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An experimental investigation on performance, emissions, and combustion in a manifold injection for different exhaust gas recirculation flowrates in hydrogen −−diesel dual-fuel operations N Saravanan and G Nagarajan Proceedings of the Institution of Mechanical Engineers, Part D: Journal of Automobile Engineering 2008 222: 2131 DOI: 10.1243/09544070JAUTO921 The online version of this article can be found at: http://pid.sagepub.com/content/222/11/2131

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2131

An experimental investigation on performance, emissions, and combustion in a manifold injection for different exhaust gas recirculation flowrates in hydrogen–diesel dual-fuel operations N Saravanan1* and G Nagarajan2 1 ERC Engines, Tata Motors Ltd, Pimpri, India 2 Internal Combustion Engineering Division, Department of Mechanical Engineering, College of Engineering, Anna University, Chennai, India The manuscript was received on 21 May 2008 and was accepted after revision for publication on 17 July 2008. DOI: 10.1243/09544070JAUTO921

Abstract: Hydrogen is receiving considerable attention as an alternative fuel to replace the rapidly depleting petroleum-based fuels. Its clean burning characteristics help to meet the stringent emission norms. In this experimental investigation a single-cylinder diesel engine was converted to operate in hydrogen–diesel dual-fuel mode. Hydrogen was injected in the intake manifold and the diesel was injected directly inside the cylinder. The injection timing and the injection duration of hydrogen were optimized on the basis of performance and emissions. Best results were obtained with hydrogen injection at gas exchange top dead centre with an injection duration of 30u crank angle. The flowrate of hydrogen was optimized as 7.5 l/min with optimized injection timing and duration. The optimized exhaust gas recirculation (EGR) flowrate was 20 per cent at 75 per cent load. The optimized timings were chosen on the basis of performance, emission, and combustion characteristics. The EGR technique was adopted in the hydrogen–diesel dual-fuel mode by varying the EGR flowrate from 0 per cent to 25 per cent in steps of 5 per cent. The maximum quantity of exhaust gases recycled during the test was 25 per cent (up to 75 per cent load); beyond that unstable combustion was observed with an increase in smoke. The brake thermal efficiency with 20 per cent EGR decreases by 9 per cent compared with diesel. The nitrogen oxide (NOx) emission in hydrogen manifold injection decreases by threefold with 20 per cent EGR operation at full load. The NOx emission tends to reduce drastically with increase in the EGR percentage at all load conditions owing to the increase in heat capacity of the exhaust gases. The smoke decreases by 80 per cent in the dualfuel operation compared with diesel at 75 per cent load. Keywords: hydrogen, exhaust gas recirculation, manifold injection, emissions, performance, combustion

1

INTRODUCTION

The world at present is heavily dependent on petroleum fuels and the importance of alternative fuel research for internal combustion engines needs emphasis. Diesel engines are the main prime movers for public transportation vehicles, stationary power *Corresponding author: ERC Engines, Tata Motors, Hall 11A, New building, Pimpri, Pune, Maharshtra, 411018, India. email: [email protected]; [email protected] JAUTO921 F IMechE 2008

generation units, and agricultural applications. Therefore it is important to find a best alternative fuel, which emits fewer pollutants to the atmosphere from diesel engines. In this regard, hydrogen is receiving considerable attention as an alternative source of fuel to replace the rapidly depleting petroleum resources. Its clean burning characteristics provide a strong incentive to study its utilization as a possible alternate fuel. Hydrogen can be adopted in both a spark ignition (SI) engine and a compression ignition (CI) engine. Shudo and Yamada [1] concluded Proc. IMechE Vol. 222 Part D: J. Automobile Engineering

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N Saravanan and G Nagarajan

that, in an SI engine, hydrogen could be used as the sole fuel. In a CI engine, because of the high self-ignition temperature of hydrogen, some ignition source is needed to initiate the combustion. Hence the dual-fuel mode is adopted, in which diesel is used as an ignition source for hydrogen combustion. As knocking does not permit a large quantity of hydrogen to be used, hence its usage is restricted to 30 per cent [1]. Masood et al. [2] used hydrogen with diesel as a source of ignition. The highest brake thermal efficiency of 30 per cent was noticed. Lee et al. [3] studied the performance of a dual-injection hydrogen-fuelled engine by using solenoid in-cylinder injection and an external fuel injection technique. The increase in thermal efficiency for dual injection was about 22 per cent at low loads and 5 per cent at high loads compared with direct injection. Lee et al. [4] suggested that, by considering the dual injection, the stability and maximum power of the directinjection (DI) cylinder with maximum efficiency of the external mixture hydrogen engine could be obtained. Das [5] has carried out experiments on continuous carburation, continuous manifold injection, timed manifold injection, and low-pressure direct cylinder injection. The maximum brake thermal efficiency of 31.32 per cent was obtained in timed manifold injection. The major problem with hydrogen is the emission of nitrogen oxides (NOx), which is more predominant than other emissions, owing to the higher peak cylinder temperature in hydrogen combustion. Different methods that are widely used to reduce NOx from diesel engines are exhaust gas recirculation (EGR), retarded injection timing, fuel denitrogenation, staged injection of fuel, air preheat, charge air inter-cooling, water injection, exhaust catalysts, reduction of premixed burn fraction by reducing ignition delay, and optimization using a homogeneous charge compression ignition (HCCI) engine. Among these, EGR is one of the most effective techniques currently employed for reducing NOx emissions from internal combustion engines. There are two types of EGR one is internal EGR and the other is external EGR [6]. Internal EGR uses variable valve timings or other devices to retain a certain fraction of exhaust gases from the preceding cycle. As internal EGR provides a very short response time, its practical application is not possible until camless technology becomes widely available. Furthermore in the internal EGR technique the gases cannot be cooled; hence external EGR has emerged as the preferred current approach. External EGR involves Proc. IMechE Vol. 222 Part D: J. Automobile Engineering

