Experimental Investigation in Optimizing the Hydrogen Fuel on a

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Experimental Investigation in Optimizing the Hydrogen Fuel on a Hydrogen Diesel Dual-Fuel Engine N. Saravanan*,† and G. Nagarajan‡ Tata Motors Limited, Pimpri, Pune-411019, India and Department of Mechanical Engineering, Internal Combustion Engineering DiVision, Anna UniVersity Chennai, Chennai 600025, India ReceiVed NoVember 5, 2008. ReVised Manuscript ReceiVed March 10, 2009

During the past decade the use of alternative fuels for diesel engine has received considerable attention. The interdependence and uncertainty of petroleum-based fuel availability and environmental issues, most notably air pollution, are among the principal forces behind the movement toward alternative sources of energy. Several alternative fuels are available, but all of them are hydrocarbon-based fuels, which cannot eliminate the net carbon emissions. One alternative is to make use of a non-carbon fuel like hydrogen. In the present investigation, hydrogen was used in a diesel engine in the dual-fuel mode with diesel as a primary fuel. Experiments were conducted to determine the optimized injection timing, injection duration, and hydrogen flow rate. From the results it is observed that the optimum timing in port injection is 5° before gas exchange top dead center (BGTDC) with an injection duration of 30° crank angle (CA) and in manifold injection at gas exchange top dead center (GTDC) with an injection duration of 30° CA. Hydrogen flow rate was varied from 2 to 9.5 lpm with above the above-optimized conditions for both port and manifold injection. The optimized hydrogen flow rate was found to be 7.5 lpm for both port and manifold injection. Flow rates higher than 9.5 lpm shows an improvement in performance and reduction in emissions, but the onset of knock was observed; hence, the flow rate was limited to 9.5 lpm. At 75% load the brake thermal efficiency increases by 21% in port injection and 18% in manifold injection. NOX emission is reduced by 2% in port injection and 4% in manifold injection compared to diesel at full load. At full load, smoke is reduced by 45% in both port injection and manifold injection. In the entire load spectra a reduction in CO by about 50% is noticed in both port and manifold injection. Ignition delay or a delay period is found to be 11° or 1.22 ms for diesel and 10° or 1.11 ms in both port and manifold injection.

Introduction Diesel engines are the most trusted power sources in the automobile industry; at the same time it is subjected to stringent emission norms. In addition, the rapid depletion of fossil sources has paved the way to think of alternative fuels.1,2 Emissions (anthropogenic) from automobiles are currently a dominant source of air pollution, representing 54% of carbon monoxide (CO), 38% of oxides of nitrogen (NOX), and 38% of hydrocarbon emissions globally.2 In addition, 7% of the manmade carbon dioxide (CO2) emissions from automobiles adds to the greenhouse effect, which results in global warming. In order to reduce these harmful pollutants several alternative fuels have been attempted, and hydrogen is one of them. Hydrogen on burning produces only water as its product, which is given in eq 1. Hydrogen is a nontoxic, nonodorant fuel and results in complete combustion. H2 + 0.5O2 f H2O

(1)

* To whom correspondence should be addressed. Phone: +912066134688. E-mail: [email protected], [email protected]. † ERC Engines. ‡ Anna University Chennai. (1) Barreto, L.; Makihira, A.; Riahi, K. The hydrogen economy in the 21st century: a sustainable development scenario. Int. J. Hydrogen Energy 2003, 28, 267–284. (2) Zuohua, H.; Jinhua, W.; Bing, L.; Ke, Z.; Jinrong, Y.; Deming, J. Combustion Characteristics of a Direct-Injection Engine Fueled with Natural Gas-Hydrogen Mixtures. Energy Fuels 2006, 20 (2), 540–546.

Hydrogen is a carbon-free alternative fuel. Hence, formation of hydrocarbon, carbon monoxide, and carbon dioxide during combustion can be completely avoided; however, a trace amount of these compounds may be formed due to the partial burning of lubricating oil in the combustion chamber.3 NOX are one of the major pollutants in hydrogen-operated spark ignition (SI) and compression ignition (CI) engines. Hydrogen operation results in achieving higher brake thermal efficiency and in lower levels of exhaust emissions except NOX emissions.4 Several works have been done earlier on hydrogen. Das5 observed that a hydrogen engine develops lower power mainly due to its low volumetric energy density. Various fuel induction techniques such as carburation, continuous manifold injection, timed manifold injection (TMI), and direct cylinder injection were investigated, and TMI was adopted by the author since it gave a higher thermal efficiency and avoided undesirable combustion. It was observed that under stoichiometric condition, hydrogen occupies 29.6% by volume whereas a gasoline-air mixture occupies only about 2% by volume. Yi et al.6 carried out work on both port injection and in-cylinder injection-type hydrogen (3) Das, L. M. Near-term introduction of hydrogen engines for automotive and agriculture application. Int. J. Hydrogen Energy 2002, 27, 479– 487. (4) Das, L. M. Hydrogen engine: research and development (R&D) programmes in Indian Institute of Technology (IIT), Delhi. Int. J. Hydrogen Energy 2002, 27, 953–965. (5) Das, L. M. On board hydrogen storage systems for automotive application. Int. J. Hydrogen Energy 1996, 21 (9), 789–800.

10.1021/ef800962k CCC: $40.75  2009 American Chemical Society Published on Web 04/10/2009

Optimizing the Hydrogen Fuel

Energy & Fuels, Vol. 23, 2009 2647

Table 1. Properties of Hydrogen properties 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15

formula auto ignition temperature (K) minimum ignition energy (MJ) flammability limits (vol % in air) stoichiometric air fuel ratio on mass basis molecular weight (g/mol) density at 16 °C and 1.01 bar (kg/m3) net heating value (lower; MJ/kg) flame velocity (cm/s) quenching gap in NTP air (cm) diffusivity in air (cm2/S) octane number (research) cetane number boiling point (K) viscosity at 15.5 °C (centipoise) vapor pressure at 38 °C (kPa)

diesel

hydrogen

CnH1.8nC8-C20 H2 530 858 0.02 0.7-5 4-75 14.5 34.3 170 2.016 833-881 0.0838 42.5 119.93 30 265-325 0.064 0.63 30 130 40-55 436-672 20.27 2.6-4.1 negligible

