Experimental investigation of autoignition of hydrogen-air charge in a compression ignition engine under dual-fuel mode

Accepted Manuscript Experimental Investigation of Autoignition of Hydrogen-Air Charge in a Compression Ignition Engine under Dual-Fuel mode

V. Chintala, K.A. Subramanian PII:

S0360-5442(17)31244-6

DOI:

10.1016/j.energy.2017.07.068

Reference:

EGY 11256

To appear in:

Energy

Received Date:

02 November 2015

Revised Date:

26 May 2017

Accepted Date:

11 July 2017

Please cite this article as: V. Chintala, K.A. Subramanian, Experimental Investigation of Autoignition of Hydrogen-Air Charge in a Compression Ignition Engine under Dual-Fuel mode, Energy (2017), doi: 10.1016/j.energy.2017.07.068

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ACCEPTED MANUSCRIPT

1

Experimental Investigation of Autoignition of Hydrogen-Air Charge in a Compression

2

Ignition Engine under Dual-Fuel mode

3 4

V. Chintala and K. A. Subramanian*

5

Engines and Unconventional Fuels Laboratory,

6

Centre for Energy Studies, Indian Institute of Technology- Delhi, New Delhi – 110 016, India.

7

*E-mail address: [email protected]

8 9

Abstract

10

High amount of hydrogen substitution in a compression ignition (CI) engine under dual fuel

11

mode is limited due to more probability of autoignition of hydrogen-air charge and knocking

12

problem. The study deals with analysis of autoignition of hydrogen-air charge in a 7.4 kW rated

13

power output of CI engine under dual fuel mode (diesel-hydrogen) at 100% load (Case I) and

14

50% load (Case II). Experimental results indicate that the significant increase in in-cylinder

15

temperature is the predominant factor for autoignition of hydrogen-air charge. The in-cylinder

16

temperature increased due to combustion advancement with hydrogen addition into the engine.

17

Computational fluid dynamics (CFD) simulation study also confirms the combustion

18

advancement with hydrogen addition in the engine. Experimental tests were extended further

19

with water injection into the engine under dual fuel mode (Case III). A clear conclusion emerged

20

from the study is that the hydrogen-air charge gets autoignite without any external ignition aid

21

when the reactants temperature is about 953 K ± 8 K. It could also be observed that knock

22

limited hydrogen energy share in the engine at 100% load was increased from 18.8% with

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conventional dual fuel mode to 60.7% with water injection due to decrease in in-cylinder

24

temperature.

25 26

Keywords: Dual-fuel engine; Hydrogen energy share; Autoignition; In-cylinder temperature;

27

Knock limited factor, Computational fluid dynamics.

28 29

1.

Introduction

30

Hydrogen (H2) is being considered as a supplementary fuel for internal combustion engines in

31

order to yield the twin benefits of energy efficiency improvement and emissions reduction [1, 2].

32

The Ministry of New and Renewable Energy, Government of India envisaged in its roadmap that

33

one million automotive vehicles are to be fueled with hydrogen by 2020 [3]. The comparative

34

analysis of hydrogen, electric and biofuel transitional pathways to a future sustainable road

35

transport in a renewable-based energy system shows that hydrogen scenario could be

36

advantageous in reducing fuel import and consumer total fuel costs [4]. Hydrogen as a fuel is

37

more suitable for spark ignition (SI) engines due to its high octane number [5]. As hydrogen

38

fueled SI engines face major setbacks of power drop and back firing, a dedicative system needs

39

to be developed for effective utilization of hydrogen in SI engines [6]. In an alternative way,

40

hydrogen could be used in compression ignition (CI) engines under dual fuel mode (hydrogen-

41

diesel) without power drop and back firing problems. In addition, there is no need of major

42

engine hardware modifications for a CI engine to be operated under dual fuel mode. Hydrogen

43

based dual fuel engines offer significant benefits such as high energy efficiency [7], high

44

combustion efficiency, lower specific energy consumption [8], near zero carbon based emissions

45

(hydro carbon (HC), carbon monoxide (CO) and smoke/particulate matter), and less greenhouse

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gas emissions (CO2, CH4 and N2O) [4, 9]. Several investigations were reported on utilization of

47

hydrogen in CI engines under dual fuel mode with exploration of improved performance and

48

emission characteristics. For example, energy efficiency improved about 14% with 20%

49

hydrogen addition in a single cylinder direct injection CI engine (7.4 kW rated power at 1500

50

rpm) under dual fuel mode [10]. Yadav et al. found 11.6% improvement in thermal efficiency of

51

a CI engine (4.4 kW rated power at 1500 rpm) with an addition of 120 g/h hydrogen under dual

52

fuel mode due to better combustion characteristics of hydrogen [11]. The maximum brake

53

thermal efficiency of 39.53% was obtained with hydrogen addition in a CI engine (5.2 kW rated

54

power at 1500 rpm) at 60% load [12]. Wu H-W and Wu Z-Y reported a significant improvement

55

in thermal efficiency of a CI engine under dual-fuel mode with 30% hydrogen energy share at

56

100% load [13]. Edwin et al. also confirmed the same that thermal efficiency of a CI engine

57

increased from 29.9% with base diesel mode to 31.6% with 10.1% hydrogen energy share at

58

100% load [14]. The reasons for this improvement could be due to better mixing characteristics

59

of these gaseous fuel with air resulting in better combustion. Even though there are significant

60

benefits in terms of thermal efficiency improvement and emissions reduction, hydrogen based

61

dual fuel engines have a severe problem of knocking with high amounts of hydrogen substitution

62

[10, 15-17]. The literature information on the maximum amount of hydrogen substituted in CI

63

engines under dual-fuel mode are summarized in Table 1 [1, 7]. It could be observed from the

64

table that the maximum hydrogen energy share achieved is in the range from 6% to 20% at

65

moderate and high loads (BMEP range: 5 bar to 9.2 bar). But, higher amount of hydrogen about

66

30% can be substituted in the engine at low load (BMEP: 2.2 bar). It could be concluded from

67

the information given in the table that the hydrogen energy share in CI engines decreases with

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increase in engine load. Hydrogen energy share in a 7.4 kW rated power output of compression

69

ignition engine with water addition at 100% load can be increased from 18.8% to 48.4% [10].

70 71

1.1. Review on knocking problem in dual fuel engines

72

Generally knocking in a SI engine occurs due to spontaneous ignition of a portion of the end gas

73

mixture in the combustion chamber ahead of the propagating flame. With the use of hydrogen fuel

74

in internal combustion engines, knocking may occur not only at the end stage of combustion

75

process but also at the earlier stage of combustion [5, 15, 26]. In hydrogen dual fuel engines,

76

knocking combustion was observed by some of the researchers during early stage of hydrogen

77

combustion [5, 15, 26]. Knocking during combustion is defined as abnormal combustion

78

phenomenon (abnormal rate of pressure rise) which degrades the engine performance. In dual fuel

79

engines, a gaseous fuel (main fuel: hydrogen) is generally injected into the intake manifold during

80

suction stroke and a liquid fuel (pilot fuel: diesel) is directly injected at the end of compression

81

stroke for initiating the combustion [27-29]. First, the diesel fuel gets self-ignited and act as an

82

ignition source for initiating the combustion of hydrogen-air mixture which is spread around the

83

combustion chamber. In contrast, if combustion of the hydrogen-air mixture is initiated by

84

hydrogen itself prior to diesel fuel injection, the combustion would proceed with severe knock.