diverting a fraction of the exhaust gases into the intake manifold where the recirculated exhaust gases mix with the incoming air before being inducted into the combustion chamber [7]. Recirculation of exhaust gases raises the total heat capacity of the working gases in the engine cylinder, thus lowering the peak gas temperature. There are three kinds of mechanism involved in EGR in the engine cylinder. 1. Dilution mechanism. The potential increased mixing time and longer burn duration caused by the dilution effect of EGR results in a lowered flame temperature. 2. Thermal mechanism. The increased heat capacity of the recirculated exhaust gases results in a lowered flame temperature. 3. Chemical mechanism. Increased dissociation from the more complex molecules (such as carbon dioxide (CO2) and water) results in a lowered flame temperature. In this, dilution plays a major role in NOx reduction in diesel-fuelled operation [8]. However, in the case of hydrogen-fuelled operation the thermal effect also plays a vital role, because of the excess amount of water vapour present in the exhaust, but the increase in water vapour level causes an increase in particulates and carbon emission owing to its chemical effect [9].

2

EMISSIONS

The use of hydrogen in combustion engines results in the emission of H2O vapour and does not produce any major pollutants such as hydrocarbons (HCs), carbon monoxide (CO), sulphur dioxide (SO2), lead, smoke, particulate matter, ozone, and other carcinogenic compounds. However, the formation of NOx is a major problem owing to the presence of nitrogen and oxygen in the air [10]. When the combustion temperature is high, some portion of nitrogen present in the air reacts with oxygen to form NOx. It can be reduced by operating the engine with lean mixtures that results in lowering the temperature inside the cylinder and slowing the chemical reaction, which weakens the kinetics of NOx formation. The most effective method of reducing NOx emissions is to adopt the EGR technique. By considering the gas constituent similar to fresh air diluted with some amount of EGR the total heat capacity can be calculated. Table 1 shows the total heat capacity of the exhaust gases by summarizing the cp of the JAUTO921 F IMechE 2008

Hydrogen–diesel dual-fuel operations

Table 1

Specific heat capacities (Cp) at constant pressure for the exhaust gas

Substance

Mole fraction

Weight (amu)

Mass (%)

Cp (kJ/kg K)

Nitrogen Oxygen Carbon dioxide Water vapour

0.78 0.18 0.02 0.02

28 32 44 18

75.7 20 3.1 1.2

1.04 0.92 1.85 1.86

substances weighed with their respective mass fractions [6]. The gas composition is 78 per cent N2, 18 per cent O2, 2 per cent CO2, and 2 per cent H2O (water vapour). The total average cp of the gas is found to be 1.05 kJ/kg K. As water has a latent heat of vaporization (V ) of 2260 kJ/kg; the vaporization of 2 per cent mole fraction (1.2 per cent mass fraction) corresponds to a reduction in average temperature by 26.8 K. DT ~

VXm, H2 OðlÞ 2260 ðkJ=kgÞ|1:2% ~26:8 K ð1Þ ~ 1:05 ðkJ=kg KÞ Cp, tot

In the present work, hydrogen was injected in the intake manifold with diesel injected in the conventional manner. The advantage of hydrogen injection over a carburetted system is that, with proper injection timing, the backfire and pre-ignition problems can be eliminated. With hydrogen operation, NOx emissions are higher; hence the EGR technique was adopted.

3

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rogen injector is shown in Fig. 1 and the specifications are shown in Table 3. The injector was controlled by using an electronic control unit (ECU). An infrared sensor was used to sense the crank angle position. The start of injection and the duration of the injector opening were controlled by using the ECU. Exhaust gases from the engine were regulated and cooled by using a counter-flow-type heat exchanger. The flowrate of cooling water in EGR systems was varied in such a way that the cooled exhaust gas temperature was maintained around 30 uC. The cooled exhaust gas was allowed to pass through a filtering device to remove the soot and particulate matter from the exhaust gas. The schematic view of the EGR system is shown in Fig. 2. The EGR percentage was calculated by taking the ratio of CO2 concentration present in the intake manifold to the CO2 concentration present in the exhaust gas. The EGR rate was calculated on volume basis using the formula [11] EGR rate~

EXPERIMENTAL SYSTEM

The engine used for the experimental work was a single-cylinder, water-cooled, four-stroke, vertical, naturally aspirated, stationary DI diesel engine developing a rated power of 3.7 kW at 1500 r/min having a compression ratio of 16.5:1. The technical specifications of the test engine are presented in Table 2. The cylinder head of the engine was modified to fit the solenoid-operated hydrogen gas injector. The injector was placed in the intake manifold at a distance of 15 cm from the intake valve from the valve seating position. The cross-section of the hydTable 2

4

½CO2 Intake gas {½CO2 Ambient ½CO2 Exhaust gas {½CO2 Ambient

ð2Þ

HEAT RELEASE ANALYSIS

The details about combustion stages and events can be determined by analysing the heat release rates as determined from cylinder pressure measurements.

