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Computational combustion and emission analysis of hydrogen–diesel blends with experimental verification
International Journal of Hydrogen Energy 32 (2007) 2539 – 2547 www.elsevier.com/locate/ijhydene
Computational combustion and emission analysis of hydrogen–diesel blends with experimental verification M. Masood ∗ , M.M. Ishrat, A.S. Reddy Mechanical Engineering Department, M.J. College of Engineering and Technology, Hyderabad 500034, India Received 19 May 2006; received in revised form 7 November 2006; accepted 7 November 2006 Available online 26 December 2006
Abstract The paper discusses the effect of blending hydrogen with diesel in different proportions on combustion and emissions. A comparative study was carried out to analyze the effect of direct injection of hydrogen into the combustion chamber with that of induction through the inlet manifold for dual fueling. Percentage of hydrogen substitution varied from 20% to 80%, simultaneously reducing the diesel percentages. CFD analysis of dual fuel combustion and emissions were carried out for both the said methods using the CFD software FLUENT, meshing the combustion chamber was carried out using GAMBIT. The standard combustion and emission models were used in the analysis. In the second part of the paper, the effect of angle of injection in both the methods of hydrogen admission, on performance, combustion and emissions were analyzed. The experimental results were compared with that of simulated values and a good agreement between them was noticed. 䉷 2006 International Association for Hydrogen Energy. Published by Elsevier Ltd. All rights reserved. Keywords: Hydrogen; Diesel; Combustion; Emissions; Heat release rate; Peak pressures
1. Introduction 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 result 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.
∗ Corresponding author.
E-mail address: [email protected] (M. Masood).
Direct injection and port injection have been studied but mostly for SI engine. In case of diesel engine, dual fueling of diesel with LPG [1,2], methane [3,4], natural gas [5–7] and hydrogen–methane combinations were studied. Most research in dual fuel engine has concentrated on defining the extent of dual fueling and its effect on emissions and performance [8,9]. A hydrogen addition to methane has been reported to be effective to promote combustion at homogeneous lean operation (Kido et al., 1994; Shioji et al., 1995). The power output of a direct injected hydrogen engine was 20% more than for a gasoline engine and 42% more than a hydrogen engine using a carburetor. While direct injection solves the problem of pre-ignition in the intake manifold, it does not necessarily prevent pre-ignition within the combustion chamber. In addition, due to the reduced mixing time of the air and fuel in a direct injection engine, the air/fuel mixture can be nonhomogenous. Studies have suggested this can lead to higher NOx emissions than the non-direct injection systems [10]. Direct injection systems require a higher fuel rail pressure than the other methods. The objective of the present work was to analyze the combustion and emissions for Hydrogen–Diesel combination in dual fuel engine. The study is structured by first laying a
0360-3199/$ - see front matter 䉷 2006 International Association for Hydrogen Energy. Published by Elsevier Ltd. All rights reserved. doi:10.1016/j.ijhydene.2006.11.008
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fundamental foundation, through the use of classical conceptual models, for understanding conventional diesel combustion. Diesel is used as the primary fuel and combustion is controlled by adding hydrogen as an additive to control combustion phasing. Hydrogen is added in two ways: 1. By direct injection into the combustion chamber in different proportions (non-homogeneous mixture formation). 2. By induction method, through the inlet manifold (homogenous mixture formation), while the diesel is injected as usual into the combustion chamber. In the first phase of work, CFD analysis was carried out using FLUENT for both conditions and in the second phase experimental work was carried out to analyze and compare the results. 2. Combustion model The standard characteristics—time combustion model calculates the equilibrium concentration of each species and the corresponding laminar and turbulent characteristic times to determine the reaction rates. Seven major combustion species are considered, fuel, O2 , N2 , CO2 , H2 , H2 O and CO. The time rate of change of the concentration of species m is given as [9] ∗ dYm /dt = (Ym − Ym )/c ,
is the local and instantaneous thermodynamic equilibrium concentration and c is the chemical conversion time to achieve equilibrium. The chemical conversion time is calculated as the sum of a kinetic laminar time scale l and a turbulent mixing time scale t as follows: (2)
The variable f is a delay coefficient that simulates the influence of turbulence on the combustion after ignition has occurred [11]. f = 0, prior to combustion and f = 1 when combustion is complete. One or more turbulent flames that are initiated by the ignition of diesel fuel consume the hydrogen gas and air mixture. The main heat release occurs in a thin reaction sheet, i.e., a flame that separates the burned and unburned gases. The flame front is tracked by a level set method that was developed for premixed combustion [12,13]. The turbulent time scale is based on the eddy breakup concept and modeled as [11] t= Cm2 k/,
Cm2 = 1.8(1 − x) + 0.4x,
(4)
x = diesel /(diesel+ hydrogen ),
(5)
where diesel and hydrogen are local partial densities of diesel and hydrogen gas, respectively. The model constant in Eq. (4) was calculated empirically by modeling the present duel fuel Combustion; however pure hydrogen and pure diesel cases were calculated separately. Combustion proceeds whenever turbulence is present (k/ > 0). In premixed flames, the reactants burn as soon as they enter the computational domain, upstream of the flame stabilizer. In our model we are using the eddy dissipation model. The eddy dissipation model requires the products to initiate the reaction i.e.,
(1)
∗ where Ym is the concentration of species m, Ym
c = l + ft .
gas are different and hence the combustion model needs to be modified. It is now required that use of a different turbulent mixing model constant for each fuel based on relative amount of hydrogen gas and diesel fuel in each computational cell would be required. Thus, the model constant Cm2 was modified as
(3)
where Cm2 is a model constant for mixing characteristics in the engine, k is the turbulent kinetic energy and is its dissipation rate. In order to apply the above models to simulate dual fuel combustion, some modifications were needed. In the Hydrogen/diesel, dual fuel engine, it is known that the combustion is initiated by the auto-ignition of the diesel fuel. However, the combustion characteristics of the diesel fuel and Hydrogen
Rm,r = mr Mw,m AB(/k) Rm,r = m,r Mw,m A(/k)
p Yp N j j,r Mw,m
YR R,r Mw,R
,
(6)
,
(7)
where Rm,r is the net rate of production of species m, due to reaction r, YP is the mass fraction of any product species, P, YR is the mass fraction of any particular reactant, R, A is an empirical constant equal to 4.0, and B is an empirical constant equal to 0.5. When we initialize the solution, FLUENT sets the product mass fractions to 0.01, which is sufficient to start the reaction. The local mass fraction of each species Ym , is predicted through the solution of a convection–diffusion equation for the mth species i.e., j(Ym ) + ∇(Ym ) = −∇Jm + Rm + Sm , jt
(8)
where Rm is the net rate of production of species ‘m’ by chemical reaction. Sm is the rate of creation by addition from the dispersed phase. This equation is solved for N −1 species where N is the total number of fluid phase chemical species present in the system. In our model, the diffusion is taken into account by specifying the number of species using the material’s panel [which is part of preprocessing steps using FLUENT] using the above equation, and the premixed combustion is defined using the viscous model, thus both the things occur simultaneously.