85

Karim stated that the primary requirement of any gaseous fuel for satisfactory operation under dual

86

fuel mode is that its mixture with air would not autoignite spontaneously during or following the

87

rapid pilot energy release [30]. Hydrogen addition in a CI engine leads to production of knocking

88

or detonation because of its lower ignition energy, wider flammability range, and shorter

89

quenching distance [31]. With hydrogen fuel, knocking problem could happen not only at the end

90

stage of combustion process as in case of SI engines but also at the earlier stage of combustion

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process [5, 26, 32, 33]. High amounts of hydrogen supply to CI engines under dual-fuel mode

92

create several problems including abnormal rate of pressure rise, high in-cylinder peak pressure,

93

too advanced combustion, high in-cylinder peak temperature, autoignition of premixed hydrogen-

94

air charge, and loss of available work due to advance in start of combustion [16, 34-36]. Severe

95

knocking occurs when hydrogen is self-ignited, unlike hydrogen burning under controlled ignition

96

[15, 26]. If knock occurs, the engine would be in severe damage including breakage of piston

97

rings, piston melting, and cracking of cylinder head. Therefore, a dual fuel CI engine has to be

98

operated with lower hydrogen energy share in order to avoid knock. Edwin Geo et al. reported the

99

maximum possible hydrogen energy share without knock was about 12.7% in a single cylinder CI

100

engine (4.4 kW rated power at 1500 rpm) [14]. Yadav et al. substituted the maximum hydrogen

101

energy share of 16.4% in a CI engine (4.4 kW rated power at 1500 rpm) without knocking problem

102

[37]. Szwaja et al. concluded that addition of a small amount of hydrogen (i.e. 5% hydrogen

103

energy share) has no effect on knocking [15]. However, with increasing hydrogen energy share,

104

high frequency component of in-cylinder pressure increased substantially to 4 MPa and resulted

105

knocking at about 17% hydrogen energy share in the engine at rated load (with base compression

106

ratio of 17:1) [15]. Similarly, Chintala and Subramanian reported the knocking tendency in a

107

hydrogen based dual fuel engine (7.4 kW rated power at 1500 rpm with compression ratio of

108

19.5:1) in terms of ringing intensity and concluded in their investigation that about 19% hydrogen

109

share was the maximum amount that could be substituted in the engine at 100% load for knock

110

free operation (with base compression ratio of 19.5:1) [1]. Saravanan and Nagarajan observed

111

knocking problem at about 50% hydrogen volume share in a CI engine (3.7 kW rated power at

112

1500 rpm) under dual fuel mode at rated load [38]. Varde and Frame in their experimental study

113

on a CI engine (single cylinder direct injection engine with compression ratio of 17.4:1), measured

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acoustic noise levels in the test cell at two different locations, and observed a substantial increment

115

in the acoustic noise level beyond 11% hydrogen energy share [17]. With increasing hydrogen

116

energy share, start of combustion (SOC) advances significantly which subsequently leads to

117

knocking problem due to autoignition of premixed hydrogen-air charge [10]. With increasing

118

amount of hydrogen fuel substitution, rate of pressure rise increases at higher loads whereas it

119

decreases at low loads in a CI engine under dual-fuel mode. For example, Edwin et al. found an

120

increase in the maximum rate of pressure rise from about 5.2 bar/oCA with base diesel mode (0%

121

hydrogen energy share) to about 6.1 bar/oCA with 12.5% hydrogen energy share at 100% load

122

[14]. However, the maximum rate of pressure rise decreased from about 2.9 bar/oCA with base

123

diesel mode to about 2.3 bar/oCA with 28% hydrogen energy share at 25% load [14]. Even though

124

many studies have been reported on some combustion characteristics (in-cylinder pressure, peak

125

pressure, rate of pressure rise, and het release rate), less attention was given to an important

126

phenomenon of knocking and the reasons for knocking in case of hydrogen based dual-fuel

127

engines. Literature details on autoignition of gaseous fuel-air mixture in dual fuel engines are given

128

below;

129 130

1.2. Review on autoignition of gaseous fuel-air mixture in dual fuel engines

131

Autoignition is the term used for a rapid chemical reaction of fuel-air mixture which is not

132

initiated by any external ignition source [39]. The autoignition of a gaseous fuel-air mixture

133

occurs when the energy released by the reaction as heat is larger than the heat lost to the

134

surroundings. Tsujimura et al. studied autoignition of hydrogen-air mixture in a constant volume

135

chamber and concluded that intake air temperature has significant effect on autoignition of the

136

mixture [40]. For the intake air temperatures below 1100 K, the auto-ignition linearly depends on

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the temperature in the Arrhenius coordinates whereas for the temperatures greater than 1100 K,

138

the temperature dependence of the auto-ignition is weak [40]. Similarly, Naber and Siebers

139

obtained a strong correlation between ignition delay and intake air temperature during the

140

investigation on a constant volume combustion chamber under diesel engine condition [33]. Start

141

of combustion advances with hydrogen addition in a CI engine under dual fuel mode. For

142

example, Szwaja et al. reported about 4o CA combustion advancement with increasing hydrogen

143

energy share to 25% in a direct injection CI engine [15]. Lata et al. found a significant

144

advancement (about 10o CA) in combustion of fuel-air mixture with increase in hydrogen

145

volume share from 4% to 16% in a multi-cylinder CI engine [41]. This may be due to hydrogen-

146

air mixture surrounding the pilot diesel fuel promotes faster initial combustion reaction rate and

147

further increase of in-cylinder temperature [41]. The start of combustion advanced by 5o CA with

148

30% hydrogen addition in a diesel engine (3.7 kW at 1500 rpm) under dual fuel mode due to

149

instantaneous combustion of gaseous-air mixture [38]. Research findings of Miyamato et al.

150

indicate that hydrogen in a CI engine was autoignited by itself without ignition assistance by

151

pilot fuel (diesel) when hydrogen fraction is higher than 8% volume in a CI engine [34].

152

Similarly, spontaneous autoignition of air-gaseous fuel (propane) was found in a diesel-propane

153

dual fuel engine due to its high reactivity nature like hydrogen fuel [35]. Wong and Karim stated

154

the reasons for autoignition of hydrogen-air charge are due to high polytropic index of hydrogen,

155

higher in-cylinder temperature, and increasing preignition chemical reactions [36].