Engine specifications

Type

Compression ignition

Number of cylinders Bore Stroke Speed Rated power Compression ratio Type of cooling Injection timing

1 80 mm 110 mm 1500 r/min 3.7 kW at 1500 r/min 16.5: 1 Water cooled 23u BITDC*

*BITDC, before ignition top dead centre.

JAUTO921 F IMechE 2008

Fig. 1

Hydrogen gas injector

Proc. IMechE Vol. 222 Part D: J. Automobile Engineering

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Table 3

Hydrogen injector specifications

Make

Quantum Technologies

Supply voltage Peak current Holding current Flow capacity Working pressure

8–16 V 4A 1A 0.8 g/s at 483–552 kPa 103–552 kPa

The change in internal energy can be written as dU~

ð4Þ

where

Analysis of heat release can help to study the combustion behaviour 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 the cylinder contents are a homogeneous mixture of air and combustion products and are at a uniform temperature and pressure during the combustion process. The first law for such a system can be written as dQhr ~dUzdW zdQht

CV ðP dV zV dP Þ R

R 5 gas constant P 5 pressure V 5 volume CV 5 specific heat at constant volume The work done by the working fluid is dW 5 P dV. The heat transfer rate to the wall can be written as dQht ~hA Tg {Tw dt where T 5 temperature of the gas A 5 area h 5 heat transfer coefficient Tw 5 temperature of the wall 5 400 K

ð3Þ

where dQhr 5 instantaneous heat release modelled as heat transfer to the working fluid dU 5 change in the internal energy of the working fluid dW 5 work done by the working fluid dQht 5 heat transmitted away from the working fluid (to the combustion chamber walls)

ð5Þ

With suitable assumptions, the first law of thermodynamics can be written as dQht ~

dt c dV 1 dp P zV zhAs Tg {Tw c{1 dh c{1 dh dh ð6Þ

where h 5 crank angle (deg) c 5 the ratio of the specific heat of the fuel to that of the air As 5 area (m2) through which heat transfer from the gas to combustion chamber walls takes place. The pressure value is obtained from the cylinder pressure data at the corresponding crank angle. This relation makes it possible to calculate the heat release rate. All the quantities on the right-hand side are known or can be easily derived once the pressure– time history is recorded.

5

Fig. 2

Schematic view of the EGR unit


EXPERIMENTAL METHODOLOGY

Hydrogen was stored in a high-pressure storage tank at a pressure of 150 bar. A double-stage pressure regulator was used to regulate the hydrogen pressure. The pressure was reduced to the range 1–4 bar based on the flow requirements. Hydrogen from the pressure regulator was passed through a shut-off valve, which can be closed if any backfire results in JAUTO921 F IMechE 2008


the fuel line [12, 13]. The hydrogen after passing through the shut-off valve was allowed to pass through the filter and then to digital mass flow controller (DFC). The DFC precisely measures the flowrate of hydrogen in standard litres per minute [14]. The hydrogen from the DFC was passed to a flame trap and then to a flame arrestor. The flame trap acts as a non-return valve and it also acts as a visible indicator for hydrogen flow. The flame arrestor consists of a bursting diaphragm, which punctures when the pressure inside the system exceeds 10 bar during backfire conditions [15]. Figure 3 shows the schematic diagram of the experimental set-up. To determine the optimum start of injection and injection duration for hydrogen operation the hydrogen flowrate was set at 5.5 l/min. Three injection durations of 3.3 ms at 30u crank angle (CA), 6.6 ms at 60u CA, and 9.9 ms at 90u CA were selected, as the fuel injector can open only for a maximum duration of 10 ms. The injection timings tested were 5u before gas exchange top dead centre (BGTDC) to 25u after gas exchange top dead centre (AGTDC) in steps of 5u CA. As the intake valve opening was 4.5u BGTDC, it was taken as the initial timing and the injector opening timing was varied in steps of 5u CA until the optimum condition was reached. From the results, it was observed that the optimized injector opening timing was gas exchange top dead centre (GTDC) with an injection duration of 30u CA based on the performance and emissions. After optimizing the injection timing and injection duration the flowrate of the hydrogen was optimized by varying the hydrogen flowrate in the order

Fig. 3

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2 l/min, 3.5 l/min, 5.5 l/min, 7.5 l/min, and 9.5 l/min. Based on the performance and emission results the best flowrate obtained was 7.5 l/min. The exhaust gas was varied for the EGR operation by maintaining the optimum hydrogen injection at 5u BGTDC with an injection duration of 30u CA for the hydrogen flowrate of 7.5 l/min. The EGR rate was varied from 0 per cent to 25 per cent on a volume basis in steps of 5 per cent. The power output of the test engine was measured with an electrical dynamometer. The power capacity of the dynamometer was 10 kW with a current rating of 43.5 A. The exhaust emissions HCs, CO, CO2, and NOx were measured with a nondispersive infrared-type analyser and smoke with a Bosch-type smoke meter. The cylinder pressure was measured using a piezoelectric pressure transducer and a charge amplifier and the pressure data were given as input to the oscilloscope for further analysis. The mass flow of hydrogen was measured using a digital mass flow controller, which measures in standard litres per minute in addition to controlling the flow.