fuel supply systems. Their results indicated that the thermal efficiency of the intake port injection was higher than in-cylinder injection at all equivalence ratios. In order to minimize the possibility of flashback occurrence, the injection timing of the hydrogen was fixed in coordination with the intake valve opening timing. Hydrogen used in the dual-fuel mode with diesel by Masood et al.7 showed the highest brake thermal efficiency of 30% at a compression ratio of 24.5. Lee et al.8 studied the performance of a dual-injection hydrogen-fueled engine by using solenoid in-cylinder injection and an external fuel injection technique. An increase in the thermal efficiency by about 22% was noted for dual injection at low loads and 5% at high loads compared to direct injection. Lee et al.9 suggested that in dual injection, the stability and maximum power could be obtained in direct injection of hydrogen and maximum efficiency in the manifold fuel preparation. The salient properties of hydrogen compared with diesel are shown in Table 1. Hydrogen Injection in the Intake System Hydrogen can be injected in the intake manifold by using a mechanically or an electronically operated injector or through carburetion. The advantage of hydrogen injection over carbureted system is that with proper injection timing, the backfire and preignition problems can be eliminated. Electronic injectors are robust in design with a greater control over the injection timing and injection duration with quicker response to operate under high-speed conditions.10-13 The injector works on the (6) Yi, H. S.; Lee, S. J.; Kim, E. S. Performance evaluation and emission characteristics of Incylinder injection type hydrogen fueled engine. Int. J. Hydrogen Energy 1995, 21 (7), 617–624. (7) Masood, M.; Ishrat, M. M.; Reddy, A. S. Computational combustion and emission analysis of hydrogen-diesel blends with experimental verification. Int. J. Hydrogen Energy 2006, 32 (13), 2539–2547. (8) Lee, S. J.; Yi, H. S.; Kim, E. S. Combustion characteristics of Intake Port Injection type Hydrogen Fueled Engine. Int. J. Hydrogen Energy 1995, 20 (4), 317–322. (9) Jong, T. L.; Kim, Y. Y.; Jerald, A. C. The development of a dual injection hydrogen fueled engine with high power and high efficiency 2002 Fall Technical Conference of ASME-ICED; Sept 8-11, 2002; pp 2-12. (10) Heffel, J. W.; McClanahan, M. N.; Norbeck, J. M. Electronic fuel injection for hydrogen fueled internal combustion engines. SAE Trans. 981924 1998, 421–432. (11) Lee, J. T.; Kim, Y. Y.; Lee, C. W.; Caton, J. A. An Investigation of a cause of Backfire and its control due to crevice volumes in a hydrogen fueled Engine. ASME 2001, 204 (123), 204–210. (12) Masood, M.; Mehdi, S. N.; Omar, M.; Reddy, P. R. Effect of injection delay on performance and emissions in hydrogen diesel dual fuel engine at different compression ratios. Inst. Eng. 2006, 87, 24–25. (13) Swain, M. R.; Schade, G. J.; Swain, M. N. Design and testing of a dedicated hydrogen-fueled engine. SAE Trans. 961077 1996, SP-1181, 1–11.

Figure 1. Cross-sectional view of the hydrogen injector. Table 2. Hydrogen Fuel Injector Specifications make supply voltage peak current holding current flow capacity working pressure durability range of injection length diameter (max) resistance inductance

Quantum Technologies 8-16 V 4A 1A 0.8 g/s@483-552 kPa 103-552 kPa >500 million cycles 1-12 ms 79.8 mm 24.5 mm (exclusive of connector) 2.05 ( 0.25 Ω at 20 °C 3.98 ( 0.3 mH at 1000 Hz typical

principle of a solenoid, which receives a constant power supply from a 12 V battery. The cross-sectional view of the hydrogen injector is shown in Figure 1. The technical specifications of the hydrogen injector are given in Table 2. The hydrogen injector was placed on the cylinder head at a distance of 10 mm from the valve seat position and placed perpendicular to the axis of the intake valve in the case of port fuel injection.14 For manifold injection the injector was placed on the intake manifold at a distance of 130 mm from the valve seat position and positioned at an angle of 45° to the axis of the intake valve in such a way that hydrogen can mix easily with air during suction.15,16 Knocking in a Hydrogen-Operated CI Engine It has been reported by many investigators that hydrogenfueled diesel combustion could be achieved to a limited extent because of the autoignition characteristics of the fuel. Use of pilot injection ensures ignition and also reduces the ignition delay to some extent.17 A small leakage from the injector most often exhibited a similar effect; sometimes a pronounced improved effect has also been observed on ignition. It has been found that by permitting a small leakage in the combustion chamber the engine can run without any symptoms of knock(14) Naber, J. D.; Siebers, D. L. Hydrogen combustion under diesel engine conditions. Int. J. Hydrogen Energy 1998, 23 (5), 363–371. (15) Bing, L.; Zuohua, H.; Ke, Z.; Hao, C.; Xibin, W.; Haiyan, M.; Deming, J. Experimental Study on Emissions of a Spark-Ignition Engine Fueled with Natural Gas-Hydrogen Blends. Energy Fuels 2008, 22 (1), 273– 277. (16) Jorach, R.; Enderle, C.; Decker, R. Development of a low NOX Truck hydrogen engine with high specific power output. Int. J. Hydrogen Energy 1997, 22 (4), 423–427. (17) Natkin, R. J.; Tang, X.; Boyer, B.; Oltmans, B.; Denlinger, A.; Heffel, J. W. Hydrogen IC Engine Boosting Performance and NOx study. SAE Trans. 2003-01-0631 2003, 865–875.

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SaraVanan and Nagarajan

Table 3. Engine Specifications make and model general type no. of cylinders bore stroke swept volume clearance volume compression ratio rated output rated speed combustion chamber type of cooling

Kirloskar, AV1 make 4-stroke/vertical compression ignition 1 80 mm 110 mm 553 cc 36.87 cc 16.5:1 3.7 kW at 1500 rpm 1500 rpm hemispherical open water cooled

ing.18 However, with excess fuel as the preliminary fuel it also results in autoignition by itself, which gives rise to rough combustion. Test Engine Details The engine used for the experimental work was a singlecylinder, water-cooled, four-stroke, vertical, naturally aspirated

stationary, DI diesel engine developing a rated power of 3.7 kW at a constant speed of 1500 rpm having a compression ratio of 16.5:1. The engine was modified to operate with hydrogen by positioning the injector on the cylinder head for port injection and intake manifold for manifold injection. Table 3 shows the technical specifications of the engine. The photographic view of the hydrogen flow arrangement is shown in Figure 2. Figure 3 shows the photographic view of the hydrogen injector position on the intake manifold. Table 4 shows the list of instruments used for the experiments. Safety Arrangements Hydrogen fuel was often associated with either the Hindenburg or the Challenger disasters or even the hydrogen bomb. However, other fuels such as gasoline and natural gas pose similar dangers. Hydrogen actually has a good overall safety record due to strict adherence to regulations and procedures and good training for the persons who handle hydrogen. For overall safety during hydrogen operation some safety devices and safety measures have been adopted for hydrogen operation.