156 157

It is clearly understood from the literature information that the start of combustion of diesel-

158

hydrogen-air mixture advances with increase in amount of hydrogen addition along with diesel

159

fuel in a CI engine. If the start of combustion advances drastically, autoignition of hydrogen-air

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may takes place, which further leads to knocking and suppress the amount of hydrogen

161

substitution in the engine. A very few investigations on influence of high amount of hydrogen

162

substitution on autoignition of premixed hydrogen-air charge in dual fuel engines were reported

163

in the literature. However, the reasons for autoignition of the charge in dual fuel engines are not

164

addressed properly. In this context, the present study is aimed at investigation of effect of high

165

hydrogen energy share on autoignition of hydrogen-air charge in a single cylinder hydrogen dual

166

fuel engine (7.4 kW rated power at 1500 rpm). The effects of various parameters such as in-

167

cylinder pressure, in-cylinder temperature, and gaseous fuel concentration on autoignition of the

168

charge, knocking, and maximum achievable hydrogen energy share are analyzed. Experimental

169

tests were also conducted with water injection in the dual fuel engine, to assess the effect of in-

170

cylinder temperature on autoignition of hydrogen-air charge. A computational fluid dynamics

171

(CFD) study was carried out to support the fact of combustion advancement/autoignition with

172

hydrogen addition in a CI engine.

173 174

2.

Experimental tests and simulation details

175

2.1. Experimental test setup for dual fuel operation

176

An experimental test setup was developed by modifying a 7.4 kW direct injection CI engine to

177

hydrogen based dual-fuel mode. A photographic view of the experimental setup is shown in

178

Figure 1 and technical specifications of the engine are given in Table 2. The test engine was

179

loaded with an eddy current dynamometer as shown in the figure. In the present study, hydrogen

180

was used as main fuel and diesel as pilot fuel. Hydrogen was injected into the intake manifold

181

using a solenoid gas injector and the liquid fuel (diesel) was directly injected into the combustion

182

chamber using a conventional high pressure liquid injector for initiating the ignition of

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hydrogen-air mixture. The physiochemical properties of the fuels are given in Table 3 [7, 9]. An

184

electronic control unit (ECU) was used to adjust the hydrogen injection timings including start

185

and duration of fuel injection. Hydrogen was injected into the intake manifold after the outlet

186

valve closed (43o CA after TDC) in order to avoid scavenging losses [29]. Start of hydrogen

187

injection was maintained constant as 43o CA after TDC throughout the experimentation whereas

188

the hydrogen injection duration was varied with respect to the engine loading. These injection

189

timings were optimized based on better performance and lower emissions of the engine in the

190

earlier study [29].

191 192

2.2.

Brief details of water injection system

193

The water injection system consists of water reservoir, submersible pump, solenoid water

194

injector, pressure indicator, flow control valves, ECU, and electronic weighing balance as shown

195

in Figure 2. The ECU, which was used for hydrogen gas injection, controls the start and end of

196

water injection timings. Timed intake port injection system was used for water injection into the

197

dual fuel engine. Water was injected into the intake port during the suction stroke at the pressure

198

of 2-3 bar. Experimental tests were carried out on the dual-fuel engine for Specific Water

199

Consumption (SWC) of 480 g/kWh at 100% load and a constant speed of 1500 rpm. The start of

200

injection for hydrogen and water kept constant as 43o and 45o CA-after TDC throughout the

201

experiments [8, 10]. The end of injection for both fluids (hydrogen and water) varied

202

independently with respect to hydrogen energy share and water consumption. Based on the

203

maximum available time for reaching the injected hydrogen/water from the point of injection to

204

the engine cylinder, the duration of injection was optimized for hydrogen/water [29].

205

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206

2.3.

Engine combustion diagnosis system

207

The engine combustion diagnosis system consists of an inbuilt charge amplifier, voltage

208

amplifier, and data acquisition system (Figure 3). The system was mainly attached with three

209

kinds of sensors; (i) piezoelectric pressure transducer (ii) optical encoder and (iii) piezoelectric

210

strain gauge. The piezoelectric pressure transducer with nominal sensitivity 45 Pico coulomb/bar

211

was flush mounted on the cylinder head of the engine, for measurement of instantaneous in-

212

cylinder pressure data during the engine operation. The optical encoder (720 pulses/revolution)

213

was mounted on one end of crankshaft of the engine for crank angle (CA) measurement with an

214

accuracy of 0.1o CA. The piezoelectric strain gauge pressure transducer (pressure measurement

215

range: 0 to 2000 bar) was fitted on high-pressure diesel pilot fuel line to measure the pilot fuel

216

injection pressure. The in-cylinder pressure analog signal was amplified by the charge amplifier

217

and then, the analog signal was converted to a digital signal by data acquisition system for

218

further processing of the acquired data. The in-cylinder pressure signal, in-line pilot fuel pressure

219

signal, and TDC encoder signals were processed together by the combustion analyzer. The post

220

processing software was used for processing of pressure-crank angle data that captured for 200

221

consecutive engine operating cycles.

222 223

2.4.

Details of experimental tests

224

Experimental tests were carried out on the engine under hydrogen based dual fuel mode at a

225

constant speed of 1500 rpm for 100% load (Case I) and 50% load (Case II). Experiments on the

226

dual fuel engine were further extended with water injection (Specific Water Consumption (SWC)

227

of 410 g/kWh) at 100% load and a constant speed of 1500 rpm (Case III). The SWC could be

228

determined using Eq. (1) [8]. As the study is mainly focused on autoignition of the hydrogen-air

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229

charge in a dual fuel engine, Case I is analyzed in detail whereas remaining other two cases

230

(Case II and Case III) are discussed briefly for understanding the in-cylinder temperature effect.

231

Details of measurement range and resolution of the instruments used in the study are given in

232

Table 4. The uncertainty associated with the various parameters including hydrogen energy

233

share were computed using Eq. (2) [7]. A computational fluid dynamics (CFD) study was also

234

carried out to understand the phenomenon of combustion advancement/autoignition with

235

increasing amount of hydrogen substitution in the engine (details are given in Section 2.5).

236 237

mwater

Specific water consumption (SWC) = Brake power × 3600

(1)

238

[

∂q

∂q

∂q

1

2

n

]

239

∆q = (∂x ∆x1)2 + (∂x ∆x2)2 + ………… + (∂x ∆xn)2

0.5

240

Where,

241

q

: f (x1, x2,…..xn) (q is the function of x1, x2,…..xn)

242

x1, x2,…..xn

: Measured variables

243

∂q ∂q ∂q ∂x 1, ∂x 2,……., ∂x n

(2)

: Partial differential of calculated parameter q which depends on variables

244 245

2.5. Brief details of computational fluid dynamics (CFD) simulation

246

A three dimensional hexahedral mesh for the engine’s combustion chamber geometry (a sector of

247

72 degrees consisting cylinder and piston: 47270 Cells, 52548 nodes, and 136701 quadrilateral

248

interior faces) was created using GAMBIT software (Figure 4) and then numerical simulation was

249

carried out using ANSYS FLUENT software. Dynamic mesh model was used for dynamic motion

250

of piston where the shape of the domain is continuously changed with respect to time. Mesh

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251

independence was also verified with variation of mesh elements of the given geometry. A time

252

step of 0.1o CA was selected for simulation during combustion period. Various models used in

253

CFD simulation study in the hydrogen based dual fuel engine are given in Table 5. It may be noted

254

that as the study is mainly focused on the experimental investigations, a few details of CFD

255

simulation are presented in this section in support of combustion advancement/autoignition at high

256

amounts of hydrogen energy share in the engine.