6

ERROR ANALYSIS AND UNCERTAINTY

All measurements of physical quantities are subject to uncertainties. Uncertainty analysis was needed to prove the accuracy of the experiments. Hence to obtain realistic error limits for the computed result, the principle of the r.m.s method was used to obtain the magnitude of error given by Holman [16] as

Schematic diagram of the experimental set-up (PC, personal computer; IR, infrared)



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Table 4

" 2 2 2 #1=2 LR LR LR DR~ Dx1 z Dx2 z . . .z Dxn Lx1 Lx2 Lxn ð7Þ where R is the computed result function of the independent measured variables x1, x2, x3, …, xn, as per the relation R~f ðx1 , x2 , . . . , xn Þ

ð8Þ

The error limits for the measured variables or parameters are x1, ¡Dn1, x2 ¡ Dn2, …, xa ¡ Dxa, and the error limits for the computed result are R ¡ DR. Using equation (8) the uncertainty in the computed values such as the brake power, brake thermal efficiency, and fuel flow measurements were estimated. The measured values such as the speed, fuel time, voltage, and current were estimated from their respective uncertainties based on the Gaussian distribution. The uncertainties in the measured parameters, i.e. the uncertainty DV in the voltage and the uncertainty DI in the current, estimated by the Gaussian method, are ¡10 V and ¡0.16 A respectively. The uncertainty Dtr in the fuel time and the uncertainty Dt in the fuel volume are taken as ¡0.2 s and ¡0.1 s respectively. A sample calculation is as follows with the parameters Speed N 5 1500 r/min Voltage V 5 230 V Current I 5 12 A Fuel volume fx 5 10 cm3 Brake power BP 5 4.4 kW

BP~f ðV , I Þ LBP I 16 ~ ~{ ~0:0188 LV 0:85|1000 0:85|1000 LBP V 230 ~ ~ ~0:2705 LI 0:85|1000 0:85|1000 sffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi 2 2ffi LBP LBP DV z DI DBP~ LV LI qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi ~ ð0:0188|10Þ2 zð0:2705|0:16Þ2


Uncertainty (%)

Speed Temperature Mass flowrate of air Mass flowrate of diesel Mass flowrate of hydrogen NOx HC Smoke Particulate matter

1.4 0.6 1.3 1.4 0.8 2.7 2.2 2.2 2.9

Therefore, the uncertainty in brake power is ¡0.1929 kW and the uncertainty limits in the calculation of the brake power are 4.4 ¡ 0.1929 kW. The details of the estimated average uncertainties of some measured and calculated parameters at some typical operating conditions are given in Table 4. It can be observed that the uncertainty ranges from 0.6 per cent to 2.9 per cent.

7

RESULTS AND DISCUSSION

The results of the experimental investigation on using hydrogen in the dual-fuel mode in a DI diesel engine with the EGR technique compared with neat diesel operation are presented (a summary of the results is given in Appendix 2). The optimum EGR percentage is found to be 20 per cent. Table 5 shows the energy share ratio of hydrogen energy to total energy.

Performance characteristics

Figure 4 presents the variation in the brake thermal efficiency with load for manifold injection. The brake thermal efficiency is the ratio of the brake energy to the fuel energy. The brake thermal efficiency increases in hydrogen operation without EGR and increases with EGR operation up to part-load condition but drops at full load. At 25 per cent load the brake thermal efficiency is 11.9 per cent for diesel and 15.4 per cent in hydrogen operation without

VI kW gg |1000

~0:1929 kW

Parameter

7.1

The brake power is given by BP~

Average uncertainties of some measured and calculated parameters

Table 5 Energy share ratio of hydrogen and diesel for different load conditions at optimized injection timings in manifold injection

ð9Þ

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



Fig. 4

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Variation in the brake thermal efficiency with load for different EGR flowrates in manifold injection

EGR. With 20 per cent EGR the brake thermal efficiency is 14.5 per cent compared with 14.9 per cent with 25 per cent EGR operation at 25 per cent load. From the figure it can be observed that at part loads (up to 50 per cent of load), the brake thermal efficiency increases with increase in EGR percentage compared with diesel. This is because at no loads a lean mixture is present inside the engine cylinder and increasing the quantity of exhaust gases results in a reduction in the air-to-fuel ratio and increases the inlet charge temperature, which results in an increase in the thermal efficiency of the engine [17]. The brake thermal efficiency at 75 per cent load is 21.6 per cent in diesel and 25.6 per cent in hydrogen operation without EGR. With 20 per cent EGR at 75 per cent load the brake thermal efficiency is 25 per cent and, with 25 per cent EGR, it is 24.8 per cent. At full load the brake thermal efficiency is 23.4 per cent for diesel and 23.9 per cent in hydrogen operation without EGR. With 20 per cent EGR the brake thermal efficiency is 21.6 per cent. The reduction in brake thermal efficiency at full load is due to the higher specific heat ratio of exhaust gases and the presence of a low oxygen concentration reduces the combustion temperature [13]. The variation in specific energy consumption (SEC) with load for different EGR flowrates is shown in Fig. 5. The SEC decreases in hydrogen operation without EGR but increases with increase in EGR percentage up to 75 per cent load. At 25 per cent load the SEC is 30.27 MJ/kW h for diesel compared with 23.36 MJ/kW h in hydrogen operation without EGR. JAUTO921 F IMechE 2008