Figure 2. Photographic view of the hydrogen flow arrangement.

Figure 3. Photographic view of the hydrogen injector position on the intake manifold.


Energy & Fuels, Vol. 23, 2009 2649 Table 4. Instrumentation List

instrument

purpose

make/model

1

electrical dynamometer

measurement of power output

2 3 4

exhaust gas analyzer smoke meter pressure transducer and charge amplifier digital mass flow controller hydrogen leak detector

measurement of HC, CO, CO2, and NOX measurement of smoke measurement of cylinder pressure

Laurence Scott and elctromotor Ltd., Norwich and Manchester, U.K.; capacity, 10kW; current rating, 43 A QRO 401, Qrotech Corp. Ltd., Korea TI diesel tune, 114 smoke density tester TI Tran service Type 5015A, Kistler Instruments, Switzerland

measurement of H2 flow detection of H2 leakage

DFC 46 mass flow controller AALBORG, USA Finch Mono II, Portable single gas monitor, INIFITRON Inc., Korea

5 6

A special hydrogen sensor was used to monitor the hydrogen gas in the operating environment and also to sense any leak of hydrogen in the pipeline during installation and operation of the engine. The sensor works on the principle of electrochemical reaction. Hydrogen has the highest diffusivity characteristics of about 3-8 times in air. Any hydrogen leakage will result in quicker dispersion in air compared to that of hydrocarbon dispersion. Hence, it will not form any cloud of hydrogen vapor in the working space. A blower was also made available to disperse the hydrogen gas in the environment in case of a leak, and proper ventilation was provided during the engine running condition. The hydrogen cylinders were also stored away from the working environment. The crankcase for the hydrogenoperated engine was properly ventilated to avoid ignition from taking place inside the crankcase due to blow by. The clouds of gases collected in the crankcase were removed from the rocker arm holes and vented into the atmosphere. The hydrogen present in the rocker arm assembly was found to be around 40-80 ppm during hydrogen operation. A flame arrestor was used for suppressing explosion inside a hydrogen-containing system. The flame arrestor consisted of a tank partly filled with

Figure 4. Peak and hold current for the hydrogen injector.

Figure 5. Photographic view of the hydrogen injector position on the cylinder head.

water with a fine wire mesh to prevent the flame penetration beyond the wire mesh. The flame will get quenched while reaching the water surface in case of any backfire. A nonreturn valve in the line was used to prevent the reverse flow of hydrogen into the system. Such a possibility of reverse flow can occur sometimes in a hydrogen-injected engine, particularly during the later part of injection duration. A flow indicator was used to see the flow of hydrogen during the engine running conditions. As the hydrogen was allowed to pass through a glass tube containing water, bubbles were formed during hydrogen flow, which clearly showed the flow of hydrogen. Experimental Setup Hydrogen was stored in a high-pressure storage tank at a pressure of 150 bar having a capacity of 7 m3 (0.5 kg) of hydrogen. A double-stage pressure regulator was used to reduce the pressure in the range of 1-4 bar based on the flow requirements. The hydrogen from the pressure regulator was passed through a shut-off valve, which can be closed if any back-fire results in the fuel pipeline. The hydrogen after passing through the shut-off valve was allowed to pass through the digital mass flow controller (DMFC). The DMFC precisely measures the flow rate of hydrogen in standard liters per minute. Since the hydrogen flow to the injector should be free from any impurities, the hydrogen was passed through a filtering device. The hydrogen from the filter was passed to a flame arrestor. The hydrogen from the flame arrestor was then passed to the flame trap, which consists of wire mesh. The wire mesh will prevent the penetration of flame to the hydrogen cylinder. The flame arrestor also acts as a nonreturn valve (NRV). The hydrogen from the flame trap was passed to the 2-way valve. One end of the two-way valve was connected to the pipeline, and it was kept away from the working area. This was done to remove the excess hydrogen in the fuel line during the engine shutoff time. The other end of the two-way valve was connected to a selector switch, which supplies the hydrogen to either the port fuel injector or the manifold injector. The port injector was placed on the engine head 13 mm above the intake valve, and the manifold injector was placed at a distance of 100 mm away from the engine head on the intake manifold. The injector used was a Quantum gas injector. The electronic control unit (ECU) controls the opening and closing of the fuel injector. One end of the positive power supply from the 12 V battery was connected to the injector; the other negative terminal of the injector was connected to the ECU, which controls the injector opening timing and duration.19 An infrared (IR) detector was used to give the signal to the ECU for injector opening. On the basis of the preset timing and duration the injector will be opened for injection and closed after injection. The injection timing and injection duration can be varied within the specified range by using a knob. The current requirement for opening the injector opening was 4 A, and for holding the armature to inject the fuel it was 1 A. Figure 4 shows the peak and hold current of the hydrogen injector. Figure 5 shows the hydrogen injector position on the cylinder head. Experimental work has been carried out on the port and manifold injection by varying the injection timing from 5° BGTDC to 30° after gas exchange top dead center (AGTDC) in steps of 5° CA and varied the injection duration as 30°, 60°, and 90° CA. On the basis of this combination, 21 operating parameters were obtained

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Table 5. Hydrogen Flow Rates Used in Port and Manifold Injection port injection 1 2 3 4 5

manifold injection

optimized timing

hydrogen flow rate (lpm)

optimized timing

hydrogen flow rate (lpm)