M. Masood et al. / International Journal of Hydrogen Energy 32 (2007) 2539 – 2547
The G-equation that has been utilized satisfies the condition that the flame front location should be such that the variable G = 0. The equation is
Table 1 Specifications
jG jG + uj = SL |∇G|, jt jxj
Make Rated power Bore and Stroke Compression ratio Cylinder capacity Dynamometer Cylinder pressure Orifice dia. Fuel Hydrogen injection
(9)
where SL is the laminar flame speed. The rate of consumption of the reactants is SL |∇G|, xj is the Cartesian coordinate, uj is the Cartesian component of the velocity vector v and t is the time. The laminar flame speed in our model is correlated to the pressure, and the variation between our model and the other model is that we are not considering the PREMIX code. The variation of flame curvature is not considered in the G-function of our model but it depends upon the diffusion of the species, which is taken into account during the preprocessing steps of the problem analysis. A full mesh model has been utilized. As we are considering the full model, a coarse mesh has been utilized and a hybrid grid has been considered for the meshing.
3. CFD analysis CFD analysis was carried out for the combustion and emission analysis of hydrogen–diesel dual fuel mode by varying the percentage of hydrogen substitution. Diesel was injected at a pressure of 160 bar, air was inducted at atmospheric temperature and pressure and hydrogen were: (a) Injected through the injector at a pressure of 60 bar directly into the combustion chamber. (b) It was inducted through the inlet manifold and slightly above the atmospheric condition. Diesel is used as the primary fuel and combustion is controlled by adding hydrogen as an additive during the combustion phase. Meshing of the combustion chamber is carried out using GAMBIT, by a tetrahedral element using cooper tool. Hydrogen and diesel were predefined in the fuel selection options along with air. The results obtained were presented for discussion along with experimental verification.
4. Experimental setup 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 and schematic experimental setup is shown in Fig. 1. Diesel was injected with a nozzle of hole size of 0.15 mm, hydrogen was injected by a hydrogen injector, the cross section in Fig. 13 shows the location of injectors. Simultaneous provision is also provided for hydrogen induction through inlet manifold (Fig. 14); this is to compare the effect of injection versus induction. Airflow rate was measured using a laminar flow element.
Type
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4-stroke, single cylinder, compression ignition engine with variable compression ratio 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: 2000 psi 0.15 mm Diesel, hydrogen By injector and by induction method
The engine was coupled to a DC dynamometer and all the experiments were carried out at a constant speed of 1400 rpm. Electric resistance load is applied on the dynamometer through an alternator in percentages of full load. As the load was increased, the rpm decreased, and it was maintained constant by increasing the hydrogen flow rate by a thermal mass flow controller. Crank-angle-resolved in-cylinder pressures and the diesel injection pressures were measured. A computer interfaced piezoelectric sensor, of range 145 bar was used to note the in-cylinder pressures. Pressure signals were obtained at one-degree 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. It is a software developed using C and Matlab for collecting the pressure histories from the piezoelectric transducers and data from the data acquisition system and for storing in the computer. It is programmed to calculate the heat release rate from the data collected using the well-established formulas of heat release rate calculations. The special software stores the data of pressures and volumes corresponding to a particular crank angle location for plotting the P–V and P– curves. The experiments were carried out first by injecting hydrogen at a pressure of 60 bar, and then it was repeated for the same operating conditions by induction into the manifold at atmospheric pressure, while the diesel was directly injected into the cylinder in both the cases at a pressure of 160 bar. Hydrogen flow rate was controlled by a thermal mass flow controller. The airflow rate was measured using a laminar flow element. The engine speed was maintained constant by controlling the hydrogen gas mass flow rate. Engine exhaust emissions were measured using an advanced AVL 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 exhaust gas analyzer is calibrated periodically using standard calibration gas. The hydrocarbons and NOx are measured in terms of parts per million (ppm) as hexane equivalent and carbon monoxide emissions are measured in terms of percentage volume. Standard Bosch smoke measuring instrument is used to measure the exhaust smoke emission from the engine.
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4
13 2
10 8 14
5 1 1
9
12 3 7 6
15
Fig. 1. Schematic diagram of the complete engine test setup with instrumentation. 1. Hydrogren gas cylinder; 2. Flash back arrestor; 3. Fome arrestor; 4. Air tank; 5. H2 injector; 6. Variable compression ratio engine; 7. Dynamometer; 8. PC interfaced data acqusition system; 9. PC interfaced with VCR; 10. Provision to change the compression ratio; 11. Pressure sensor to record Cly Pt; 12. Cable connecting the sensor to PC; 13. Diesel tank connected to injector; 14. TDC pick-up; 15. Exhaust gas analyzer.
5. Results and discussions
Brake thermal efficiency Vs % H2
5.1. Effect of hydrogen substitution on brake thermal efficiency The experimental analysis carried out to study the effect of hydrogen induction through inlet manifold versus that of direct hydrogen injection on brake thermal efficiency is presented in Fig. 2. The brake thermal efficiency increased with the increase in percentage substitution of hydrogen by both the methods; however, the efficiency was higher by around 19% in induction through inlet manifold when compared to that of direct injection method. This is primarily because of uniform mixing of hydrogen and air (by induction method) which formed a homogeneous mixture, burnt completely by the flame initiated by the diesel injection and resulted in complete heat release. The combustion modeling, that has predicted the velocity contours for both methods of hydrogen injection, confirms the experimentally obtained trends. As shown in Figs. 3 and 4, the central core portion, in case of induction method (Fig. 4) represents a homogeneous charge compression ignition behavior, whereas in the case of direct injection, it represents the non turbulent behavior. The flame front travel in case of induction is more rigorous and uniform when compared to rather slow and non-uniform flame front in case of direct injection method especially near the cylinder walls. For induction method, the predicted combustion velocities are shown in Fig. 5, these were at least 23% higher than that of direct injection method. Additionally, as can be seen from the model, the slow rate of flame travel near the cylinder walls may leave some of the fuel unburned, set out as excessive unburned hydrocarbons. The model also highlights very clearly the regions of high and low turbulence in both the cases.
B.thermal efficiency
35 30 25 20 Induction
15
Injection
10 5 0 0
20 40 60 80 % Hydrogen substitution Fig. 2. Efficiency comparisons.
Fig. 3. H2 injection method.
100
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Fig. 4. H2 induction method.
Combustion velocity m/s
Combustion velocity Vs % Hydrogen substitution 800 700 600 500 400 300 200 100 0
Velocity in m/s
10
30 50 70 % Hydrogen substitution
90
Fig. 5. CFD results analysis of combustion velocity vs % hydrogen substitution.