257 258

3.

Methodology

259

3.1. Determination of autoignition of hydrogen-air mixture

260

In hydrogen based dual-fuel engines, typically diesel is used to initiate the ignition whereas

261

hydrogen is used as main fuel (main energy source). As hydrogen gas is injected into the intake

262

manifold of the engine along with intake air during suction stroke, the gaseous fuel could well mix

263

with the air, and forms high degree of homogeneous premixed hydrogen-air charge. When diesel

264

fuel is directly injected (sprayed) into the combustion chamber at the end of compression process,

265

entrainment of the premixed hydrogen-air charge into the diesel spray takes place and forms diesel-

266

hydrogen-air mixture. In this diesel-hydrogen-air mixture, the amount of diesel fuel is dominant

267

than the amount of hydrogen and hence the mixture is simply considered as diesel-air mixture.

268

Combustion initiates due to autoignition of diesel-air mixture as the in-cylinder temperature at the

269

end of compression (about 700 K) is higher than the autoignition temperature of diesel fuel (about

270

530 K). However, if the amount of hydrogen exceeds beyond a certain limit, autoignition of

271

premixed hydrogen-air charge could takes place without diesel aid, which may result in

272

combustion with knock. Autoignition of hydrogen-air charge could be defined as the initiation of

273

combustion of the premixed hydrogen-air charge without diesel ignition source (i.e., combustion

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starts prior to the diesel injection). With low amount of hydrogen addition in to a CI engine

275

(Operation 1), generally combustion starts (SOC 1) after pilot diesel fuel injection (SOI 1),

276

whereas with high amount of hydrogen addition, combustion (SOC 2) may start earlier than the

277

diesel injection (SOI 2) as shown in Figure 5. Start of diesel fuel injection and start of combustion

278

of fuel-air charge (diesel-air (Operation 1) or hydrogen-air (Operation 2)) are together represented

279

in Figure 5. Later case (Operation 2) indicates autoignition of premixed hydrogen-air charge which

280

subsequently may create a problem of severe knock, whereas Operation 1 indicates a typical

281

combustion process in a dual-fuel engine without knock. Polk et al. expressed the similar tendency

282

of autoignition of propane-air mixture in a 1.9 liter turbo charged CI engine under dual fuel mode

283

[35].

284 285

3.2. Activation energy of hydrogen-air charge under dual fuel mode

286

The minimum ignition energy required for initiating chemical reaction is known as activation

287

energy. As the minimum ignition energy of hydrogen (0.02 mJ) is lower than the conventional

288

diesel fuel (0.6 mJ), the hydrogen promotes combustion reactions in a faster way. It may be noted

289

that pre-ignition reaction rate increases significantly with increasing hydrogen energy share in the

290

engine under dual fuel mode. The activation energy for hydrogen-air charge could be determined

291

using Eqs. (4-7) which were obtained by plotting log ID (Ignition delay) as a function of 1/T (T:

292

In-cylinder temperature) (Figure 6). Ignition delay of any fuel in a CI engine could be defined as

293

the time interval (or crank angle duration) between the start of injection of fuel and start of

294

combustion. The start of injection of fuel is usually taken as the time when the injector needle lifts

295

of its seat. The start of combustion could be defined as the change in slope of the heat release rate

296

from negative to positive. In the present study the start of injection was determined using fuel

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297

injection pressure measurement i.e., the degree crank angle at which the injector nozzle starts open

298

at 250 bar pressure (nozzle opening pressure : 250 bar). The start of combustion was determined

299

using heat release rate, determined from in-cylinder pressure data with respect to degree crank

300

angle. Similarly in-cylinder temperature with respect to degree crank angle was determined using

301

the ideal gas equation with the input data of measured in-cylinder pressure [16]. The activation

302

energy of the hydrogen for initiating ignition under dual fuel mode is found to be 74992 J/mol at

303

950 K. The data is almost equal to the literature data reported by Chen, et al. that the activation

304

energy for the reaction of hydrogen-air charge at 900 K is about 79000 J/mol [43].

305 1

306

For τi = f (T)

307

Where

308

τi

= Ignition delay, ms

309

T

= In-cylinder temperature at the time of pilot fuel injection, K

310

From Figure 6 the ignition delay could be expressed as Eq. (4),

311

ln(τi)

312

The above equation could be rewritten as Eq. (5).

313

τi

314

The general equation for determining ignition delay that could be expressed as Eq. (6).

315

τi = μ exp (

316

Where µ = Constant

317

(3)

= 9020 (1/T) - 0.7899

= 2.2032 exp (

9020 T )

9020 T )

(4)

(5)

(6)

T = In-cylinder temperature

318

From Eq. (5) and Eq. (6), the activation energy of hydrogen-air mixture could be obtained as

319

follows, 14

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320

For Ea/Ru = 9020 and Ru = 8.314 J/mol-K,

321

The activation energy (Ea) = 74992 J/mol

322 323

It may be noted that the products (air-fuel charge/burned products) in the combustion chamber are

324

assumed as ideal gas which is used for calculating in-cylinder temperature with respect to degree

325

crank angle (Eqs. (7-9)).

326

p v = mRT

(7)

327

T = pv/mR

(8)

d dθ(T)

328

=

d 𝑝𝑣 dθ(mR)

(9)

329

Assuming no variation in mass of air fuel charge (m) and characteristic gas constant (R) with

330

respect to degree crank angle the Eq.( 9) could be written as given in Eq.( 10).

331

dT dθ

dv

=

dp

𝑝dθ + v dθ

(10)

mair ‒ fuel charge R

332 333

3.3 Knock limited factor for maximum hydrogen energy share

334

Abnormal rate of pressure rise (RPR) during combustion was considered as knocking in the

335

present study. Heywood stated the maximum RPR for knocking combustion is about 10 bar/

336

oCA

337

reaction rate which is described by Arrhenius Eq. (11) [39].

in case of CI engines [39]. It is well known that the RPR is strong dependence on oxidation

338 Ea

339

RPR (Maximum H2 energy share limit) ∝ exp

(R T) u

(11)

340 341

Where 15

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342

Ea

= Activation energy of species

343

Ru

= Universal gas constant

344

T

= Reaction temperature (or) In-cylinder temperature

345

It may be noted that compression ratio and equivalence ratio are the other major influencing

346

design and operating parameters for knocking problem in the engine. The onset of knock in the

347

engine is inversely proportional to these parameters (Eqs. (12, 13)) [44]. Reduction in

348

compression ratio of the engine leads to decrease in the in-cylinder temperature which allows

349

higher energy share in the engine without knock. Typically low temperature combustion

350

strategies such as water injection, retarded diesel fuel injection, and compression ratio reduction

351

could increase the amount of hydrogen substitution in a CI engine [8].