With 20 per cent EGR the SEC is 24.77 MJ/kW h compared with 24.17 MJ/kW h with 25 per cent for EGR operation. The reduction in SEC at part load is due to the increase in the charge temperature of inlet gases. At full load the SEC is 15.41 MJ/kW h for diesel compared with 15.05 MJ/kW h in hydrogen operation without EGR. With 20 per cent EGR operation the SEC is 16.63 MJ/kW h and at full load with 25 per cent EGR the engine was not able to reach the rated power and speed owing to the larger replacement of air by both hydrogen and exhaust gases. With an increase in the EGR percentage at part load the SEC drops because of the replacement of excess air by inert gases which improves the overall charge temperature at part load [14]. At full load, owing to the reduction in oxygen concentration, as a result of the replacement of air by hydrogen and in EGR operation, more fuel is admitted to attain the rated power, which results in an increase in the SEC.

7.2

Combustion characteristics

Figure 6 presents the variation in the peak pressure with load. A reduction in the peak pressure is observed with increase in the EGR concentration. At no load the peak pressure for diesel is 59 bar and in hydrogen operation without EGR it is 51.3 bar. With 20 per cent EGR, the peak pressure is 52.6 bar and, with 25 per cent EGR, it is 52 bar. At 75 per cent load for diesel the peak pressure is 78.5 bar and, with 20 per cent and 25 per cent EGR, it is 73.1 bar. At full load the peak pressure of 82.2 bar is observed for Proc. IMechE Vol. 222 Part D: J. Automobile Engineering

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Fig. 5

Variation in the specific energy consumption with load for different EGR flowrates in manifold injection

Fig. 6

Variation in the peak pressure with load for different EGR flowrates in manifold injection

diesel and 80.8 bar in hydrogen operation without EGR, while, with 20 per cent EGR it is 76.1 bar. At full load with 25 per cent EGR the engine is not able to reach the rated speed and power [18]. The reduction in peak pressure is due to the deterioration in the combustion. Figure 7 shows the pressure–CA diagram for hydrogen operation with an optimized EGR of 20 per cent and without EGR and for diesel at 75 per cent load. The peak pressure is 76.9 bar for the hydrogen-operated engine compared with 78.5 bar for diesel and 73.5 bar with 20 per cent EGR. The peak Proc. IMechE Vol. 222 Part D: J. Automobile Engineering

pressure decreases with 20 per cent EGR by 5 per cent compared with that of hydrogen without EGR and there is a shift in the peak pressure towards TDC by 2u CA compared with that of hydrogen injection. The shift is due to the increase in delay period as a result of the presence of exhaust gases in the intake manifold. Figure 8 depicts the variation in heat release rate at 75 per cent load. The peak heat release decreases from 58.6 J/deg CA in hydrogen manifold injected engine to 54 J/deg CA in diesel and 53.5 J/deg CA with 20 per cent EGR. This is due to the reduction in JAUTO921 F IMechE 2008


Fig. 7

Variations in the pressure with CA at 75 per cent load for the optimized EGR flowrate in manifold injection

Fig. 8

Variation in the heat release rate with CA at 75 per cent load with the optimized EGR flowrate in manifold injection

the oxygen concentration as a result of replacement by CO2. With increase in the EGR percentage there is an increase in the ignition delay, providing more time for the fuel to mix with oxygen, which would have increased the amount of premixed fuel. However, the reduction in oxygen concentration reduces the intensity of premixed combustion, thereby offsetting the effect of extra premixed fuel to a certain extent. This can be clearly observed in the peak heat release rate with EGR. With an increase in CO2 concentration the time taken for the inlet charge to ignite reduces substantially. Following the more JAUTO921 F IMechE 2008

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rapid initial burning with EGR the subsequent combustion phase progresses slowly with increasing CO2 replacing the oxygen in the inlet charge. This may also be a reason for the shift in the end of combustion by 4u CA with EGR operation compared with hydrogen operation. Figure 9 depicts the variation in the rate of pressure rise with CA at 75 per cent load. The start of injection in diesel is at 23u BITDC. The ignition delay is 11u CA or 1.22 ms in diesel and it is 12u CA or 1.33 ms in hydrogen injection. As the combustion starts, the rate of pressure rise increases in a Proc. IMechE Vol. 222 Part D: J. Automobile Engineering

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Fig. 9

Variation in the rate of pressure rise with CA at 75 per cent load with an optimized EGR flowrate of 20 per cent

progressive manner compared with diesel in hydrogen operation owing to the presence of exhaust gases. The combustion duration in hydrogen operation is 21u CA and in manifold injection with 20 per cent EGR it is 18u CA compared with diesel for which it is 26u CA. The rate of pressure rise is maximum in hydrogen operation, 4.43 bar/deg CA at 356u CA with 20 per cent EGR followed by 3.9 bar/deg CA for diesel at 357u CA. The increase in rate of pressure rise may be due to the instantaneous combustion that takes place because of hydrogen [13]. The rate of pressure rise is higher by about 10 per cent in the case of

Fig. 10

hydrogen operation without EGR compared with 20 per cent EGR owing to the hydrogen combustion and reduction in inert gases such as water vapour and CO2.