Though the simulation and experimented values show a good agreement, the combustion model fails to predict the transition between diesel pilot and hydrogen gas combustion. The model also does not very clearly predict the performance at very low hydrogen percentage substitutions by both the methods of injection. 5.2. Effect of hydrogen substitution on combustion Pressure signals were obtained at one-degree crank angle intervals using a digital data acquisition system. The average pressure data from 100 consecutive cycles were used for calculating combustion parameters. Pressure–crank angle and heatrelease rate with crank angle diagrams were plotted from the data collected for the combustion analysis as shown in Figs. 10 and 11 for both the methods of hydrogen injection. The combustion analysis from the data collected for dual fuel shows that there are two main stages of combustion in the dual fuel mode like in diesel (single fuel) mode. The first stage is mainly the combustion of diesel, along with the small amount of hydrogen entrained in the diesel spray, while the second phase of combustion is mainly due to the combustion of remaining amount of hydrogen by flame propagation from the ignition centers formed by the diesel spray. The second stage is stronger than
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the diesel mode as most of heat release occurs by the combustion of hydrogen by flame propagation. In the rapid combustion phase, which is around 355◦ in Figs. 10 and 11, it can be seen that combustion becomes more rapid with the increased percentage of hydrogen than that of lower hydrogen substitutions. The rate of pressure rise per crank angle is higher with higher percentages of hydrogen substitutions because of higher flame velocities of hydrogen. With lower hydrogen percentages, the peak pressures are getting shifted more towards the TDC. The rise of pressure per degree of crank angle is more in the dual mode. This is because of the high flammability of hydrogen and rapid combustion. The peak heat release rate, that is, the peak combustion is considerably low at low outputs in the dual fuel mode as compared to that of single diesel mode. This is the reason for the reduced brake thermal efficiency and reduced rate of pressure rise at low outputs in the dual fuel mode when compared to diesel. The rate of heat release increased with hydrogen substitution at high outputs. This is because as the hydrogen percentage is increased, the mixture becomes rich in hydrogen in hydrogen-air combination. This sets the rapid combustion rates, which tends to increase the pressure rapidly in the dual fuel mode. The rate of heat release is higher in dual fuel because of combined burning of two fuels. This is one of the reasons for higher thermal efficiency at high outputs in the dual fuel mode. Fundamentally the combustion analysis was the same for both direct injection and port injection. However in case of direct injection of hydrogen at low loads, the combustion duration increased. This is due to the reduced combustion rate of the primary fuel (diesel)–air mixture which in turn was a result of the reduction in the ignition centers. As the percentage of hydrogen is increased, the rate of pressure rise per crank angle is simultaneously increased. At leaner mixtures, however, the flame velocity decreases significantly. In case of induction of hydrogen through the inlet manifold, it forms more ignition centers because of homogeneous mixture formation which results in faster and complete combustion even at low outputs. This resulted in higher rate of pressure rise per crank angle and higher rate of heat release by induction. The experimental results shows that pressure rise and heat release rate per crank angle in case of induction is around 17% higher than that of injection.
5.3. Effect of hydrogen substitution on NOx The experimental results show an obvious trend of reduction in NOx with increase in hydrogen percentage. This could be because increase in hydrogen substitution simultaneously increases the mole fraction of H2 O, i.e., moisture increases (Fig. 17), which finally brought down the peak temperatures. And hence NOx decreases with the increase in hydrogen substitution. This experimental trend was in conformity with the simulated results for both the methods of hydrogen injection, which can be seen from Fig. 9. Additionally, the CFD analysis carried out revealed that the NOx formation tendency is slightly higher
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B.Thermal Efficiency Vs Injection Angle B.thermal efficiency
35 30 25 20 15 Induction Injection
10 5 0 -50
-40
-30 -20 10 0 10 injection Angle (Deg. CA)
20
Fig. 6. Effect of injection angle on efficiency. Fig. 8. H2 injection.
700 600
CFD Experimental
NOI [pp ]
500 400 300 200 100 0 0
20 40 60 80 100 % hydrogen Substitution
Fig. 9. CFD results analysis of NO vs % hydrogen substitution. Fig. 7. H2 induction method.
5.4. Effect of injection angle on thermal efficiency In the second case as shown in Fig. 6, the study was carried out by changing the angle of injection of diesel. In the induction method, hydrogen was inducted through the inlet manifold and only the diesel injection angle was changed (Figs. 7–10). Similarly in the direct injection method, the injection angles of both hydrogen and diesel were changed. As shown in Fig. 6, in both the methods, the brake thermal efficiency increased first and then in the very advance injection angles it decreases. However, in this case also the induction method had the higher efficiency than that of injection. The flame speed is quite low in non-turbulent mixtures and increases with increasing turbulence. This is mainly due to the additional intermingling of the burning and unburned particles at the flame front which expedites reaction by increasing the rate of contact. Turbulence increases the heat flow to the cylinder walls. It also accelerates the chemical reaction by intimate mixing of fuels and oxygen so that injection advance may be
Heat release rate (deg.\CA)
in case of induction than in direct injection. This is probably because of the reason that the combustion peak temperatures are higher in induction than in injection. This can be seen from Figs. 7 and 8.
Heat release rate Vs Crank Angle 450 400 350 300 250 200 150 100 50 0 330
Injection Induction
340
350 360 370 380 390 Crank Angle (deg.)
400
Fig. 10. Heat release rate comparisons.
reduced. Higher rate of turbulence and complete heat release were the reasons of higher efficiency in induction than in direct injection which will further be discussed in the case of rate of heat release. 5.5. Effect of injection angle on heat release rate As discussed earlier, injection angles were varied in the following ways: (1) In case of induction, only the diesel injection angle was varied.
M. Masood et al. / International Journal of Hydrogen Energy 32 (2007) 2539 – 2547
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Pressure-Crank Angle diagram, 80% H2
Pressure (bar)
100 90 80 70 60 50 40 30 20 10 0
Injection Induction
0
200 400 600 Crank Angle (deg.)
800
Fig. 11. Pressure crank angle diagram comparisons. Fig. 13. Induction of hydrogen through inlet manifold with flame arrestor. 300
Diesel Injector
Hydrogen Injector
Pressure transducer A-A
A
Heat release rate (deg. /CA)
Intake valve
250 200 150 20 deg. 10 deg. 2.5 deg BTDC 3 ATDC
100 50
Exhaust valve Fuel jet
0 -50
A
For the case 1, the effect of injection angles on heat release rate are shown in Fig. 14. The injection timing resulting in the shortest ignition delay yields the highest level of smoke. A shorter ignition delay yields a higher percentage of diffusion burn because less time is available for mixing prior to ignition. As expected, combustion duration is maximized at the same injection timing that ignition delay is minimized. Since diffusion combustion has a slower burn rate than premixed combustion, it substantiates the claim that combustion duration increases as the diffusion burn portion increases. Therefore, smoke increases as the contribution of diffusion burn to the total heat release increases (Figs. 11–14). As the injection timing was earlier, the value of the second peak of heat release becomes smaller, because the rate of premixed combustion increases.