352 353

1

(12)

Maximum H2 energy share limit ∝ compression ratio (CR)

354 355

1

Maximum H2 energy share limit ∝ Equivalence ratio(∅)

(13)

356 357

It could be observed from the literature that increase in hydrogen energy share in a CI engine

358

leads to autoignition of the hydrogen-air charge and subsequently to onset of knock [10, 34] i.e.,

359

knocking is directly proportional to hydrogen energy share in the engine (Eq. (14)).

360 361

Knock in the engine ∝ H2 energy share

(14)

362 363

Knock limited factor (KLF) could be defined as a limiting value at which maximum hydrogen

364

could be substituted in a CI engine without knock. With the above inferences, the KLF for 16

ACCEPTED MANUSCRIPT

365

maximum hydrogen energy share could be expressed as given in Eq. (15). The KLF is a

366

dimensionless indicator that could be used for finding maximum achievable hydrogen energy

367

share in a CI engine under dual fuel mode.

368 Ea

(R T)

369

Knock limited factor (KLF) = k (H2energy share)α(∅) ‒ β(CR) ‒ γexp

370

Where

371

Constant, k = 1.2 (depends on engine configuration and loading conditions)

372

Hydrogen energy share exponent, α = 3.5

373

Equivalence ratio exponent, β = 0.351

374

CR exponent, γ = 0.3

u

(15)

375 376

4.

Results and discussion

377

The concept of autoignition of premixed hydrogen-air charge in the hydrogen dual fuel engine

378

was introduced in the earlier study [1], but its detailed analysis was not provided. In the present

379

study, a detailed analysis with wide range of experimental tests including water injection in the

380

test engine is provided. Effects of high amount of hydrogen addition in the engine on combustion

381

advancement and subsequently on autoignition of hydrogen-air charge are discussed in this

382

section. The main reasons for autoignition of premixed hydrogen-air charge in the dual fuel

383

engine are also addressed.

384 385 386 387

4.1. Effect of high amount of hydrogen addition on combustion advancement in the dual fuel engine 4.1.1. Experimental test results of combustion advancement 17

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388

Figure 7 shows variation of liquid fuel injection pressure and heat release rate with respect to

389

degree crank angle of the engine under dual fuel mode at 100% load (Case I). It could be

390

observed from the figure that start of combustion (SOC) advanced with increase in hydrogen

391

energy share. This advancement in combustion occurred with larger degrees crank angle at high

392

amount of hydrogen substitution as compared to low amount of hydrogen substitution. For

393

example, the SOC was advanced about 1.8o CA duration with 14.5% hydrogen energy share

394

whereas with 18.8% hydrogen energy share the SOC was advanced about 6o CA duration as

395

compared to base diesel mode. The SOC was advanced from 1.9o CA after TDC with 0%

396

hydrogen energy share (diesel mode) to 4.1o CA before TDC with 18.8% hydrogen energy share

397

in the engine under dual fuel mode. The main reason for the combustion advancement could be

398

due to increase in the in-cylinder temperature with hydrogen addition.

399 400

4.1.2. CFD simulation results of combustion advancement

401

It may be noted that the fact of advancement in combustion with increasing hydrogen energy

402

share in the dual fuel engine is also supported by CFD simulation results. Progress of combustion

403

in the engine cylinder is illustrated with in-cylinder temperature contours as shown in Figure 9.

404

The degree crank angle at which a spontaneous increase in the in-cylinder temperature occurs

405

could be indicated as start of combustion in the engine. It is evident from the CFD simulation

406

results (Figure 9) that the combustion in the cylinder advanced with increase in hydrogen energy

407

share. For example, the in-cylinder temperature was spontaneously increased at 2.6o CA after

408

TDC with base diesel mode (0% hydrogen energy share), whereas it advanced to 0.9o and 0.6o

409

CA after TDC with 14.5% and 16.7% hydrogen energy shares (Figure 9). Similarly from the

410

experimental test results, it could be observed that the start of combustion was advanced from

18

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411

1.9o ATDC with base diesel mode (0% hydrogen energy share) to 0.3o ATDC (after TDC) and

412

0.6o BTDC (before TDC) and 4.1o BTDC with 14.5%, 16.7% and 18.8% hydrogen energy shares

413

respectively as shown in Figure 7. Simulated in-cylinder pressure and in-cylinder temperature

414

results obtained from the CFD analysis on Mesh type 3 were validated with the experimental

415

results for the same engine operating conditions. Figure 8 shows validation of CFD simulation

416

results of in-cylinder pressure and in-cylinder temperature with the experimental results for base

417

diesel and dual-fuel modes at 100% load. In-cylinder peak pressure was slightly over predicted

418

(particularly in premixed combustion phase) and subsequently this error was amplified in case of

419

in-cylinder temperature. It could also be clearly seen that start of combustion (SOCcfd) was

420

advanced in CFD simulation as compared to experimental results (SOCexp) [45]. For example, at

421

16.7% hydrogen energy share SOCcfd occurred at 2o CA after TDC whereas SOCexp occurred at

422

4o CA after TDC. This over prediction of in-cylinder pressure and in-cylinder temperature could

423

be due to neglecting crevice volumes, blow-by losses, mismatching of compression ratio and

424

uncertainty associated with equivalence ratio. Similar type of trends was reported by Maghbouli

425

et al. in their CFD simulation work in a dual-fuel engine [46]. Liu and Karim also expressed the

426

similar results of higher in-cylinder peak pressure value (about 78 bar) with CFD simulation than

427

the experimental in-cylinder pressure about 74 bar in a dual-fuel engine [47]. The variations in

428

the in-cylinder pressure and in-cylinder temperature are affected by the total kinetic energy

429

variations and a slight pressure difference existing in the main combustion chambers [47].

430 431

Localized maximum in-cylinder temperature increased to about 2100 oC spontaneously (in all

432

cases) at the start of combustion as seen in the figure. Average temperature scale was represented

433

at high degree crank angles (i.e., at 3o, 5o, and 10o etc.). It may be noted that the phenomenon of

19

ACCEPTED MANUSCRIPT

434

autoignition couldn’t be represented in a single domain using CFD simulation results

435

(temperature/ pressure/ species concentration) due to the software constraints (i.e., simultaneous

436

representation of both start of diesel fuel injection and start of combustion in a single domain

437

can’t be done). However, the in-cylinder temperature contours could be used as indirect

438

representation of autoignition concept in the engine. When ignition of the fuel-air charge occurs

439

inside the combustion chamber, in-cylinder temperature increases spontaneously. Santoso et al.

440

also used in-cylinder temperature contours (obtained from CFD simulation) for representing

441

combustion progress in a hydrogen based dual fuel engine [48].

442 443

It could also be observed from Figure 9 that the temperature contours are widespread towards

444

cylinder walls (wide propagation of flames towards end charge) in the combustion chamber with

445

hydrogen addition in the engine. This could be interpreted as due to rapid combustion of

446

hydrogen-air mixture in the engine under dual fuel mode. For a particular crank angle rotation

447

(for example at 3o CA), it could be observed qualitatively from Figure 9 that flame propagation

448

distance increased significantly towards the cylinder walls with increase in hydrogen energy

449

share. However, a detailed CFD study needs to be carried out in order to assess the

450

characteristics of flames (i.e., multiple ignition centers, flame propagation, and flame quenching

451

etc.) in dual fuel engines.