5° BGTDC, 30° CA 5° BGTDC, 30° CA 5° BGTDC, 30° CA 5° BGTDC, 30° CA 5° BGTDC, 30° CA

2 3.5 5.5 7.5 9.5

GTDC, 30° CA GTDC, 30° CA GTDC, 30° CA GTDC, 30° CA GTDC, 30° CA

2 3.5 5.5 7.5 9.5

Table 6. Energy Share Ratio of Hydrogen and Diesel in Port Injection load, %

hydrogen energy, %

diesel energy, %

no load 5% load 50% load 75% load full load

35.85 22.59 16.95 12.71 10.00

64.15 77.41 83.05 87.29 90.00

Table 7. Energy Share Ratio of Hydrogen and Diesel in Manifold injection load, %

hydrogen energy, %

diesel energy, %

no load 25% load 50% load 75% load full load

36.67 22.21 16.23 12.35 9.44

63.33 77.79 83.77 87.65 90.56

and the optimized injection timing and injection duration were determined. The optimized injection timing and injection duration for hydrogen port-injected dual-fuel operation were 5° BGTDC and 30° CA, respectively. For manifold-injected dual-fuel operation the optimized conditions were GTDC with 30° CA injection duration. With these optimized start of injection and injection durations for both port and manifold injection, the hydrogen flow rate was varied from 2 to 9.5 lpm for optimization. At higher flow rates above 9.5 lpm an improvement in performance and reduction in emissions were observed with the onset of knock. Hence, the optimum hydrogen flow was found to be 7.5 lpm in both port and manifold injection. Table 5 presents the different hydrogen flow rates used to find the optimum hydrogen flow rate in port and manifold injection. Tables 6 and 7 show the energy share ratio of hydrogen and diesel at different load conditions at optimized injection timing with an optimum hydrogen flow rate in port injection and manifold injection, respectively. The following observations were made during the engine operation: (1) Unstability in engine operation in manifold injection at high hydrogen flow rates, (2) onset of knock

during high hydrogen flow rates above 9.5 lpm, and (3) significant increase in smoke emissions at high flow rates of hydrogen at full load.

Results and Discussion Performance Characteristics. Figures 6 and 7 depict the variation of brake thermal efficiency for port injection and manifold injection, respectively, at different injection timings and injection durations for hydrogen operation. Brake thermal efficiency is defined as the amount of energy consumed to produce a unit kilowatt power. On the basis of the performance it was observed that the optimum injection timing and injection duration for hydrogen port injection was at 5° BGTDC with an injection duration of 30° CA. GTDC with an injection duration of 30° CA were optimum for manifold injection since the efficiency at these points is higher for the entire load spectrum, and also the specific emissions are also lower compared to other timings. Combustion is found to be smoother for the optimum injection timings in both port and manifold injection, which can be observed from the pressure crank angle diagram. The general reason for the improvement in performance is due to hydrogen, which enhances the diesel combustion.15 Table 8 gives the injection timings and duration for hydrogen operation in port and manifold injection. Figure 8 depicts the variation of brake thermal efficiency with load for optimized conditions. An increase in efficiency is observed in both port and manifold injection. The brake thermal efficiency is about 26% for hydrogen in both port and manifold injection compared to diesel of 21.6% at 75% load. Similarly, at full load for hydrogen the brake thermal efficiency is 24.5% in both port and manifold injection compared to 23.3% for diesel. Port injection shows a higher brake thermal efficiency of 2% and 3% at 75% load and full load, respectively, compared to manifold injection. The increase in brake thermal efficiency may be due to more charge admitted into the combustion

Figure 6. Variation of brake thermal efficiency with brake power for different hydrogen injection timings (Table 8) for port injection.



Figure 7. Variation of brake thermal efficiency with brake power for different hydrogen injection timings (Table 8) for manifold injection. Table 8. Start of Injection Timings and Injection Duration for Hydrogen Operation in Port and Manifold injection

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22

start of injection, crank angle (deg)

injection duration, crank angle (deg)

5° BGTDC 5° BGTDC 5° BGTDC GTDC GTDC GTDC 5° AGTDC 5° AGTDC 5° AGTDC 10° AGTDC 10° AGTDC 10° AGTDC 15° AGTDC 15° AGTDC 15° AGTDC 20° AGTDC 20° AGTDC 20° AGTDC 25° AGTDC 25° AGTDC 25° AGTDC 23° BITDC

30 60 90 30 60 90 30 60 90 30 60 90 30 60 90 30 60 90 30 60 90 diesel

chamber with lesser replacement of air, resulting in enhanced combustion of hydrogen in the case of port injection. From the heat release diagram, it can be observed that the peak heat release rate is 99.48 J/°CA in the case of port injection and 89.1 J/°CA in manifold injection. Even though the ignition delay is the same in both cases, the peak pressure is higher, which is about 83.3 bar in the case of port injection compared to 80.8 bar in manifold injection. Hence, both the peak heat release rate and peak pressure results in higher thermal efficiency in the case of port injection compared to manifold injection. The variation of specific energy consumption at all loads for optimized start of injection, injection duration, and optimum hydrogen flow rate in port and manifold injected hydrogen is shown in Figure 9. The specific energy consumption is 13.7 MJ/kWh in port injection and 14.1 MJ/kWh in manifold injection compared to diesel of 16.7 MJ/kWh at 75% load. At full load the lowest SEC is found to be 14.5 MJ/kWh in port injection and 15 MJ/kWh in manifold injection compared to diesel of 15.4 MJ/kWh. For 7.5 lpm hydrogen flow rate the overall SEC is better than other hydrogen flow rates. Hydrogen

Figure 8. Variation of brake thermal efficiency with load.

combustion exhibits a high cooling loss from the burning gas to the combustion chamber wall compared to hydrocarbon combustion because of its higher burning velocity and shorter quenching distance. These two characteristics have a strong influence on the thermal efficiency of hydrogen-operated engines. In order to improve the thermal efficiency of hydrogenoperated dual-fuel engines the coolant flow is reduced to 75% of the normal flow and set as 300 lpm for both hydrogen dualfuel operations and diesel to reduce the cooling loss ratio, which causes an increase in thermal efficiency for hydrogen diesel dualfuel operations. The SEC in port injection is lower by about 3-4% compared to that of manifold injection for the same hydrogen flow rate of 7.5 lpm. The reduction in SEC is due to the stability in the combustion process. Peak Pressure. Figure 10 shows the comparison of peak pressure with load for optimum hydrogen flow rate in both port and manifold injection systems. In general, the peak pressure

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Figure 11. Variation of pressure crank-angle diagram at full load.

be 99.65 J/°CA compared to manifold injection of 82.1 J/°CA. The ignition delay for the diesel combustion is calculated by the equation

[ ( RT1 - 17 1190 ) ( p -21.212.4 ) ](2)

τid(CA) ) (0.36 + 0.22SP)exp EA

0.63

Figure 9. Variation of specific energy consumption with load.