10
30
50
Fig. 14. Heat release rate analysis for induction method. Rate of Heat Release (J/deg.)
(2) In case of direct injection, both the hydrogen and diesel injection angles were varied which were measured in terms of crank angles.
-10
Injection angle (deg. CA)
85 Fig. 12. Piston and head C/s, with fuel injectors, pressure transducer, intake and exhaust valves.
-30
300 250 200 150
20 deg.
100
2.5 deg BTDC 3 ATDC
10 deg.
50 0 -60 -40 -20
0
20
40
60
80
Injection angle (deg.CA)
Fig. 15. Heat release rate analysis for injection method.
The increased ignition delay increases the time for fuel and air mixing, allowing for attainment of premixed compression ignition combustion. Notice that the rate of heat release shifts by much more than 10◦ as timing is retarded from 20◦ BTDC to 10◦ BTDC (Fig. 15). This shift, as well as the decreased rate of heat release, results in combustion temperatures low enough to avoid fuel pyrolysis, and thus soot formation. Whereas during induction of hydrogen (Fig. 14), the same amount of retardation resulted in higher rate of heat release,
M. Masood et al. / International Journal of Hydrogen Energy 32 (2007) 2539 – 2547
900 800 700 600 500 400 300 200 100 0
Mole Fraction of H2O Vs % H2
Injection Induction
-60 -50 -40 -30 -20 -10 0 CAD ATDC
10 20
Mole Fraction of H2O
NOx (ppm)
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0.7 0.6 0.5 0.4
Induction injection
0.3 0.2 0.1 0 0
Fig. 16. Effect of injection angles on NOx .
higher combustion temperatures and soot formation occurred because of fuel pyrolizes during high temperature combustion. 5.6. Effect of injection timing on NOx As ignition is advanced by advancing the fuel injection, peak pressure, rate of pressure and NOx production all increased. NOx decreases monotonically as timing is retarded in both the cases. Initially, smoke increased as expected based on classical diesel combustion behavior (Fig. 18). However, as injection timing is further retarded, smoke begins to drop. As shown in Fig. 18, for the induction method the PM level decreased continuously in contrast to the direct injection where it increased first then decreased. Smoke (or soot) in the exhaust is the difference between soot formation and soot oxidation [14]. Soot formation occurs as fuel pyrolizes during high temperature combustion. Fuel pyrolysis is defined as hydrocarbon chain fragmentation in the absence of oxygen. These fragmentations then develop into nucleation sites for hydrocarbons and sulfates to adhere onto, thus forming the soot particle. Soot oxidation occurs as high temperature gases promote soot burning. When the injection angle was retarded from near TDC, NOx increased and had the maximum value near 23◦ BTDC, which can be seen from Fig. 16; when the injection angle was further retarded, NOx value decreased. This is because, there is no high temperature area due to well mixing of hydrogen, air and diesel. Diesel, which is around 20% in the present case, as 80% is hydrogen substitution. The NOx formation in case of induction method was found to be 33% higher than that of the direct injection method at lower percentages of hydrogen substitutions. However, it was found that as the percentage of hydrogen was increased the value of NOx decreased (Fig. 9), since mole fraction of H2 O increased (Figs. 17 and 18). As discussed above, the induction of hydrogen through the inlet manifold seems more promising than that of direct injection. However, the risk of hydrogen leakage and chances of catching fire in the inlet manifold when mixed with air are some of the drawbacks of induction, whereas the direct injection of hydrogen in CI engine avoids the risk of backfire besides being low specific fuel consumption when compared to induction. The high delivery pressure of hydrogen in direct injection re-
20 40 60 80 % Hydrogen substitution
100
Fig. 17. Effect of H2 on H2 O formation.
1.6 1.4 1.2 1 0.8 0.6 0.4 0.2 0
Induction injection
-30 -20 -10
0 10 20 CAD ATDC
30
40
Fig. 18. Effect of injection angle on PM.
quires additional handling and operating care. The tendency of slightly higher NOx formation especially at lower percentages of hydrogen substitutions in the induction method is one the limiting factor. Practically storage of two fuels separately on board causes additional cost and care hence not very feasible. However for research, it is not a constraint. 6. Conclusions The CFD analysis along with the experimental investigations carried out to compare the hydrogen–diesel dual fuel combustion and emissions by induction and direct injection methods had the following conclusions: This approach with little modification to the existing model has given an acceptable range of results. However there exist many areas which are unaddressed by the model. 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. The results could have been more accurate with higher degree of refinement of the computational mesh. Here, however, the KIVA modeling has got a clear edge both in refinement, flexibility and speed up over the suggested model for co-fueling. The hydrogen–diesel co-fueling will solve the drawback of lean operation of hydrocarbon fuels such as diesel, which are hard to ignite and results in reduced power output, by reducing misfires, improving emissions, performance and fuel economy.
M. Masood et al. / International Journal of Hydrogen Energy 32 (2007) 2539 – 2547
However handling and storing two fuels separately pose practical difficulty. The other conclusions that can be drawn out of the analysis are: • Practically the brake thermal efficiency was around 19% higher in the induction method when compared to that of the direct injection method. • CFD analysis of both the methods show that the combustion velocity increased with higher hydrogen substitutions. • The predicted combustion velocities for the induction method are at least 23% higher than that of the direct injection method. • The CFD analysis carried out revealed that the NOx formation tendency is higher in case of induction than in direct injection. This tendency was confirmed by the practical results obtained. The NOx formation in case of induction was found to be 33% higher than that of the injection method at lower percentages of hydrogen substitutions. • As ignition is advanced by advancing the fuel injection, peak pressure, rate of pressure and NOx production all increased. NOx decreases monotonically as timing is retarded in both the cases. • The experimental results show that pressure rise and heat release rate per crank angle in case of induction is around 17% higher than that of direct injection. References [1] Poonia PM, Ramesh A, Gaur RR. Experimental investigation of the factors affecting the performance of a LPG-diesel dual fuel engine. SAE Paper No. 99-01-1123, SAE transactions. J Fuels Lubricants; 1999.
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[2] 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. [3] 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. [4] 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. [5] Abd-Alla GH, Soliman HA, Badr OA, Abd-Rabbo MF. Energy Conversion Manage 2000;41:559–72. [6] 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. [7] 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. [8] Karim GA, Liu T, Jones W. Exhaust emissions from dual fuel engines at light loads. Presented at SAE. Paper SAE 932822; 1993. [9] Yashuhiro KT, Yuki SN, 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. [10] Hermann, Rottengurber, Ulrichwiebike. Hydrogen diesel engine with direct injection, high power density and low exhaust gas emissions. MTZ Motortechniscie, Zietschrift, Report, 61; 2001. [11] Kong SC, Han ZW, Reitz RD. The development and application of a diesel ignition and combustion model for multidimensional engine simulations. SAE 950278; 1995. [12] Peters N. Turbulent combustion. Cambridge, UK: Cambridge University Press; 2000. [13] Tan Z, Reitz RD. Modeling in spark–ignition engines using a level set method. Presented at SAE. Paper SAE 2003-01-0722. [14] Khan I, Greeves G, Wang C. Factors affecting smoke and gaseous emissions from direct injection engines and a method of calculation. SAE Paper 730169; 1973.
