0
0
20
40
60
80
100
% Hydrogen
Fig. 1. Effect of hydrogen substitution on CO2 at T = 1500 K.
0
50
100
% Hydrogen
Fig. 4. Effect of hydrogen substitution on NO at T = 1500 K.
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M. Masood, M.M. Ishrat / Fuel 87 (2008) 1372–1378 Mole Fraction of N2 Vs% H2, T = 1500 K
Mole Fraction of NO Vs Equivalence Ratio
0.03 20% H2
0.02 40% H2
0.01 60% H2
Mole fraction of N2
Mole Fraction Of NO
0% H2
0 0
0.5
1
1.5
0.9 0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1 0
Fig. 5. Effect of equivalence ratio on NO T = 1500 K.
Phi =0.5 Phi = 0.8 Phi = 1.0 Phi = 1.2 Phi = 1.4
0
80% H2
Equivalence Ratio
Phi =0.3
T=2000
0.02
T=2500
0.015 0.01
T=3000
0.005 0 0
0.5
1
100
Mole Fraction of H2o Vs % Hydrogen, T=1500 k Mole fraction of H2o
Mole Fraction Of NO
0.03
40 60 80 % Hydrogen
Fig. 8. Effect of hydrogen substitution on N2 at T = 1500 K.
Mole Fraction of NO Vs Equivalence Ratio , 80% Hydrogen
0.025
20
1.5
0.8 0.7 0.6 0.5 0.4 0.3 0.2 0.1 0
Phi =0.3 Phi =0.5 Phi = 0.8 Phi = 1.0 Phi = 1.2 Phi = 1.4 0
Equivalence Ratio
20
40
60
80
100
% Hydrogen
Fig. 6. Effect of temperature on NOx. Fig. 9. Effect of hydrogen substitution on H2O at T = 1500 K.
because of relatively slow formation reaction. As the temperature increases, the mole fraction of NO increases due to increase in dissociation of N2 [13]. Fig. 7 shows the results for CO against the percentage hydrogen substitution. It can be noted that as the hydrogen percentage is increasing the mole fraction of CO increases, and for more than 1 equivalence ratio, the increase is very sharp. However its value is lower with higher temperatures. CO is the minor species in the lean combustion and as the mixture gets richer the mole fraction of CO increases due to incomplete combustion of carbon. Further, as the temperature increases CO2 dissociates to form CO and hence mole fraction of CO increases as temperature increases (see Fig. 7). The decreasing behavior of N2 was observed in Fig. 8, and the trend was the same as that was observed in case
Mole Fraction Vs Equivalence Ratio , 80% Hydrogen
Mole Fraction for CO
T=2000 0.08
of NO. As shown in Fig. 9, as the percentage hydrogen increases, the mole fraction of H2O also increases; it has highest value at equivalence ratio of 1. If the equivalence ratio is further increased, mole fraction of H2O falls down considerably. As the mole fraction of H2O increases with hydrogen substitution, it brings down the peak combustion temperature, and hence resultantly the values of NO at higher percentages of hydrogen substitution decrease considerably. Figs. 10–15 show the comparison of predicted and experimental values of CO2, NOx, and CO, respectively with variation of hydrogen substitution and also against the equivalence ratios. The predicted values are in good agreement with the experimental results. This validates the code and the other predictions emerging from the same code.
CO2 Vs % H2 0.3 0.25 0.2
T=2500
0.06 0.04
T=3000
0.02
Experiment Simulated
0.15 0.1 0.05
0 -0.02 0
0.5
1
1.5
Equivalence Ratio
Fig. 7. Effect of temperature on CO.
0 0
20
40 60 80 % Hydrogen
100
Fig. 10. Comparison of simulated and experimental values of CO2.
M. Masood, M.M. Ishrat / Fuel 87 (2008) 1372–1378 CO2 Vs Phi (Equivalence ratio)
0.25 Simulated experimental
0.1
Mole fraction of CO
Mole fraction of CO2
0.3
0.15
CO vs Phi
0.35
0.35
0.2
1377
0.3 0.25 0.2
Experiment Simulated
0.15 0.1 0.05 0
0.05
0
0 0
0.5
1
1.5
Equivalence Vs Phi
0.2 0.4 0.6 0.8 1 1.2 1.4 Phi (Equivalence ratio)
Fig. 15. Comparison of simulated and experimental values of CO.