7.3

Emission characteristics

The variation in NOx with load is depicted in Fig. 10. At 25 per cent load the NOx level is 25.38 g/kW h in diesel and 18.01 g/kW h in hydrogen operation without EGR, while with 20 per cent EGR the NOx level is

Variation in NOx concentration with load for different EGR flowrates in manifold injection




8.8 g/kW h and with 25 per cent EGR it is 5.91 g/kW h. The NOx emission tends to reduce drastically with increase in EGR percentage for all load conditions owing to the increase in heat capacity of the exhaust gases. At 75 per cent load in diesel the NOx emission is 17.92 g/kW h while, in hydrogen operation without EGR, it is 19.04 g/kW h and, with 20 per cent EGR, it is 6.56 g/kW h and, with 25 per cent EGR, 4.53 g/kW h. At full load the NOx emission decreases from 16.01 g/kW h for diesel to 15.63 g/kW h in hydrogen operation without EGR. With 20 per cent EGR the NOx level is 2.44 g/kW h. The NOx emission reduces with increase in EGR percentage, which reduces both the peak combustion temperature as well as Table 6

Equations for Emission formation

Mechanisms

Equations

Kinetics of NO formation O + N2 5 NO + N Zeldovich mechanism N + O2 5 NO + O N + OH 5 NO +H 2 2R1 1{½ðNOÞ=ðNOÞe d½NO dt ~ 1zð½NO=½NOe ÞR1 =ðR2 zR3 Þ where R1 ~k1z ½Oe ½N2 e ~k1{ ½NOe ½Ne R2 ~k2z ½Ne ½O2 e ~k2{ ½NOe ½Oe R3 ~k3z ½Ne ½OHe ~k3{ ½NOe ½He Formation of NO2

NO + HO2 R NO2 + OH NO2 + O R NO + O2

CO formation

RH R R R RO2 R RCHO R RCO R CO CO + OH 5 CO2 + H

Arrhenius equation in HC formation

tCH4[O2]1.37[CH4]20.17 5 5.85610217 exp(45 940/RT)

Fig. 11

2141

the inlet oxygen concentration [19]. Table 6 shows the equations for emission formation. Figure 11 depicts the variation in smoke. The smoke emission reduces with increase in hydrogen percentage and increases with increase in EGR percentage. Even with the use of EGR the smoke concentration is less than diesel smoke up to 75 per cent load. At 75 per cent load, for diesel, a smoke concentration of 2.2 BSN is observed while, in hydrogen operation without EGR, it is 1.1 BSN. With 20 per cent EGR, the smoke level is 2 BSN and, with 25 per cent EGR, it is 1 BSN. Similarly, at full load, a smoke level of 3.6 BSN is observed for diesel and 2.1 BSN in hydrogen operation without EGR whereas, with 20 % EGR, it is 5.2 BSN. The increase in smoke concentration is due to partial replacement of air by exhaust gas, which may result in combustion instability. The variation in CO with load is presented in Fig. 12 for the hydrogen operation with and without EGR compared with diesel. At 25 per cent load the CO concentration for diesel is 0.65 g/kW h while, with 25 per cent EGR, it is 0.43 g/kW h. The reduction in CO may be due to EGR which raises the intake air temperature, leading to better oxidation. The CO emission at full load is higher compared with diesel. An increase in EGR percentage at no load does not have a significant effect on CO emissions. With increase in EGR percentage the CO concentration increases in diesel where it is 0.32 g/kW h while, with 20 per cent EGR it is 0.55 g/kW h at 75 per cent load. The increase in the CO concentration is due to the partial replacement of oxygen in the inlet charge by

Variation in the smoke with load for different EGR flowrates in manifold injection



2142


Fig. 12

Variation in CO with load for different EGR flowrates in manifold injection

the exhaust gases. This leads to incomplete combustion owing to the lack of oxygen in the combustion chamber. Also the lower oxygen concentration results in a richer mixture because of the increase in the CO concentration. The variation in the CO2 concentration is shown in Fig. 13. The CO2 concentration decreases in general for various EGR percentages. At 75 per cent load the CO2 concentration with 20 per cent EGR is 0.68 g/ kW h and, with 25 per cent EGR, 0.66 g/kW h. At full load the CO2 concentration is 0.75 g/kW h for diesel and 0.69 g/kW h in hydrogen operation without EGR. With 20 per cent EGR the CO2 concentration is 0.51 g/kW h. A reduction in the CO2 concentration at full load in hydrogen operation by 13 per cent with

Fig. 13

20 per cent EGR is observed. This is because circulation of higher EGR percentage reduces the peak combustion temperature and lack of oxygen present in combustion chamber leads to poor combustion which results in lower CO2 emissions during the combustion process [20]. The variation in HC with load is depicted in Fig. 14. With 20 per cent EGR the HC level is 0.12 g/kW h and, with 25 per cent EGR, 0.31 g/kW h at 25 per cent load. In diesel, it is 0.27 g/kW h and, in hydrogen operation without EGR, 0.3 g/kW h. At 75 per cent load the HC emission increases from 0.13 g/kW h for diesel to 0.13 g/kW h with 20 per cent EGR compared with 0.15 g/kW h with 25 per cent EGR. At full load the HC is found to be 0.21 g/kW h with 20

Variation in CO2 with load for different EGR flowrates in manifold injection




Fig. 14

Variation in hydrocarbon with load for different EGR flowrates in manifold injection

per cent EGR compared with 0.13 g/kW h for diesel. The increase in HC is due to the reduction in the oxygen concentration in the intake charge [21].