Computational combustion and emission analysis of hydrogen–diesel blends with experimental verification M. Masood ∗ , M.M. Ishrat, A.S. Reddy Mechanical Engineering Department, M.J. College of Engineering and Technology, Hyderabad 500034, India Received 19 May 2006; received in revised form 7 November 2006; accepted 7 November 2006 Available online 26 December 2006
Abstract The paper discusses the effect of blending hydrogen with diesel in different proportions on combustion and emissions. A comparative study was carried out to analyze the effect of direct injection of hydrogen into the combustion chamber with that of induction through the inlet manifold for dual fueling. Percentage of hydrogen substitution varied from 20% to 80%, simultaneously reducing the diesel percentages. CFD analysis of dual fuel combustion and emissions were carried out for both the said methods using the CFD software FLUENT, meshing the combustion chamber was carried out using GAMBIT. The standard combustion and emission models were used in the analysis. In the second part of the paper, the effect of angle of injection in both the methods of hydrogen admission, on performance, combustion and emissions were analyzed. The experimental results were compared with that of simulated values and a good agreement between them was noticed. 䉷 2006 International Association for Hydrogen Energy. Published by Elsevier Ltd. All rights reserved. Keywords: Hydrogen; Diesel; Combustion; Emissions; Heat release rate; Peak pressures
1. Introduction 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 result 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.
∗ Corresponding author.
E-mail address: [email protected] (M. Masood).
Direct injection and port injection have been studied but mostly for SI engine. In case of diesel engine, dual fueling of diesel with LPG [1,2], methane [3,4], natural gas [5–7] and hydrogen–methane combinations were studied. Most research in dual fuel engine has concentrated on defining the extent of dual fueling and its effect on emissions and performance [8,9]. A hydrogen addition to methane has been reported to be effective to promote combustion at homogeneous lean operation (Kido et al., 1994; Shioji et al., 1995). The power output of a direct injected hydrogen engine was 20% more than for a gasoline engine and 42% more than a hydrogen engine using a carburetor. While direct injection solves the problem of pre-ignition in the intake manifold, it does not necessarily prevent pre-ignition within the combustion chamber. In addition, due to the reduced mixing time of the air and fuel in a direct injection engine, the air/fuel mixture can be nonhomogenous. Studies have suggested this can lead to higher NOx emissions than the non-direct injection systems [10]. Direct injection systems require a higher fuel rail pressure than the other methods. The objective of the present work was to analyze the combustion and emissions for Hydrogen–Diesel combination in dual fuel engine. The study is structured by first laying a
0360-3199/$ - see front matter 䉷 2006 International Association for Hydrogen Energy. Published by Elsevier Ltd. All rights reserved. doi:10.1016/j.ijhydene.2006.11.008
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fundamental foundation, through the use of classical conceptual models, for understanding conventional diesel combustion. Diesel is used as the primary fuel and combustion is controlled by adding hydrogen as an additive to control combustion phasing. Hydrogen is added in two ways: 1. By direct injection into the combustion chamber in different proportions (non-homogeneous mixture formation). 2. By induction method, through the inlet manifold (homogenous mixture formation), while the diesel is injected as usual into the combustion chamber. In the first phase of work, CFD analysis was carried out using FLUENT for both conditions and in the second phase experimental work was carried out to analyze and compare the results. 2. Combustion model The standard characteristics—time combustion model calculates the equilibrium concentration of each species and the corresponding laminar and turbulent characteristic times to determine the reaction rates. Seven major combustion species are considered, fuel, O2 , N2 , CO2 , H2 , H2 O and CO. The time rate of change of the concentration of species m is given as [9] ∗ dYm /dt = (Ym − Ym )/c ,
is the local and instantaneous thermodynamic equilibrium concentration and c is the chemical conversion time to achieve equilibrium. The chemical conversion time is calculated as the sum of a kinetic laminar time scale l and a turbulent mixing time scale t as follows: (2)
The variable f is a delay coefficient that simulates the influence of turbulence on the combustion after ignition has occurred [11]. f = 0, prior to combustion and f = 1 when combustion is complete. One or more turbulent flames that are initiated by the ignition of diesel fuel consume the hydrogen gas and air mixture. The main heat release occurs in a thin reaction sheet, i.e., a flame that separates the burned and unburned gases. The flame front is tracked by a level set method that was developed for premixed combustion [12,13]. The turbulent time scale is based on the eddy breakup concept and modeled as [11] t= Cm2 k/,
Cm2 = 1.8(1 − x) + 0.4x,
(4)
x = diesel /(diesel+ hydrogen ),
(5)
where diesel and hydrogen are local partial densities of diesel and hydrogen gas, respectively. The model constant in Eq. (4) was calculated empirically by modeling the present duel fuel Combustion; however pure hydrogen and pure diesel cases were calculated separately. Combustion proceeds whenever turbulence is present (k/ > 0). In premixed flames, the reactants burn as soon as they enter the computational domain, upstream of the flame stabilizer. In our model we are using the eddy dissipation model. The eddy dissipation model requires the products to initiate the reaction i.e.,
(1)
∗ where Ym is the concentration of species m, Ym
c = l + ft .
gas are different and hence the combustion model needs to be modified. It is now required that use of a different turbulent mixing model constant for each fuel based on relative amount of hydrogen gas and diesel fuel in each computational cell would be required. Thus, the model constant Cm2 was modified as
(3)
where Cm2 is a model constant for mixing characteristics in the engine, k is the turbulent kinetic energy and is its dissipation rate. In order to apply the above models to simulate dual fuel combustion, some modifications were needed. In the Hydrogen/diesel, dual fuel engine, it is known that the combustion is initiated by the auto-ignition of the diesel fuel. However, the combustion characteristics of the diesel fuel and Hydrogen
Rm,r = mr Mw,m AB(/k) Rm,r = m,r Mw,m A(/k)
p Yp N j j,r Mw,m
YR R,r Mw,R
,
(6)
,
(7)
where Rm,r is the net rate of production of species m, due to reaction r, YP is the mass fraction of any product species, P, YR is the mass fraction of any particular reactant, R, A is an empirical constant equal to 4.0, and B is an empirical constant equal to 0.5. When we initialize the solution, FLUENT sets the product mass fractions to 0.01, which is sufficient to start the reaction. The local mass fraction of each species Ym , is predicted through the solution of a convection–diffusion equation for the mth species i.e., j(Ym ) + ∇(Ym ) = −∇Jm + Rm + Sm , jt
(8)
where Rm is the net rate of production of species ‘m’ by chemical reaction. Sm is the rate of creation by addition from the dispersed phase. This equation is solved for N −1 species where N is the total number of fluid phase chemical species present in the system. In our model, the diffusion is taken into account by specifying the number of species using the material’s panel [which is part of preprocessing steps using FLUENT] using the above equation, and the premixed combustion is defined using the viscous model, thus both the things occur simultaneously.