Fig. 11. Comparison of simulated and experimental results at different equivalence ratios.
5. Conclusions NOx Vs % H2 0.016 0.014 0.012 0.01 Simulated Experiment
0.008 0.006 0.004
The code developed gave reasonably good results which are in good agreement with the experimental values. However, there exist many areas which are unaddressed by the code. At low and high percentages of hydrogen and during transition between diesel and hydrogen the model predictions are not very clear, this eventually shows the limitation of the model and opens the doors for further investigation.
0.002 0 0
20 40 60 80 100 % Hydrogen substitution
Fig. 12. Comparison of simulated and experimental values of NOx.
NOx vs Phi Mole fraction of CO2
0.016 0.014 0.012 0.01 Simulated Experiment
0.008 0.006 0.004 0.002
• For all equivalence ratios except for 1.2 and 1.4, the mole fraction of CO2 increases between 20% and 40% hydrogen substitution and then decreases with further increase in hydrogen percentage. However for equivalence ratios of 1.2 and 1.4, it increases continuously. • As the value of equivalence ratio increases, the mole fraction of NOx decreases for both temperatures of 1500 K and 2000 K. • As the hydrogen percentage increases the mole fraction of CO increases, and for more than 1 equivalence ratio, the increase is sharp. • The mole fraction of H2O increases with hydrogen substitution; this decreased the peak combustion temperatures.
0 0
0.5
1
1.5
References
Equivalence ratio (phi)
Fig. 13. Comparison of simulated and experimental values of NOx With the variations of equivalence ratios.
CO Vs % Hydrogen 0.35 0.3 0.25 0.2
Experiment Simulated
0.15 0.1 0.05 0 0
20
40
60
80
100
% Hydrogen substitution
Fig. 14. Comparison of simulated and experimental values of CO.
[1] Karim GA, Liu T, Jones W. Exhaust emissions from dual fuel engines at light loads. Presented at SAE Paper. SAE 932822; 1993. [2] Yashuhiro Kou T, Yuki Shigeki N, Ryoji K, Takeshi S. Controlling combustion and exhaust emissions in a direct-injection diesel engine dual fueled with natural gas. Presented at SAE Paper. SAE 952436; 1995. [3] Abd-Alla GH, Soliman HA, Badr OA, Abd-Rabbo MF. Energ Convers Manage 2000;41:559–72. [4] Singh S. The effect of fuel injection timing and pilot quantity on the pollutant emissions from a pilot ignited natural gas engine. Masters thesis, The University Of Alabama; 2002. [5] Fraser RA, Siebers DL, Edwards CF. Auto-ignition of methane and natural gas in a simulated diesel environment. Presented at SAE Paper. SAE 2003-01-0755. [6] Naber JD, Siebers DL, Caton JA, Westbrook CK, Di jjulio SS. Natural gas auto-ignition under diesel condition: experiments and chemical kinetics modeling. Presented at SAE Paper. SAE 942034; 1994. [7] Poonia PM, Ramesh A, Gaur RR. Experimental investigation of the factors effecting the performance of a LPG–diesel dual fuel engine.
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SAE Paper No. 99-01-1123, SAE transactions, Journal of Fuels and Lubricants; 1999. [8] Poonia PM, Ramesh A, Gaur RR. The effect of air temperature and pilot fuel quantity on the combustion characteristics of a LPG–diesel dual fuel engine. SAE Paper No. 982455; 1998. [9] Karim GA. The dual fuel engine of the compression ignition type – prospects, problems and solutions – a review. SAE Paper No. 831073; 1983. [10] Karim GA, Jones W, Rain RR. An experimentation of ignition delay period in dual fuel engines.
[11] Gunea C, Razavi MRM, Karim GA. The effect of pilot fuel quantity on dual fuel engine ignition delay. SAE Paper No. 982453; 1998. [12] Daisho Y, et.al. Controlling combustion and exhaust emissions in a direct injection diesel engine dual-fueled with natural gas. SAE Paper No. 952436; 1995. [13] Deshpande NV et.al, Exhaust gas simulation. IE (I) Journal–MC; 2004. [14] Ferguson CR. Internal combustion engines – applied thermo sciences. John Wiley and Sons; 1986.