Figure 10. Peak pressure with load.

increases with an increase in hydrogen flow rate. This may be due to the rapid and instantaneous combustion that takes place at high hydrogen flow rates. The peak pressure of 80.2 bar is observed with hydrogen port injection and 76.9 bar in manifold injection compared to a diesel peak pressure of 78.5 bar. At full load in port injection the peak pressure is 83.3 and 80.8 bar in manifold injection compared to 82.2 bar in diesel. Compared to manifold injection, the peak pressure in port injection increases by 4% at 75% load and 3% at full load. The increase in peak pressure for port injection is due to the increase in heat release rate. The peak heat release rate is observed to

where τid is the ignition delay in crank angles, T is the temperature in Kelvin, p is the pressure in bar, SP is the mean piston speed in m/s, and R is the universal gas constant (8.3143 J/mol K). From the above equation it is observed that pressure and temperature are the main factors for ignition delay. As the pressure and temperature increase, the ignition delay gets reduced. In hydrogen diesel dual-fuel operation the peak pressure and temperature inside the combustion chamber are higher, which results in the reduction of delay period. The ignition delay is found to be 9° or 1 ms for diesel in hydrogen port fuel injection, while it is 10° or 1.11 ms for diesel in hydrogen manifold fuel injection compared to 11° or 1.22 ms in diesel. Pressure Crank Angle Diagram. Figure 11 portrays the pressure crank angle diagram at full load for a hydrogenoperated engine in port- and manifold-injected dual-fuel engine and diesel. The average pressure was obtained from 100 cycles. The peak pressure obtained for hydrogen operation at full load is 83.4 bar in port injection and 81.8 bar in manifold injection, whereas in the case of diesel it is 82.2 bar. The peak pressure in manifold injection is reduced by 2% compared to port injection due to the better mixing of air and fuel in manifold injection compared to port injection. The peak pressure in hydrogen port injection is advanced by 5° and 4° CA in manifold injection compared to diesel peak pressure. The advance in peak pressure in a hydrogen-operated engine is due to faster combustion of hydrogen. Heat Release Rate. The details about combustion stages and events can be determined by analyzing the heat release rates determined from cylinder pressure measurements. Analysis of heat release helps to study the combustion behavior of the engine. The analysis for the heat release rate is based on the application of first law of thermodynamics for an open system. It is assumed that cylinder contents are a homogeneous mixture of air and combustion products and are at uniform temperature and pressure during the combustion process. The first law for such a system is written as dQhr ) dU + dW + dQht

(3)

where dQhr is the instantaneous heat release modeled as heat transfer to the working fluid, dU is the change in internal energy of the working fluid, dW is the work done by the working fluid,



Figure 12. Heat release rate at full load.

and dQht is the heat transmitted away from the working fluid (to the combustion chamber walls). The change in internal energy is written as dU ) CV /R(p dV + V dp)

Figure 13. Variation of rate of pressure increase with crank angle at full load.

(4)

where the work done by the working fluid dW ) p dV. The heat transfer rate to the wall is written as dQht /dt ) hA(Tg - Tw)

(5)

where T, P, and V are the temperature, pressure, and volume, respectively, R is the gas constant, CV is the specific heat at constant volume, h is the heat transfer coefficient, and Tw is the temperature of the wall (400 K) h ) 3.26B-0.2p0.8T-0.55w0.8 where w is the average gas velocity in m/s, B (bore) is 0.088 m, P (cylinder pressure) is 29.327 bar, and T (gas temperature) is 2100 K

[

w ) C1Sp + C2

]

VdTc (p - pm) pcVc

(6)

where SP is the mean piston speed in m/s, Vd is the displaced volume in m3, Tc, pc, and Vc are the temperature, pressure, and volume, respectively, during combustion, p is the cylinder pressure during combustion, pm is the pressure in motorized condition, and C1 is 2.28 and C2 is 3.24 × 10-3 during combustion. From the first law of thermodynamics

Figure 14. Variation of exhaust gas temperature with load.

dV 1 dp γ dt P +V + hAs(Tg - Tw) (7) γ - 1 dθ γ - 1 dθ dθ where θ is the crank angle in degrees, γ is the ratio of specific heats of the fuel and air, and As is the area in m2 through which heat transfer from the gas to the combustion chamber walls takes place. Most studies show that compression and expansion processes are well fitted with a polytropic relation pVn ) constant. The exponent n for compression and expansion processes is 1.3 ((0.05) for conventional fuels. The pressure change due to volume change is given by ∆pv. Hence, ∆pv ) pj - pi ) pi[(Vi)/(Vj)n - 1]. Figure 12 shows the rate of heat release at full load. The combustion starts at the same crank angle of about 10° BITDC both in hydrogen as well as in diesel operation. It can be observed that as in the case of diesel operation, there are three phases of combustion, namely, premixed combustion phase, mixing controlled combustion phase, and late combustion phase in hydrogen operation. The heat release for hydrogen shows distinct characteristics of explosive, premixed-type combustion, in contrast to typical diffusion-type combustion of diesel fuel. The peak heat release rate for a hydrogen-operated engine is

99.6 J/°CA in port injection and 82.1 J/°CA in manifold injection compared to diesel of 81.5 J/°CA. The maximum heat addition occurs nearer to ITDC in hydrogen operation, which also causes the cycle efficiency to increase. Rate of Pressure Rise. Figure 13 depicts the variation of rate of pressure increase with crank angle for optimized conditions at full load. Diesel injection starts at 23° BITDC. The rate of pressure rise is maximum in port injection, 8.02 bar/°CA at 355° CA followed by 6.61 bar/°CA at 356° CA in manifold injection, and 6.63 bar/°CA in diesel at 356° CA. The increase in the rate of pressure increase in port injection may be due to the instantaneous combustion of hydrogen. The maximum rate of pressure increase shows a reduction of 21% in manifold injection compared to port injection. This may be due to more uniform mixture formation and progressive combustion in manifold injection rather than instantaneous combustion in port injection. The combustion duration in port injection is found to be 60° CA, and in manifold injection it is 63° CA compared to diesel of 66° CA. Exhaust Gas Temperature. Figure 14 shows the variation of exhaust gas temperature with load. The exhaust gas temper-

dQhr )

