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Research Papers: Gas Turbines: Combustion, Fuels, and Emissions

Spark Advance Modeling of Hydrogen-Fueled Spark Ignition Engines Using Combustion Descriptors

[+] Author and Article Information
Saket Verma

Engines and Unconventional Fuels Lab,
Centre for Energy Studies,
Indian Institute of Technology Delhi,
Block-V, Huaz Khas,
New Delhi 110016, India
e-mails: saketverma@hotmail.com;
ssaketverma@gmail.com

L. M. Das

Centre for Energy Studies,
Indian Institute of Technology Delhi,
Huaz Khas,
New Delhi 110016, India

Contributed by the Combustion and Fuels Committee of ASME for publication in the JOURNAL OF ENGINEERING FOR GAS TURBINES AND POWER. Manuscript received January 7, 2016; final manuscript received November 13, 2017; published online April 13, 2018. Assoc. Editor: Jeffrey Naber.

J. Eng. Gas Turbines Power 140(8), 081501 (Apr 13, 2018) (11 pages) Paper No: GTP-16-1007; doi: 10.1115/1.4038798 History: Received January 07, 2016; Revised November 13, 2017

In-cylinder pressure-based combustion descriptors have been widely used for engine combustion control and spark advance scheduling. Although these combustion descriptors have been extensively studied for gasoline-fueled spark ignition (SI) engines, adequate literature is not available on use of alternative fuels in SI engines. In an attempt to partially address this gap, present work focuses on spark advance modeling of hydrogen-fueled SI engines based on combustion descriptors. In this study, two such combustion descriptors, namely, position of the pressure peak (PPP) and 50% mass fraction burned (MFB) have been used to evaluate the efficiency of the combustion. With a view to achieve this objective, numerical simulation of engine processes was carried out in computational fluid dynamics (CFD) software ANSYS fluent and simulation data were subsequently validated with the experimental results. In view of typical combustion characteristics of hydrogen fuel, spark advance plays a very crucial role in the system development. Based on these numerical simulation results, it was observed that the empirical rules used for combustion descriptors (PPP and 50% MFB) for the best spark advance in conventional gasoline fueled engines do not hold good for hydrogen engines. This work suggests revised empirical rules as: PPP is 8–9 deg after piston top dead center (ATDC) and position of 50% MFB is 0–1 deg ATDC for the maximum brake torque (MBT) conditions. This range may vary slightly with engine design but remains almost constant for a particular engine configuration. Furthermore, using these empirical rules, spark advance timings for the engine are presented for its working range.

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Figures

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Fig. 1

Cylinder pressure versus crank angle for various ignition timings showing pressure peaks and combustion descriptor of PPPs (equivalence ratio = 0.48, speed = 2000 rpm, and compression ratio = 9.5)

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Fig. 2

Mass fraction burned versus crank angle for various ignition timings showing combustion descriptor of 50% MFB (equivalence ratio = 0.48, speed = 2000 rpm, and compression ratio = 9.5)

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Fig. 3

Variation of engine torque, PPP and 50% mass fraction burned (50% MFB) with ignition timing (equivalence ratio = 0.48, speed = 2000 rpm, and compression ratio = 9.5)

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Fig. 4

Sensitivity analysis of grid discretization on the simulation results

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Fig. 5

Three-dimensional moving mesh

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Fig. 6

Comparison of experimental and simulation data for cylinder pressure versus crank angle (speed = 2000 rpm, equivalence ratio = 0.48, spark timing = 12 deg BTDC)

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Fig. 7

Comparison of experimental and simulation data for cylinder pressure versus crank angle (speed = 2000 rpm, equivalence ratio = 0.48, spark timing = 16 deg BTDC)

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Fig. 8

Comparison of experimental and simulation data for cylinder pressure versus crank angle (speed = 2200 rpm, equivalence ratio = 0.6, spark timing = 10 deg BTDC)

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Fig. 9

Comparison of experimental and simulation data for cylinder pressure versus crank angle (speed = 2200 rpm, equivalence ratio = 0.6, spark timing = 14 deg BTDC)

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Fig. 10

(a) Cylinder pressure versus crank angle and (b) progress variable versus crank angle; at speed = 2000 rpm, equivalence ratio = 0.6

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Fig. 11

(a) Cylinder pressure versus crank angle and (b) progress variable versus crank angle; at speed = 2000 rpm, equivalence ratio = 0.8

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Fig. 12

(a) Cylinder pressure versus crank angle and (b) progress variable versus crank angle; at speed = 2000 rpm, equivalence ratio = 1.0

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Fig. 13

(a) Cylinder pressure versus crank angle and (b) progress variable versus crank angle; at speed = 2000 rpm, equivalence ratio = 1.2

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Fig. 15

Maximum brake torque timings for various equivalence ratios and engine speeds

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Fig. 14

Contours of progress variable with crank angle: (a) CA = −131 deg, (b) CA = −18 deg, (c) CA = 60 deg, and (d) CA = 150 deg

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