8

2143

CONCLUSIONS

Experiments were carried out on a single-cylinder DI diesel engine using hydrogen fuel injection in the intake manifold. The EGR technique was adopted for NOx reduction. The following conclusions are drawn. The optimum timing for hydrogen injection was at GTDC, with an injection duration of 30u CA and a hydrogen flowrate of 7.5 l/min. 20 per cent EGR was found to be the optimum at 75 per cent load. With 25 per cent EGR the engine was not able to reach the rated power and speed owing to larger replacement of air by both hydrogen and exhaust gases. Compared with base diesel, the brake thermal efficiency improved by 15 per cent in hydrogen operation without EGR and by 14 per cent with 20 per cent EGR at 75 per cent load. The NOx emission in the case of hydrogen operation is reduced threefold to fourfold while using EGR owing to the increase in the heat capacity of the exhaust gases. The smoke emission is reduced with an increase in the hydrogen percentage but it increases with an increase in the EGR percentage. The smoke increases by 35 per cent at full load with 20 per cent EGR compared with diesel. CO and unburnted HC emissions increase with increasing EGR; this is due to the reduction in the inlet oxygen concentration. The JAUTO921 F IMechE 2008

CO2 emission decreases using hydrogen compared with diesel operation. In general by using hydrogen as a partial fuel there is a significant reduction in smoke emissions and improvement in the performance of the engine. However, NOx emission increases, which can be reduced by the EGR technique. Hence hydrogen injection together with EGR can be adopted in engines to give a significant reduction in the pollutant concentration.

REFERENCES 1 Shudo, T. and Yamada, H. Hydrogen as an ignition-controlling agent for HCCI combustion engine by suppressing the low-temperature oxidation. Int. J. Hydrogen Energy, 2007, 32(14), 3066–3072. 2 Masood, M., Ishrat, M. M., and Reddy, A. S. Computational combustion and emission analysis of hydrogen–diesel blends with experimental verification. Int. J. Hydrogen Energy, 2007, 32(13), 2531–2547. 3 Lee, J. T., Kim, Y. Y., Lee, C. W., and Caton, J. A. An investigation of a cause of backfire and its control due to crevice volumes in a hydrogen fueled engine. Trans. ASME J. Engng Gas Turbines Power, 2001, 123, 204–210. 4 Lee, J. T., Kim, Y. Y., and Caton, J. A. The development of a dual injection hydrogen fueled engine with high power and high efficiency. In Proceedings of the 2002 Fall Technical Conference of the Internal Combustion Engine Division of ASME, New Orleans, Louisiana, USA, 8–11 September 2002, pp. 2–12 (American Society of Mechanical Engineers, New York). Proc. IMechE Vol. 222 Part D: J. Automobile Engineering

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5 Das, L. M. Near-term introduction of hydrogen engines for automotive and agricultural application. Int. J. Hydrogen Energy, 2002, 27, 479–487. 6 Egnell, R. The influence of EGR on heat release rate and NO formation in a D.I. diesel engine. SAE paper 2000-01-1807, 2000. 7 Abd-Alla, G. H. Using exhaust gas recirculation in internal combustion engines: a review. Energy Conservation Managmt, 2001, 43, 1027–1042. 8 Das, L. M. Fuel induction techniques for a hydrogen operated engine. In Hydrogen fuel for surface transportation, 1996, pp. 27–36 (SAE International, Warrendale, Pennsylvania). 9 Heffel, J. W. NOx emission and performance data for a hydrogen fueled internal combustion engine at 1500 rpm using exhaust gas recirculation. Int. J. Hydrogen Energy, 2002, 28, 901–908. 10 Ladommatos, N., Abdelhalim, S. M., Zhao, H., and Hu, Z. Effect of EGR on heat release in diesel combustion. SAE paper 980184, 1998. 11 Ladommatos, N., Abdelhalim, S. M., Zhao, H., and Hu, Z. The dilution, chemical and thermal effects of exhaust gas recirculation on diesel engine emission – part 4: effect of carbon dioxide and water vapour. SAE paper 971660, 1997. 12 Haragopala Rao, B., Shrivastava, K. N., and Bhakta, H. N. Hydrogen for dual fuel engine operation. Int. J. Hydrogen Energy, 1982, 8, 381–384. 13 Yi, H. S., Min, K., and Kim, E. S. Optimized mixture formation for hydrogen fuelled engine. Int. J. Hydrogen Energy, 2000, 25, 685–690. 14 Ikegami, M., Miwa, M., and Shioji, M. A study on hydrogen fuelled compression ignition engines. Int. J. Hydrogen Energy, 1982, 7(4), 341–353. 15 Selim, M. Y. E. A study of some combustion characterstics of dual fuel engine using EGR. SAE paper 2003-01-0766, 2003. 16 Holman, J. P. Experimental methods for engineers, 1973 (McGraw-Hill). 17 Ravi, M., Ramaswamy, M. C., Jagadeesan, T. R., and Rao, A. N. Experimental investigation on dual fuel operation of hydrogen in a C.I. engine. In Proceedings of the National Conference on I.C. engines and combustion, Dehradun, India, 15–18 September 1992, pp. 86–91 (Indian Institute of Petroleum, Dehradun). 18 Jorach, R., Enderle, C., and Decker, R. Development of a low NOx truck hydrogen engine with high