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ature is found to be higher for hydrogen operation. At full load the exhaust gas temperature is 469 °C with hydrogen in port injection and 481 °C in manifold injection compared to diesel of 452 °C. The increase in exhaust gas temperature may be attributed to creation of a hotter environment in the combustion chamber by hydrogen fuel, which in turn increases the exhaust gas temperature. Compared to manifold injection, the exhaust gas temperature in port injection is higher by 7% at 75% load and 3% at full load. The increase in temperature may be due to a higher peak temperature prevailing in the combustion chamber due to the combustion of hydrogen. Oxides of Nitrogen. Nitric oxide (NO) and nitrogen dioxide (NO2) are usually grouped together as NOX emissions. NO2/ NO in a diesel engine is approximately in a ratio of 10%/30%.20 The Zeldovich mechanism of NO formation is given below O + N2 ) NO + N

(8)

N + O2 ) NO + O

(9)

N + OH ) NO + H

(10)

The NO formation rate is derived from the following equation as 6 × 1016 -69 090 d[NO] ) exp [O2]1/2 e [N2]e 1/2 dt T T

(

)

(11) Figure 15. Variation of oxides of nitrogen with load.

The time calculation for NO is τNO )

8 × 10-16T exp(58 300/T) p1/2

(12)

From the above equation it can be noted that temperature is the main factor for the formation of NO compared to the availability of oxygen concentration. In diesel engines the combustion is always leaner, resulting in more oxygen availability in entire load spectrum of operation. NO formation in a diesel combustion peaks at a critical time period (between the start of combustion and shortly after the occurrence of peak cylinder pressure). After the occurrence of peak pressure the burned gas temperature decreases as the cylinder gases expand, which causes the combustion temperature to decrease, resulting in the freezing of NO without further reduction. In hydrogen diesel combustion the rate of pressure increase is maximum in port injection, 7.1 bar/°CA at 355° CA followed by 7 bar/°CA at 356° CA in manifold injection, and 6.63 bar/°CA in diesel at 356° CA at full load (Figure 13). The ignition delay or delay period is found to be 11° CA or 1.22 ms in diesel, in port injection it is 9° CA or 1 ms, and in manifold injection it is 10° CA or 1.11 ms, which results in rapid combustion and ends instantaneously, causing an increase in NO. The combustion duration is shorter due to its high burning velocity (2.65-3.25 m/s for hydrogen and 0.3 m/s for diesel) compared to diesel. In general, the NO formed in hydrogen-diesel dual-fuel operation is higher for both port and manifold injection compared to diesel operation. Figure 15 presents the variation of oxides of nitrogen (NOX) with load for port and manifold injection systems. The NOX emission is found to be lower at no load for all hydrogen flow (18) Shudo, T.; Nakajima, Y.; Futakuchi, T. Thermal efficiency analysis in a hydrogen premixed Combustion engine. JSAE Trans. 2000, 21, 177– 182. (19) Weigang, W.; Lainfang, Z. The research on internal combustion engine with the mixed fuel of diesel and hydrogen. Int. Symp. Hydrogen Syst. 1985, 2, 83–94. (20) Huang, Z.; Wang, J.; Liu, B.; Zeng, K.; Yu, J.; Jiang, D. Combustion Characteristics of a Direct-Injection Engine Fueled with Natural GasHydrogen Blends under Various Injection Timings. Energy Fuels 2006, 20 (4), 1498-1504.

rates, but the trend is reversed at 75% load. At no load NOX emission is 25.64 g/kWh in diesel, whereas with hydrogen in port injection it is 20.38 g/kWh, and in manifold injection it is 17.99 g/kWh. The reduction in NOX at low load may be due to the hydrogen operation with very lean mixtures, which in turn may result in a reduction in peak combustion temperature.21 At 75% load, NOX is observed to be maximum in port-injected hydrogen engine, 20.28 g/kWh compared to diesel of 18.11 g/kWh, and manifold injection of 19.04 g/kWh. At full load the NOX is 16.13 g/kWh in diesel compared to port injection of 15.92 g/kWh and manifold injection of 15.57 g/kWh. NOX emission increases by 6% at 75% load and 3% at full load in port injection compared to manifold injection. It is observed that the peak heat release rate is about 99.4 J/°CA in the case of port injection, which is higher than that of manifold injection of 89.15 J/°CA. It is expected that due to higher peak heat release, the peak combustion temperature may be higher, which may result in increased NOX emissions. Smoke. Diesel particulates or smoke consists principally of combustion-generated carbonaceous material (soot) on which some organic compounds have been absorbed. Most particulate material results from incomplete combustion of fuel hydrocarbons, and some is contributed by the lubricating oil. The equation for the formation of smoke is n CmHn + yO2 f 2yCO + H2 + (m - 2y)Cs 2

(13)

If m > 2y, then the C/O ratio exceeds unity, which results in the formation of smoke. The fuel air ratio for smoke formation is given by φ)2

( OC )(1 + δ)

(14)

where δ ) n/4m, φ is 3 for (C/O) ) 1 with n/m ) 2. Since carbon is not present in hydrogen smoke formation is totally (21) Heywood, J. B. Internal combustion engine fundamentals; McGrawHill: New York, 1998; pp 1-923.



Figure 16. Variation of smoke with load.