452 453

4.2. Effect of high amount of hydrogen addition on autoignition of hydrogen-air mixture

454

It is clearly seen in Figure 7 that until 16.7% hydrogen energy share, SOC occurred after the

455

injection of diesel fuel (ignition source), which indicates the autoignition of diesel-air charge

456

(diesel dominant diesel-hydrogen-air mixture) occurred first and later hydrogen-air mixture

20

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457

(spread around the diesel spray) combusted in the cylinder. However, with 18.8% (about 19%)

458

hydrogen energy share, combustion started earlier (4.1o CA before TDC) than diesel fuel injection

459

(1.8o CA before TDC) which indicates autoignition of hydrogen-air charge without any external

460

ignition source (pilot diesel fuel). Further enhancement of hydrogen energy share leads to onset of

461

knock resulting in sharp increase in the in-cylinder peak pressure to about 90 bar (maximum

462

allowable peak pressure). Due to the autoignition of hydrogen-air mixture, the amount of hydrogen

463

substitution is restricted up to about 19% in the engine under conventional dual fuel mode at 100%

464

load. However, the maximum hydrogen energy share in the engine increased drastically from 19%

465

with duel fuel mode at 100% load (Case I) to 48.3% and 60.7% with dual fuel mode at 50% load

466

(Case II) and water added dual fuel mode (Case III). At 19% hydrogen energy share, there is no

467

such problem of autoignition in Case II and Case III operations. It could be observed from Figure

468

10 that at 50% load (Case II), there was no autoignition of hydrogen-air mixture for all hydrogen

469

energy shares ranging from 0% to 48.3%. This indicates the engine runs smoothly for all hydrogen

470

energy shares as the combustion (SOC) proceeded after the pilot diesel fuel injection as in typical

471

dual fuel operation. The equivalence ratio in the engine under dual-fuel mode decreased with

472

increase in hydrogen energy share at all loads. At 100% load, the equivalence ratio decreased from

473

0.4 with base diesel mode to 0.18 with diesel hydrogen dual-fuel mode (18.8% hydrogen energy

474

share). It may be noted that the equivalence ratio at higher and lower flammability limits for

475

hydrogen-air mixture are about 1.5 and 0.1. As the equivalence ratio of premixed hydrogen-air

476

mixture is under the flammability limits, there are more chances of autoignition of the charge

477

without diesel ignition source. It could be observed from Figure 11 (Case III) that the occurrence

478

of autoignition of hydrogen-air charge was suppressed until 56.5% hydrogen energy share, but

479

beyond this energy share autoignition of the premixed hydrogen-air charge problem was occurred.

21

ACCEPTED MANUSCRIPT

480

The primary reason is due to substantial reduction in the in-cylinder temperature by the addition

481

of water which has high specific heat and also it dilutes the charge. The added water does function

482

of slowing down the reaction rate of hydrogen-diesel-air during combustion by reducing the in-

483

cylinder temperature and enhancing the charge dilution. With water addition in the engine (Case

484

III), knock limited hydrogen energy share (without autoignition problem) was increased from

485

18.8% to 56.5% as shown in Figure 11. Until 56.5% hydrogen energy share, start of combustion

486

occurred after the pilot fuel injection (diesel), which indicates the normal combustion process.

487

However, beyond 56.5% hydrogen energy share, start of combustion (5o CA before TDC) was

488

occurred prior to the fuel injection (1o CA before TDC) due to autoignition of hydrogen-air charge.

489 490

Similar findings were reported by Miyamoto et al. that the hydrogen was autoignited without the

491

ignition aid of pilot diesel fuel in a diesel-hydrogen dual fuel engine [34]. They found that with

492

high amount of hydrogen substitution, heat energy released prior to start of diesel injection, and

493

the heat energy release continued even after the diesel fuel cut off [34]. Experimental

494

investigations carried out by Polk et al. supported the fact of spontaneous autoignition of propane

495

in a diesel-propane based dual fuel engine due to its high reactivity [35]. They stated the reasons

496

for propane autoignition are due to high in-cylinder temperature and high equivalence ratio that

497

are conducive to rapid preignition. In case of hydrogen dual fuel mode, the reasons for

498

autoignition of hydrogen-air charge could be due to high in-cylinder temperature [49], rapid

499

preignition reactions [36], production of high concentration of free radicals (O, H, and OH) [50],

500

reduction in flame nucleation period [50]. In the present study, some of these reasons such as the

501

increase in the in-cylinder temperature and free radical concentration for autoignition of

502

hydrogen-air charge in the engine are discussed as given below;

22

ACCEPTED MANUSCRIPT

503 504

4.2.1. Effect of in-cylinder temperature on autoignition of hydrogen-air charge

505

It may be noted that at 19% hydrogen energy share, autoignition of hydrogen-air charge occurred

506

in Case I operation (conventional dual fuel mode at 100% load), whereas no such problem

507

occurred in Case II (conventional dual fuel mode at 100% load) and Case III (water added dual

508

fuel mode at 100% load) operations. The main reason could be due to lower in-cylinder

509

temperature with water injection and lower load (50%) operation than high load (100%)

510

operation. For example, at 19% hydrogen energy share, the in-cylinder peak temperature

511

decreased from 1876 K with conventional diesel-hydrogen dual fuel mode at 100% load to 1680

512

K with water injected dual-fuel mode at 100% load and 1564 K with conventional diesel-

513

hydrogen dual fuel mode at 50% load (Figure 12). With this reduction in the in-cylinder

514

temperature, the possibilities of autoignition and knocking were reduced, that could lead to

515

significant enhancement of hydrogen energy share under dual fuel mode.

516 517

Autoignition-temperature of a fuel could be defined as the temperature at which the fuel will

518

spontaneously ignite [39]. Leishear measured the relationship between pressure and autoignition

519

temperature of hydrogen-air mixture at stoichiometric condition in a piping system used in

520

nuclear reactor [51]. Typically the stoichiometric hydrogen-air charge gets autoignition at

521

temperature of 580 oC and atmospheric pressure of 1 bar [51]. In case of internal combustion

522

engines, the autoignition-temperature of fuel-air charge depends on various parameters including

523

pressure and equivalence ratio. For example, at an equivalence ratio of 1 (stoichiometric

524

condition), autoignition temperature of hydrogen-air charge decreased from 580 oC to 440 oC due

525

to increase in pressure from 1 bar to 40 bar [51]. Similarly, in the present study, autoignition-

23

ACCEPTED MANUSCRIPT

526

temperature of hydrogen-air charge in the engine is found about 670 oC at an equivalence ratio of

527

0.4 and the in-cylinder pressure about 54 bar (under conventional hydrogen-diesel dual fuel

528

mode). The in-cylinder temperature at the end of compression stroke is reached to the range of

529

600 oC to 700 oC. This temperature is enough for initiation of autoignition of hydrogen-air

530

charge in the engine cylinder. These experimental results are in agreement with the literature

531

data of autoignition temperature of the charge is about 650 oC to 700 oC at equivalence ratios of

532

0.4 to 0.6 [30, 44]. Daeyup and Hochgreb reported autoignition of premixed hydrogen-air

533

mixture could take place in reactant’s pressure range from 4 bar to 40 bar and temperature

534

range from 950 K -1050 K [52]. However, the relationship between autoignition temperature and

535

in-cylinder pressure with respect to equivalence ratio for dual fuel engines needs to be studied.