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The G-equation that has been utilized satisfies the condition that the flame front location should be such that the variable G = 0. The equation is
Table 1 Specifications
jG jG + uj = SL |∇G|, jt jxj
Make Rated power Bore and Stroke Compression ratio Cylinder capacity Dynamometer Cylinder pressure Orifice dia. Fuel Hydrogen injection
(9)
where SL is the laminar flame speed. The rate of consumption of the reactants is SL |∇G|, xj is the Cartesian coordinate, uj is the Cartesian component of the velocity vector v and t is the time. The laminar flame speed in our model is correlated to the pressure, and the variation between our model and the other model is that we are not considering the PREMIX code. The variation of flame curvature is not considered in the G-function of our model but it depends upon the diffusion of the species, which is taken into account during the preprocessing steps of the problem analysis. A full mesh model has been utilized. As we are considering the full model, a coarse mesh has been utilized and a hybrid grid has been considered for the meshing.
3. CFD analysis CFD analysis was carried out for the combustion and emission analysis of hydrogen–diesel dual fuel mode by varying the percentage of hydrogen substitution. Diesel was injected at a pressure of 160 bar, air was inducted at atmospheric temperature and pressure and hydrogen were: (a) Injected through the injector at a pressure of 60 bar directly into the combustion chamber. (b) It was inducted through the inlet manifold and slightly above the atmospheric condition. Diesel is used as the primary fuel and combustion is controlled by adding hydrogen as an additive during the combustion phase. Meshing of the combustion chamber is carried out using GAMBIT, by a tetrahedral element using cooper tool. Hydrogen and diesel were predefined in the fuel selection options along with air. The results obtained were presented for discussion along with experimental verification.
4. Experimental setup 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 and schematic experimental setup is shown in Fig. 1. Diesel was injected with a nozzle of hole size of 0.15 mm, hydrogen was injected by a hydrogen injector, the cross section in Fig. 13 shows the location of injectors. Simultaneous provision is also provided for hydrogen induction through inlet manifold (Fig. 14); this is to compare the effect of injection versus induction. Airflow rate was measured using a laminar flow element.
Type
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4-stroke, single cylinder, compression ignition engine with variable compression ratio 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: 2000 psi 0.15 mm Diesel, hydrogen By injector and by induction method
The engine was coupled to a DC dynamometer and all the experiments were carried out at a constant speed of 1400 rpm. Electric resistance load is applied on the dynamometer through an alternator in percentages of full load. As the load was increased, the rpm decreased, and it was maintained constant by increasing the hydrogen flow rate by a thermal mass flow controller. Crank-angle-resolved in-cylinder pressures and the diesel injection pressures were measured. A computer interfaced piezoelectric sensor, of range 145 bar was used to note the in-cylinder pressures. Pressure signals were obtained at one-degree 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. It is a software developed using C and Matlab for collecting the pressure histories from the piezoelectric transducers and data from the data acquisition system and for storing in the computer. It is programmed to calculate the heat release rate from the data collected using the well-established formulas of heat release rate calculations. The special software stores the data of pressures and volumes corresponding to a particular crank angle location for plotting the P–V and P– curves. The experiments were carried out first by injecting hydrogen at a pressure of 60 bar, and then it was repeated for the same operating conditions by induction into the manifold at atmospheric pressure, while the diesel was directly injected into the cylinder in both the cases at a pressure of 160 bar. Hydrogen flow rate was controlled by a thermal mass flow controller. The airflow rate was measured using a laminar flow element. The engine speed was maintained constant by controlling the hydrogen gas mass flow rate. Engine exhaust emissions were measured using an advanced AVL 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 exhaust gas analyzer is calibrated periodically using standard calibration gas. The hydrocarbons and NOx are measured in terms of parts per million (ppm) as hexane equivalent and carbon monoxide emissions are measured in terms of percentage volume. Standard Bosch smoke measuring instrument is used to measure the exhaust smoke emission from the engine.
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4
13 2
10 8 14
5 1 1
9
12 3 7 6
15
Fig. 1. Schematic diagram of the complete engine test setup with instrumentation. 1. Hydrogren gas cylinder; 2. Flash back arrestor; 3. Fome arrestor; 4. Air tank; 5. H2 injector; 6. Variable compression ratio engine; 7. Dynamometer; 8. PC interfaced data acqusition system; 9. PC interfaced with VCR; 10. Provision to change the compression ratio; 11. Pressure sensor to record Cly Pt; 12. Cable connecting the sensor to PC; 13. Diesel tank connected to injector; 14. TDC pick-up; 15. Exhaust gas analyzer.
5. Results and discussions
Brake thermal efficiency Vs % H2
5.1. Effect of hydrogen substitution on brake thermal efficiency The experimental analysis carried out to study the effect of hydrogen induction through inlet manifold versus that of direct hydrogen injection on brake thermal efficiency is presented in Fig. 2. The brake thermal efficiency increased with the increase in percentage substitution of hydrogen by both the methods; however, the efficiency was higher by around 19% in induction through inlet manifold when compared to that of direct injection method. This is primarily because of uniform mixing of hydrogen and air (by induction method) which formed a homogeneous mixture, burnt completely by the flame initiated by the diesel injection and resulted in complete heat release. The combustion modeling, that has predicted the velocity contours for both methods of hydrogen injection, confirms the experimentally obtained trends. As shown in Figs. 3 and 4, the central core portion, in case of induction method (Fig. 4) represents a homogeneous charge compression ignition behavior, whereas in the case of direct injection, it represents the non turbulent behavior. The flame front travel in case of induction is more rigorous and uniform when compared to rather slow and non-uniform flame front in case of direct injection method especially near the cylinder walls. For induction method, the predicted combustion velocities are shown in Fig. 5, these were at least 23% higher than that of direct injection method. Additionally, as can be seen from the model, the slow rate of flame travel near the cylinder walls may leave some of the fuel unburned, set out as excessive unburned hydrocarbons. The model also highlights very clearly the regions of high and low turbulence in both the cases.
B.thermal efficiency
35 30 25 20 Induction
15
Injection
10 5 0 0
20 40 60 80 % Hydrogen substitution Fig. 2. Efficiency comparisons.
Fig. 3. H2 injection method.
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Fig. 4. H2 induction method.
Combustion velocity m/s
Combustion velocity Vs % Hydrogen substitution 800 700 600 500 400 300 200 100 0
Velocity in m/s
10
30 50 70 % Hydrogen substitution
90
Fig. 5. CFD results analysis of combustion velocity vs % hydrogen substitution.