specific power output. Int. J. Hydrogen Energy, 1997, 22(4), 423–427. 19 Verhelst, S., Verstraetran, S., and Sierens, R. A comprehensive overview of hydrogen engine design features. Proc. IMechE, Part D: J. Automobile Engineering, 2007, 221(8), 911–920. 20 Saravanan, N. and Nagarajan, G. An experimental investigation on optimized manifold injection in a direct-injection diesel engine with various hydrogen flowrates. Proc. IMechE, Part D: J. Automobile Engineering, 2007, 221(12), 1575–1584. 21 Ladommatos, N., Xiao, Z., and Zhao, H. Effects of fuels with a low aromatic content on diesel engine exhaust emissions. Proc. Instn Mech. Engrs, Part D: J. Automobile Engineering, 2000, 214(7), 779–794.

APPENDIX 1 Notation AGTDC BGTDC BSN CA CO CO2 DFC ECU EGR GTDC HC HCCI H2 O NOx SEC SO2

after gas exchange top dead centre before gas exchange top dead centre Bosch smoke number crank angle carbon monoxide carbon dioxide digital flow controller electronic control unit exhaust gas recirculation gas exchange top dead centre hydrocarbon homogeneous charge compression ignition water nitrogen oxides specific energy consumption sulphur dioxide

APPENDIX 2 A summary of the results is given in Table 7.



Table 7

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Experimental results Value for the following

Parameter

Load (%)

Diesel

0% EGR

5%EGR

10%EGR

15%EGR

20%EGR

25%EGR

Peak pressure (bar)

0 25 50 75 100 0 25 50 75 100 25 50 75 100 25 50 75 100 25 50 75 100 25 50 75 100 25 50 75 100 25 50 75 100 0 25 50 75 100

59 67 71 78.5 82.2 197 254 322 378 452 30.3 21.3 16.7 15.4 11.9 16.9 21.6 23.4 0.27 0.17 0.13 0.13 0.65 0.37 0.32 0.88 1.30 0.93 0.78 0.75 25.38 20.65 17.92 16.01 0.3 1.1 2 2.2 3.6

51.3 62.2 71.5 76.9 80.8 201 257 304 387 481 23.4 17.4 14.1 15.1 15.4 20.7 25.6 23.9 0.30 0.22 0.12 0.14 0.22 0.25 0.32 0.75 0.84 0.73 0.64 0.69 18.01 21.80 19.04 15.63 0.1 0.5 0.8 1.1 2.1

54.6 64.8 67.9 76.8 78.7 187 242 290 363 484 23.6 17.1 14.2 13.9 15.3 21.0 25.4 25.9 0.19 0.22 0.08 0.10 0.43 0.37 0.32 0.82 0.88 0.70 0.62 0.65 12.97 15.12 13.21 10.92 0 0.1 0.6 0.9 3.5

52.6 63 67.6 73.4 77.9 182 245 292 383 463 24.5 17.3 14.6 14.3 14.7 20.8 24.7 25.2 0.18 0.13 0.16 0.14 0.22 0.37 0.40 1.70 0.93 0.74 0.66 0.67 14.48 12.81 12.25 7.76 0 0.3 1.1 1.2 4

54 62.3 67.5 73.1 76.8 184 246 291 372 452 24.6 17.3 14.4 14.8 14.7 20.8 25.1 24.4 0.15 0.20 0.13 0.16 0.43 0.37 0.47 3.02 0.95 0.75 0.67 0.65 11.12 10.26 8.58 5.66 0 0 0.6 2 3.4

52.6 60.8 66.4 73.1 76.1 185 244 289 370 418 24.8 17.1 14.4 16.6 14.5 21.1 25.1 21.6 0.12 0.20 0.13 0.21 0.45 0.37 0.55 4.37 0.99 0.75 0.68 0.51 8.80 7.76 6.56 2.44 0 0.2 0.6 2 5.2

52 61 66.2 73.1 – 182 242 286 372 – 24.2 17.0 14.5 – 14.9 21.2 24.8 – 0.31 0.18 0.15 – 0.43 0.49 0.71 – 0.95 0.75 0.66 – 5.91 5.07 4.53 – 0 0.1 0.2 1 –

Exhaust gas temperature (uC)

SEC (MJ/kWh)

Brake thermal efficiency (%) HC (gm/kWh)

CO (gm/kWh)

CO2 (gm/kWh)

NOx (gm/kWh)

Smoke (BSN)



Engineering Engineers, Part D: Journal of Automobile ...

Engineering Engineers, Part D: Journal of Automobile ...

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