536 537

A summary of in-cylinder pressure and autoignition temperatures with respect to different

538

hydrogen energy shares under different experimental test cases is given in Table 6. Experimental

539

tests were conducted on the same test engine with retarded diesel injection timing under dual fuel

540

mode in the earlier study [8]. A conclusion emerged from these results is that the premixed

541

hydrogen-air charge gets autoignition at the in-cylinder temperature is about 953 K ± 8 K with

542

the corresponding in-cylinder pressure of 56 bar ± 3 bar in the engine. Wong and Karim studied

543

the effect of in-cylinder temperature variation on autoignition of three gaseous fuels such as

544

methane (CH4), propane (C3H8), and hydrogen (H2) in a diesel engine [36]. The fuel-air charge is

545

able to reach a higher peak temperature with hydrogen due to its higher polytropic index as

546

compared to other two gaseous fuels. They also found that the hydrogen-air charge did not

547

autoignite at low in-cylinder temperature whereas the propane and methane fuels were

548

autoignited even at low temperatures [36]. Hence, the problem of autoignition at high amount of

24

ACCEPTED MANUSCRIPT

549

hydrogen substitution could be resolved with low temperature combustion strategies such as

550

retarded injection timing of diesel fuel, addition of diluents (nitrogen and carbon dioxide), water

551

injection, and compression ratio reduction.

552 553

4.2.2. Effect of free radicals concentration on autoignition of hydrogen-air charge

554

It is reported that the increase in concentration of hydrogen increases the potential for free

555

radicals (H, O, and OH) production during its oxidation process [50]. High amount of hydrogen

556

substitution in the dual fuel engine may produce a larger pool of H, O, and OH radicals at an ear-

557

lier stage of the combustion process. High concentration of these radicals have significant effects

558

on autoignition of hydrogen-air mixture and overall combustion reaction rate during the engine

559

operation. These effects include extension of the lean limit, increasing EGR/dilution tolerance,

560

and shortening of the flame nucleation period, thereby increasing the heat release rate [50].

561

Increase in the in-cylinder temperature could be one of the major cause for production of high

562

levels of free radicals in the dual fuel engine.

563 564

4.2.3. Effect of residual gases on autoignition of hydrogen-air charge

565

The other reasons for autoignition of hydrogen-air charge in the dual fuel engine could be the

566

gaseous fuel concentration in residual gases of the previous cycle and high temperature of the

567

residual gases [36, 44]. These existing species in the residual gases can play important chemical

568

and thermal roles in the preignition reaction processes of the next cycle. The presence of residual

569

gas could alter the in-cylinder temperature of the hydrogen-air charge at the beginning of

570

compression process. The kinetic effect of these residual gases may cause an increase in the

571

preignition reaction activity of the gaseous fuel-air charge which leads to autoignition of the

25

ACCEPTED MANUSCRIPT

572

charge [36]. At higher load and higher concentration of the hydrogen, flame which is initiated

573

from the various ignition centers of the diesel fuel, propagates at a faster rate and consumes the

574

most of the cylinder gaseous fuel-air mixture in the preceding cycles which results in higher

575

residual gas temperature in the following cycles [50]. The residual gas with some partial

576

oxidation products from the preceding cycles would be a source of active radicals for the

577

following cycle. It is evident from research findings of Wong and Karim that the residual gas

578

(EGR) has high potential to enhance the preignition reaction rate in dual fuel engines [36].

579 580

4.3. Knock limited factor for maximum hydrogen energy share

581

From the wide range of experimental tests carried out with different engine operating conditions

582

(Case I to Case III), it is found that the knock limited factor (KLF) for maximum hydrogen

583

energy share under dual fuel mode is a function of hydrogen energy share, in-cylinder pressure,

584

equivalence ratio, compression ratio, in-cylinder temperature, and activation energy of hydrogen-

585

air mixture. The following critical conditions for KLF were observed from the experimental test

586

results under hydrogen dual fuel mode.

587 588 589

If

KLF < 1 ----> Less probability for knocking KLF > 1 ----> More probability for knocking

590

For Case (I) operation, the KLF was increased to 0.9 with 16.7% hydrogen energy share and then

591

it reached to 1.1 with 18.8% hydrogen energy share as shown in Figure 13. Similarly with Case

592

(II) and Case (III) operations, the KLF was under limit until the hydrogen energy shares of 18%

593

and 56.5% and beyond these energy shares the KLF was crossed the limit exponentially as

594

shown in the figure.

26

ACCEPTED MANUSCRIPT

595 596

5.

597

The effect of high percentage of hydrogen energy share on autoignition of premixed hydrogen-

598

air charge was analyzed in a single cylinder hydrogen based dual fuel engine. The following

599

conclusions are drawn based on the experimental results.

600

Conclusions

 High in-cylinder temperature which is the predominant factor, influences the autoignition

601

of premixed hydrogen-air charge. The premixed hydrogen-air charge could get self-

602

ignition (autoignition) without external ignition aid (diesel pilot fuel) when the

603

temperature of the reactants is about 953 K ± 8 K with the corresponding in-cylinder

604

pressure of 56 bar ± 3 bar. Similarly, the auto-ignition temperature and pressure of any

605

compression ignition engine under dual fuel mode (Diesel-Hydrogen) could be found out

606

using the methodology emerged from this study. Hence, the critical energy share of

607

hydrogen in a dual fuel engine working under specific design and operating parameters

608

could be calculated within reasonable accuracy using this auto-ignition temperature and

609

pressure predicted using this study’s methodology.

610

 Increase in hydrogen energy share in the engine enhances the degree of advancement in

611

start of combustion and the reactant’s in-cylinder temperature. Too advance in start of

612

combustion with high amount of hydrogen energy share (beyond a critical limit) leads to

613

more probability of auto-ignition which may result in severe knocking problem.

614

 Knock limited factor (KLF) is a notable outcome emerged from this study as the

615

dimensionless indicator could be useful to predict qualitatively the maximum hydrogen

616

energy share in the engine under dual fuel mode. If the KLF is greater than 1, the

617

probability of knock would be higher whereas less than one means lower probability.

27

ACCEPTED MANUSCRIPT

 The probability of auto-ignition of hydrogen-air charge can be minimized by reducing

618 619

reactant’s temperature during ignition period using suitable techniques including water

620

injection. For example, it is confirmed from the experimental results that knock limited

621

maximum hydrogen energy share in a dual fuel diesel engine (7.4 kW at 100% load) was

622

found as 18.8% under conventional dual fuel mode and this critical energy share could be

623

enhanced to 60.7% with water injection in the dual fuel engine for the same engine

624

operating conditions.