Though the simulation and experimented values show a good agreement, the combustion model fails to predict the transition between diesel pilot and hydrogen gas combustion. The model also does not very clearly predict the performance at very low hydrogen percentage substitutions by both the methods of injection. 5.2. Effect of hydrogen substitution on combustion Pressure signals were obtained at one-degree crank angle intervals using a digital data acquisition system. The average pressure data from 100 consecutive cycles were used for calculating combustion parameters. Pressure–crank angle and heatrelease rate with crank angle diagrams were plotted from the data collected for the combustion analysis as shown in Figs. 10 and 11 for both the methods of hydrogen injection. The combustion analysis from the data collected for dual fuel shows that there are two main stages of combustion in the dual fuel mode like in diesel (single fuel) mode. The first stage is mainly the combustion of diesel, along with the small amount of hydrogen entrained in the diesel spray, while the second phase of combustion is mainly due to the combustion of remaining amount of hydrogen by flame propagation from the ignition centers formed by the diesel spray. The second stage is stronger than
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the diesel mode as most of heat release occurs by the combustion of hydrogen by flame propagation. In the rapid combustion phase, which is around 355◦ in Figs. 10 and 11, it can be seen that combustion becomes more rapid with the increased percentage of hydrogen than that of lower hydrogen substitutions. The rate of pressure rise per crank angle is higher with higher percentages of hydrogen substitutions because of higher flame velocities of hydrogen. With lower hydrogen percentages, the peak pressures are getting shifted more towards the TDC. The rise of pressure per degree of crank angle is more in the dual mode. This is because of the high flammability of hydrogen and rapid combustion. The peak heat release rate, that is, the peak combustion is considerably low at low outputs in the dual fuel mode as compared to that of single diesel mode. This is the reason for the reduced brake thermal efficiency and reduced rate of pressure rise at low outputs in the dual fuel mode when compared to diesel. The rate of heat release increased with hydrogen substitution at high outputs. This is because as the hydrogen percentage is increased, the mixture becomes rich in hydrogen in hydrogen-air combination. This sets the rapid combustion rates, which tends to increase the pressure rapidly in the dual fuel mode. The rate of heat release is higher in dual fuel because of combined burning of two fuels. This is one of the reasons for higher thermal efficiency at high outputs in the dual fuel mode. Fundamentally the combustion analysis was the same for both direct injection and port injection. However in case of direct injection of hydrogen at low loads, the combustion duration increased. This is due to the reduced combustion rate of the primary fuel (diesel)–air mixture which in turn was a result of the reduction in the ignition centers. As the percentage of hydrogen is increased, the rate of pressure rise per crank angle is simultaneously increased. At leaner mixtures, however, the flame velocity decreases significantly. In case of induction of hydrogen through the inlet manifold, it forms more ignition centers because of homogeneous mixture formation which results in faster and complete combustion even at low outputs. This resulted in higher rate of pressure rise per crank angle and higher rate of heat release by induction. The experimental results shows that pressure rise and heat release rate per crank angle in case of induction is around 17% higher than that of injection.
5.3. Effect of hydrogen substitution on NOx The experimental results show an obvious trend of reduction in NOx with increase in hydrogen percentage. This could be because increase in hydrogen substitution simultaneously increases the mole fraction of H2 O, i.e., moisture increases (Fig. 17), which finally brought down the peak temperatures. And hence NOx decreases with the increase in hydrogen substitution. This experimental trend was in conformity with the simulated results for both the methods of hydrogen injection, which can be seen from Fig. 9. Additionally, the CFD analysis carried out revealed that the NOx formation tendency is slightly higher
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B.Thermal Efficiency Vs Injection Angle B.thermal efficiency
35 30 25 20 15 Induction Injection
10 5 0 -50
-40
-30 -20 10 0 10 injection Angle (Deg. CA)
20
Fig. 6. Effect of injection angle on efficiency. Fig. 8. H2 injection.
700 600
CFD Experimental
NOI [pp ]
500 400 300 200 100 0 0
20 40 60 80 100 % hydrogen Substitution
Fig. 9. CFD results analysis of NO vs % hydrogen substitution. Fig. 7. H2 induction method.
5.4. Effect of injection angle on thermal efficiency In the second case as shown in Fig. 6, the study was carried out by changing the angle of injection of diesel. In the induction method, hydrogen was inducted through the inlet manifold and only the diesel injection angle was changed (Figs. 7–10). Similarly in the direct injection method, the injection angles of both hydrogen and diesel were changed. As shown in Fig. 6, in both the methods, the brake thermal efficiency increased first and then in the very advance injection angles it decreases. However, in this case also the induction method had the higher efficiency than that of injection. The flame speed is quite low in non-turbulent mixtures and increases with increasing turbulence. This is mainly due to the additional intermingling of the burning and unburned particles at the flame front which expedites reaction by increasing the rate of contact. Turbulence increases the heat flow to the cylinder walls. It also accelerates the chemical reaction by intimate mixing of fuels and oxygen so that injection advance may be
Heat release rate (deg.\CA)
in case of induction than in direct injection. This is probably because of the reason that the combustion peak temperatures are higher in induction than in injection. This can be seen from Figs. 7 and 8.
Heat release rate Vs Crank Angle 450 400 350 300 250 200 150 100 50 0 330
Injection Induction
340
350 360 370 380 390 Crank Angle (deg.)
400
Fig. 10. Heat release rate comparisons.
reduced. Higher rate of turbulence and complete heat release were the reasons of higher efficiency in induction than in direct injection which will further be discussed in the case of rate of heat release. 5.5. Effect of injection angle on heat release rate As discussed earlier, injection angles were varied in the following ways: (1) In case of induction, only the diesel injection angle was varied.
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Pressure-Crank Angle diagram, 80% H2
Pressure (bar)
100 90 80 70 60 50 40 30 20 10 0
Injection Induction
0
200 400 600 Crank Angle (deg.)
800
Fig. 11. Pressure crank angle diagram comparisons. Fig. 13. Induction of hydrogen through inlet manifold with flame arrestor. 300
Diesel Injector
Hydrogen Injector
Pressure transducer A-A
A
Heat release rate (deg. /CA)
Intake valve
250 200 150 20 deg. 10 deg. 2.5 deg BTDC 3 ATDC
100 50
Exhaust valve Fuel jet
0 -50
A
For the case 1, the effect of injection angles on heat release rate are shown in Fig. 14. The injection timing resulting in the shortest ignition delay yields the highest level of smoke. A shorter ignition delay yields a higher percentage of diffusion burn because less time is available for mixing prior to ignition. As expected, combustion duration is maximized at the same injection timing that ignition delay is minimized. Since diffusion combustion has a slower burn rate than premixed combustion, it substantiates the claim that combustion duration increases as the diffusion burn portion increases. Therefore, smoke increases as the contribution of diffusion burn to the total heat release increases (Figs. 11–14). As the injection timing was earlier, the value of the second peak of heat release becomes smaller, because the rate of premixed combustion increases.
10
30
50
Fig. 14. Heat release rate analysis for induction method. Rate of Heat Release (J/deg.)
(2) In case of direct injection, both the hydrogen and diesel injection angles were varied which were measured in terms of crank angles.