Figure 17. Variation of carbon monoxide with load.

eliminated. In a hydrogen-diesel dual-fuel-operated engine, due to the presence of hydrogen, the net smoke emission decreases.22 Figure 16 depicts the variation of smoke with load for the optimized hydrogen flow rate in port and manifold injection systems. The smoke is found to be lesser with hydrogen at all load conditions, which is attributed to the absence of carbon in hydrogen. The lowest smoke of 0.8 BSN is obtained with hydrogen in port injection, while in manifold injection it is 1 BSN, and in diesel it is 2.2 BSN at 75% load. At full load the smoke is 2 BSN with hydrogen in both port and manifold injection compared to diesel of 3.6 BSN. Port injection exhibits a lower smoke by 27% both at 75% load and full load with hydrogen compared to manifold injection. The reduction in smoke may be due to more hydrogen admitted in port injection, resulting in replacement of diesel. In general, smoke emission decreases with an increase in hydrogen flow rates, but after the optimum flow rate of 7.5 lpm the smoke increases due to the deterioration in combustion process and instability in combustion during high flow rates of hydrogen. Carbon Monoxide. The CO formation in the hydrocarbon radical is given by the following reaction RH f R f RO2 f RCHO f RCO f CO

(15)

where R stands for the hydrocarbon radical. CO formed through the combustion process was then oxidized to CO2 at a slower rate. The CO oxidation reaction in a hydrocarbon-air flame is given by the following equation CO + OH ) CO2 + H

(16)

On the basis of the oxygen availability the CO formed during the combustion is oxidized further to form CO2.23 In general, for hydrogen-diesel dual-fuel operation at full load condition the mixture is slightly richer than normal diesel combustion due (22) Masood, M.; Ishrat, M M. Computer simulation of hydrogen-diesel dual fuel exhaust gas emissions with experimental verification. Int. J. Fuel 2008, 87 (7), 1372–78. (23) Saravanan, N.; Nagarajan, G.; Sanjay, G.; Dhanasekaran, C.; Kalaiselvan, K. M. Combustion analysis on a DI diesel engine with hydrogen in dual fuel mode. Int. J. Fuel 2008, 87, 3591–3599.

Figure 18. Variation of carbon dioxide with load.

to the air replacement by hydrogen in the intake manifold that causes a slight increase in CO concentration. Figure 17 portrays the variation of carbon monoxide with load for port and manifold injection systems. The carbon monoxide is found to be lesser in both port and manifold injection at all loads. The lowest CO of 0.21 g/kWh is observed in manifold injection and 0.43 g/kWh in port injection compared to diesel of 0.65 g/kWh at 25% load. This may be due to the high flame velocity of hydrogen and the characteristics of hydrogen to operate with leaner mixtures. At 75% load CO is found to be 0.32 g/kWh in diesel, and in port and manifold injection it is 0.31 g/kWh. At full load the CO in diesel is found

2656



both port and manifold injection while that of diesel is 0.78 g/kWh. The reduction in CO2 emission is due to the absence of carbon in hydrogen.24 At full load the lowest CO2 of 0.68 g/kWh is observed with hydrogen in both port and manifold injection compared to diesel CO2 of 0.76 g/kWh. A significant reduction in CO2 emissions is noticed at lower loads compared to high loads. This is due to the higher energy share of hydrogen, about 30-35% at low loads in both port and manifold injection compared to higher loads, which is only 9-10%. Hydrocarbon. Hydrocarbons are organic emissions as a consequence of incomplete combustion of hydrocarbon fuel. The hydrocarbon oxidation rate is given by the following equation d[HC] -18 735 p ) -6.7 × 1015 exp xHCxO2 dt T RT

(

Table 9. Average Uncertainties of Some Measured and Calculated Parameters parameters

uncertainty

speed brake thermal efficiency temperature mass flow rate of air mass flow rate of diesel mass flow rate of hydrogen oxides of nitrogen hydrocarbon smoke particulate matter

1.2% 5.3% 0.6% 1.3% 2.1% 1.2% 1.7% 2.2% 2.1% 3.1%

( )

(17)

where xHC and xO2 are the mole fraction of HC and O2, t is in seconds, and T is in Kelvin. The reason for the increase in HC emission is due to over leaning (mixture becomes too lean to autoignite) and the longer ignition delay period. A small increase in ignition delay by 2° CA causes an increase in HC emission by 60-70%.25 For hydrogen-diesel dual-fuel operation the delay period is lower since as soon as the combustion of hydrogen starts it undergoes rapid combustion and assists in diesel combustion, resulting in a reduction in delay period. The ignition delay or delay period is found to be 11° or 1.22 ms in diesel, in port injection it is 9° or 1 ms, and in manifold injection it is 10° or 1.11 ms. The variation of hydrocarbon emission with load for port and manifold injection systems is depicted in Figure 19. The hydrocarbon emission is found to increase in both port and manifold injection due to the replacement of air by hydrogen, resulting in the reduction of oxygen for diesel combustion. At 25% load the HC emission is 0.30 g/kWh in both hydrogen port and manifold injection compared to diesel of 0.27 g/kWh. At 75% load the highest HC is found to be 0.15 g/kWh in port injection and 0.12 g/kWh in hydrogen manifold injection and diesel. At full load the HC in manifold injection is found to be 0.14 g/kWh in all cases expect for a decrease of 4% in port injection compared to manifold injection. The reduction in HC in port injection may be due to the lesser replacement of air in port injection compared to manifold injection.

Figure 19. Variation of hydrocarbon with load.

1 2 3 4 5 6 7 8 9 10

)

to be 0.88 g/kWh compared to 0.75 g/kWh in manifold injection and 0.56 g/kWh in port injection. The CO emission decreases by 25% in port injection compared to manifold injection at full load. The increase in CO in manifold injection may be due to higher replacement of air compared to port injection. Carbon Dioxide. The variation of carbon dioxide with load in port injection and manifold injection is shown in Figure 18. The CO2 emission is found to be lower in port and manifold injection at all load conditions. At 25% load the CO2 emission is 0.84 g/kWh in port and manifold injection compared to diesel of 1.3 g/kWh. At 75% load the CO2 emission is 0.64 g/kWh in

Summary Experimental investigations carried out on a DI diesel engine indicates that it is possible to operate with hydrogen in a diesel engine in the dual-fuel mode with some modifications depending on the techniques adopted. Some of the significant conclusions drawn from the present investigation are presented in this

Table 10. Mean and Standard Deviation for Exhaust Gas Temperature Measured for Six Trials trial-1 power

EG1

trial-2 2

trial-3 3

trial-4

trial-5

trial-6

average

SD

for diesel fuel and hydrogen port injection 4 5 6 EG (mean) diesel 194 260 318 394 463

deviation

value

(1.96SD)

average EG ( 1.96SD

no load 25% 50% 75% full load

197 254 322 378 452

190 251 312 381 449

192 249 323 385 461

192 247 315 393 465

190 252 329 383 456

192.500 252.167 319.833 385.667 457.667

2.665 4.140 6.113 5.935 6.377

5.223 8.114 11.981 11.632 12.499

197.723-187.277 260.281-244.052 331.814-307.852 397.299-374.034 470.166-445.168

no load 25% 50% 75% full load

195 258 308 392 489

199 259 314 400 476

191 253 305 397 481

190 249 297 380 480

hydrogen port injection 200 195 195.000 262 255 256.000 306 314 307.333 393 400 393.667 474 489 481.500