625 626

Abbreviations

627

Ea

: Activation energy of hydrogen, J/mol

628

R

: Characteristic gas constant, J/mol

629

SWC

: Specific water consumption, g/kWh

630

T

: Absolute temperature/In-cylinder temperature, K

631

p

: In-cylinder pressure, N/m2

632

V

: Instantaneous cylinder volume, m3

633 634 635

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Figures:

776 777

Figure 1 Photographic view of experimental setup for dual fuel mode

778

35

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779 780

Figure 2 Water injection system for the engine under diesel-hydrogen dual-fuel mode

781 782

Figure 3 Schematic diagram of combustion analysis system

36

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784 785

Figure 4 (a) Hexahedral mesh of piston and cylinder head for sector of 72 degree (b) piston bowl

786

geometry

787

37

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788 789

Figure 5 Determination of autoignition of hydrogen-air mixture under dual fuel mode

790 791

Figure 6 Variation of ignition delay with respect to temperature with diesel-hydrogen dual fuel

792

mode at 100% load 38

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793

794 795

Figure 7 Autoignition of hydrogen-air fuel charge for base dual fuel mode (case I) [1]

39

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796 797

Figure 8 Validation of CFD results with experimental in-cylinder pressure and in-cylinder

798

temperature results [45] 40

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799 800

Figure 9 Contours of in-cylinder temperature for different hydrogen energy shares in the engine

801

under dual fuel mode at 100% load

41

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802 803

Figure 10 Diesel injection pressure and heat release rate curves under diesel-hydrogen dual fuel

804

mode at 50% load

805 806

Figure 11 Autoignition of hydrogen-air fuel charge with low temperature combustion strategies 42

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807 808

Figure 12 In-cylinder temperature variation at 19% hydrogen energy share for different test cases

809 810

Figure 13 Knock limited factor variation with respect to hydrogen energy share at different test

811

conditions 43

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812

Tables:

813

Table 1 Literature details of the maximum hydrogen energy share in hydrogen dual-fuel engines Reference

Engine details

Amount of hydrogen substitution

Saravanan et al. [18]

Nc=1, CR=16.5:1,

6.7% energy share

BMEP =5.4 bar Edwin et al. [14]

Nc=1, CR=17.5:1,

12.7% energy share

BMEP =5.3 bar Mathur et al. [19]

Nc=1, CR=17.5:1,

14.8% energy share (without power loss)

BMEP =4.9 bar de Morais et al. [20]

Nc=4, CR=17:1,

20% energy share

BMEP =6.5 bar Nguyen and Mikami

Nc=1, CR=16.7:1,

10% volume of intake air (or) 15% energy

[21]

BMEP =7.3 bar

share (approx.)

Bose et al. [22]

Nc=1, CR=17.5:1,

hydrogen flow rate of 0.15 kg/h (or)17.6%

BMEP =6.4 bar

energy share (approx.)

Nc=1, CR=17.5:1,

16.4% energy share

Yadav et al. [11]

BMEP =5.3 bar Christodoulou and

Nc=4, CR=18.2:1,

8% volume of intake air (or) 12.8% energy

Megaritis [23]

BMEP =9.2 bar

share (approx.)

Saravanan et al. [24]

Nc=1, CR=16.5:1,

10% energy share

BMEP =5.4 bar Shin et al. [25]

Nc=4, CR=17.3:1,

10% energy share

BMEP =4.9 bar

44

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Varde et al. [17]

Szwaja et al. [15]

Nc=1, CR=17.4:1,

14% hydrogen energy share at 5.8 bar BMEP

BMEP : 5.8 bar

and 17% share at 4.7 bar BMEP

Nc=2, CR=17:1,

17% hydrogen energy share (knock limited

BMEP : 11.2 bar

hydrogen share)

814 815 816

Table 2 Technical specifications of the engine Parameter

Description

Type of engine

Compression ignition engine

Number of cylinders

1

Displacement volume, cc

947.4

Rated power output, kW

7.4

Rated speed, rpm

1500

Compression ratio

19.5:1

Intake valve opening and closing, degree CA

43 before TDC & 67 after BDC

Exhaust valve opening and closing, degree CA

87 before BDC & 39 after TDC

Liquid fuel injection timing by spill, degree CA

8 before TDC

Nozzle opening pressure, bar

250

817 818

Table 3 Properties of the fuels used in the experimental study [3, 42] Fuel characteristics

Diesel

Hydrogen

Lower heating value, MJ/kg

44.05

120

Stoichiometric air fuel ratio

14.5

34.2 45

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Auto ignition temperature, K

530

858

Laminar burning velocity, m/s

0.3

2.65 -3.25

Cetane number

51

--

Density, kg/m3

821.5

0.083

Viscosity, cSt

2.64

--

819 820 821

Table 4 Details of measurement range and resolution of instruments/sensors Instrument/

Measuring

Measuring

Resolution

Accuracy

Uncertai

sensor Name parameter

range

nty (%)

Dyno-

Engine

0-150 N-m

0.1 N-m

0.2% of Full scale

2.42

controller

torque

Dyno-

Engine

0-10000 rpm

1 rpm

0.1% of Full Scale

0.195

controller

speed

Piezoelectric

In-cylinder

0-250 bar

Sensitivity:

± 0.3% to ± 0.6% of

0.846

pressure

pressure

45 pC/bar

value

(Peak pressure)

sensor Optical

Degree

720 pulses

0.1 degree

encoder

crank angle

per

CA

---

---

revolution Air flow

Intake air

meter

volumetric

0-330 m3/h

0.2 m3/h

± 2% of flow

1.7

flow rate

46

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Gas mass

Gaseous

flow meter

fuel flow

0-10 kg/h

0.01 kg/h

± 0.35% of flow

1.4

---

---

---

0.33

rate Calculated

Hydrogen

parameter

energy share

822 823 824

Table 5 Summary of models used in CFD simulation Description

Model used

Solver

Pressure based type (Transient)

Dynamic mesh

Layering (In-cylinder options)

Spatial discretization

Second order upwind

Turbulence

Standard k-Epsilon

Turbulence chemistry interaction (Combustion) Finite rate/Eddy dissipation Diesel pilot fuel injection (spray model)

Discrete phase mode (DPM); Solid cone

825 826 827

Table 6 Summary of autoignition temperatures for different experimental test cases [8] Operating condition

Fuels used

H2 energy

In-cylinder

Autoignition

share

pressure (bar)

temperature (K)

Dual fuel mode (Case I)

Diesel-H2

18.8

54.4

944.5

Water injection (Case II)

Diesel-H2

60.7

59.1

959.4 47

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Retarded diesel injection

Diesel-H2

21.2

56.1

951.2

Diesel-H2

24.5

57.2

954.6

timing [8] Retarded diesel injection timing [8] 828 829 830

48

ACCEPTED MANUSCRIPT Highlights  Maximum H2 energy share in a CI engine at 100% load is limited due to knocking  Autoignition of hydrogen-air charge leads to knocking during combustion  Increase in in-cylinder temperature is main reason for autoignition of the charge  Start of combustion advanced with H2 addition in the engine under dual fuel mode  Maximum H2 energy share increased with reduction in in-cylinder temperature