-10
Injection angle (deg. CA)
85 Fig. 12. Piston and head C/s, with fuel injectors, pressure transducer, intake and exhaust valves.
-30
300 250 200 150
20 deg.
100
2.5 deg BTDC 3 ATDC
10 deg.
50 0 -60 -40 -20
0
20
40
60
80
Injection angle (deg.CA)
Fig. 15. Heat release rate analysis for injection method.
The increased ignition delay increases the time for fuel and air mixing, allowing for attainment of premixed compression ignition combustion. Notice that the rate of heat release shifts by much more than 10◦ as timing is retarded from 20◦ BTDC to 10◦ BTDC (Fig. 15). This shift, as well as the decreased rate of heat release, results in combustion temperatures low enough to avoid fuel pyrolysis, and thus soot formation. Whereas during induction of hydrogen (Fig. 14), the same amount of retardation resulted in higher rate of heat release,
M. Masood et al. / International Journal of Hydrogen Energy 32 (2007) 2539 – 2547
900 800 700 600 500 400 300 200 100 0
Mole Fraction of H2O Vs % H2
Injection Induction
-60 -50 -40 -30 -20 -10 0 CAD ATDC
10 20
Mole Fraction of H2O
NOx (ppm)
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0.7 0.6 0.5 0.4
Induction injection
0.3 0.2 0.1 0 0
Fig. 16. Effect of injection angles on NOx .
higher combustion temperatures and soot formation occurred because of fuel pyrolizes during high temperature combustion. 5.6. Effect of injection timing on NOx As ignition is advanced by advancing the fuel injection, peak pressure, rate of pressure and NOx production all increased. NOx decreases monotonically as timing is retarded in both the cases. Initially, smoke increased as expected based on classical diesel combustion behavior (Fig. 18). However, as injection timing is further retarded, smoke begins to drop. As shown in Fig. 18, for the induction method the PM level decreased continuously in contrast to the direct injection where it increased first then decreased. Smoke (or soot) in the exhaust is the difference between soot formation and soot oxidation [14]. Soot formation occurs as fuel pyrolizes during high temperature combustion. Fuel pyrolysis is defined as hydrocarbon chain fragmentation in the absence of oxygen. These fragmentations then develop into nucleation sites for hydrocarbons and sulfates to adhere onto, thus forming the soot particle. Soot oxidation occurs as high temperature gases promote soot burning. When the injection angle was retarded from near TDC, NOx increased and had the maximum value near 23◦ BTDC, which can be seen from Fig. 16; when the injection angle was further retarded, NOx value decreased. This is because, there is no high temperature area due to well mixing of hydrogen, air and diesel. Diesel, which is around 20% in the present case, as 80% is hydrogen substitution. The NOx formation in case of induction method was found to be 33% higher than that of the direct injection method at lower percentages of hydrogen substitutions. However, it was found that as the percentage of hydrogen was increased the value of NOx decreased (Fig. 9), since mole fraction of H2 O increased (Figs. 17 and 18). As discussed above, the induction of hydrogen through the inlet manifold seems more promising than that of direct injection. However, the risk of hydrogen leakage and chances of catching fire in the inlet manifold when mixed with air are some of the drawbacks of induction, whereas the direct injection of hydrogen in CI engine avoids the risk of backfire besides being low specific fuel consumption when compared to induction. The high delivery pressure of hydrogen in direct injection re-
20 40 60 80 % Hydrogen substitution
100
Fig. 17. Effect of H2 on H2 O formation.
1.6 1.4 1.2 1 0.8 0.6 0.4 0.2 0
Induction injection
-30 -20 -10
0 10 20 CAD ATDC
30
40
Fig. 18. Effect of injection angle on PM.
quires additional handling and operating care. The tendency of slightly higher NOx formation especially at lower percentages of hydrogen substitutions in the induction method is one the limiting factor. Practically storage of two fuels separately on board causes additional cost and care hence not very feasible. However for research, it is not a constraint. 6. Conclusions The CFD analysis along with the experimental investigations carried out to compare the hydrogen–diesel dual fuel combustion and emissions by induction and direct injection methods had the following conclusions: This approach with little modification to the existing model has given an acceptable range of results. However there exist many areas which are unaddressed by the model. 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. The results could have been more accurate with higher degree of refinement of the computational mesh. Here, however, the KIVA modeling has got a clear edge both in refinement, flexibility and speed up over the suggested model for co-fueling. The hydrogen–diesel co-fueling will solve the drawback of lean operation of hydrocarbon fuels such as diesel, which are hard to ignite and results in reduced power output, by reducing misfires, improving emissions, performance and fuel economy.
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However handling and storing two fuels separately pose practical difficulty. The other conclusions that can be drawn out of the analysis are: • Practically the brake thermal efficiency was around 19% higher in the induction method when compared to that of the direct injection method. • CFD analysis of both the methods show that the combustion velocity increased with higher hydrogen substitutions. • The predicted combustion velocities for the induction method are at least 23% higher than that of the direct injection method. • The CFD analysis carried out revealed that the NOx formation tendency is higher in case of induction than in direct injection. This tendency was confirmed by the practical results obtained. The NOx formation in case of induction was found to be 33% higher than that of the injection method at lower percentages of hydrogen substitutions. • As ignition is advanced by advancing the fuel injection, peak pressure, rate of pressure and NOx production all increased. NOx decreases monotonically as timing is retarded in both the cases. • The experimental results show that pressure rise and heat release rate per crank angle in case of induction is around 17% higher than that of direct injection. References [1] Poonia PM, Ramesh A, Gaur RR. Experimental investigation of the factors affecting the performance of a LPG-diesel dual fuel engine. SAE Paper No. 99-01-1123, SAE transactions. J Fuels Lubricants; 1999.
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[2] 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. [3] 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. [4] 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. [5] Abd-Alla GH, Soliman HA, Badr OA, Abd-Rabbo MF. Energy Conversion Manage 2000;41:559–72. [6] 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. [7] 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. [8] Karim GA, Liu T, Jones W. Exhaust emissions from dual fuel engines at light loads. Presented at SAE. Paper SAE 932822; 1993. [9] Yashuhiro KT, Yuki SN, 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. [10] Hermann, Rottengurber, Ulrichwiebike. Hydrogen diesel engine with direct injection, high power density and low exhaust gas emissions. MTZ Motortechniscie, Zietschrift, Report, 61; 2001. [11] Kong SC, Han ZW, Reitz RD. The development and application of a diesel ignition and combustion model for multidimensional engine simulations. SAE 950278; 1995. [12] Peters N. Turbulent combustion. Cambridge, UK: Cambridge University Press; 2000. [13] Tan Z, Reitz RD. Modeling in spark–ignition engines using a level set method. Presented at SAE. Paper SAE 2003-01-0722. [14] Khan I, Greeves G, Wang C. Factors affecting smoke and gaseous emissions from direct injection engines and a method of calculation. SAE Paper 730169; 1973.