4.050 4.243 6.377 6.848 6.348

8.316 12.499 13.421 12.443

202.937-187.063 264.316-247.684 319.832-294.834 407.088-380.245 493.943-469.057



section. The optimized injection timing is 5° BGTDC and 30° CA injection duration in hydrogen port-injected dual-fuel operation. In the case of hydrogen manifold-injected dual-fuel operation the optimized injection timing is GTDC with 30° CA injection duration with hydrogen flow of 7.5 lpm. Brake thermal efficiency increases in port injection by about 6% and 2% in manifold injection compared to diesel at full load. A greater increase in efficiencies was noted at 75% load by about 21% in port injection and 18% in manifold injection compared to base diesel. The reason for this is due to the combustion of hydrogen, which enhances the diesel combustion. NOX emission is reduced by 2% in port injection and 4% in manifold injection compared to diesel at full load. At 75% load smoke emission in port injection is reduced by 2 times and manifold injection by 1 time. For the entire load spectrum a reduction in CO by about 50% is noticed in both port and manifold injection. The HC emission is found to increase by about 6% in port injection and 10% in manifold injection at full load. Use of hydrogen in diesel engine results in improvement in performance and reduction in emissions, which can be observed from the above results. In particular, port injection gives a significant improvement compared to manifold injection of hydrogen-diesel dual-fuel operation. Further improvement in efficiency and reduction in emissions in using hydrogen as a fuel in DI diesel engine can be obtained by direct injection of hydrogen by using spark plug or glow plug as an ignition source for hydrogen. For mitigating the NOX emissions the most commonly employed techniques, like exhaust gas recirculation and selective catalytic reduction, can be employed. Appendix Analysis of Confidence Level. All measurements of physical quantities are subject to confidence levels. Confidence level analysis is needed to prove the accuracy of the experiments. In order to have reasonable limits of confidence level for a computed value, an expression is derived as follows. Let ‘R’ be the computed result function of the independent measured variables x1, x2, x3, ..., xn, as per the relation R ) f(x1, x2,..., xn) and let error limits for the measured variables or parameters be x1 ( ∆n1, x2 ( ∆n2, ..., xa ( ∆xa and the error limits for the computed result be R ( ∆R. Hence, to get the realistic error limits for the computed result, the principle of root-mean-square method is used to get the mean square error given by Holman26 as ∆R )

[(

∂R ∆x ∂x1 1

) ( 2

+

)

∂R ∆x ∂x2 2

2

+ ... +

(

)]

∂R ∆x ∂xn n

2 1/2

(18)

Using eq 18 the confidence level in the computed values such as brake power, brake thermal efficiency, and fuel flow measurements were estimated. The measured values such as N, t, V, and I are estimated from their respective confidence level based on the Gaussian distribution. The confidence level in the measured parameters, voltage (∆V) and current (∆I), estimated by the Gaussian method, are (3 V and (0.14 A, respectively. For fuel time (∆tr) and fuel volume (∆t), the (24) Arcoumanis, C.; Bae, C.; Crookes, R.; Kinoshita, E. The potential of di-methyl ether (DME) as an alternative fuel for compression-ignition engines: a review. Int. J. Fuel 2008, 87 (7), 1014-30. (25) Swain, M. R.; Shriber, J.; Swain, M. N. Comparison of Hydrogen, Natural Gas, Liquified Petroleum Gas, and Gasoline Leakage in a Residential Garage. Energy Fuels 1998, 12 (1), 83–89. (26) Holman, J. P. Experimental methods for engineers; Mc Graw Hill: New York, 2001; pp 1-350.

confidence levels are taken as (0.2 and (0.1 s, respectively. For an N of 1500 rpm, V of 230 V, I of 14 A, fuel volume (fx) of 10 cc, and brake power (BP) of 3.7 kW the confidence level for brake power calculation is VI kW ηg × 1000 where ηg is the efficiency of the generator which is 86% and BP ) f(V, I). BP )

[

∂BP I 14 ) )) 0.0163 ∂V (0.86 × 1000) (0.86 × 1000) ∂BP V 230 ) ) ) 0.2674 ∂I (0.86 × 1000) (0.86 × 1000) 2 2 ∂BP ∂BP ∆BP ) ∆V + ∆I ∂V ∂I

[( ) ( ) ]

]

) [√(0.0163 × 3)2 + (0.2674 × 0.14)2] ) 0.185 kW Therefore, the confidence level in the brake power is (0.185 kW and the confidence level limits in the calculation of BP are 3.7 ( 0.185 kW. The confidence levels in temperature measurement are (1% (T > 150 °C), (2% (150 °C < T < 250 °C), and (3% (T < 250 °C). The confidence levels of other operating parameters are given in Table 9. Table 10 shows the mean and standard deviation calculations for six samples. Nomenclature m/s ) meters per second g/kWh ) gram per kilowatt hour ms ) milliseconds kW ) killowatt R ) gas constant cc ) cubic centimeter A ) amphere V ) volts rpm ) revolutions per minute MJ/kWh ) megajoules per kilowatt hour NO ) nitric oxide Greek Symbols γ ) specific heat ratio ηg) efficiency of generator Subscripts NO2 ) nitrogen dioxide C/O ) carbon to oxygen ratio HC ) hydrocarbon CO ) carbon monoxide CO2 ) carbon dioxide NOX ) oxides of nitrogen GTDC ) gas exchange top dead center BGTDC ) before gas exchange top dead center SEC ) specific energy consumption AGTDC ) after gas exchange top dead center BITDC ) before injection top dead center BP ) brake power TMI ) timed manifold injection ECU ) electronic control unit IR ) infrared DMFC ) digital mass flow controller CA ) crank angle BSN ) Bosch smoke number CA ) crank angle lpm ) liters per minute SI ) spark ignition CI ) compression ignition NRV ) nonreturn valve EF800962K

Experimental Investigation in Optimizing the Hydrogen Fuel on a

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