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Internal Combustion Engines

Exploring Strategies for Reducing High Intake Temperature Requirements and Allowing Optimal Operational Conditions in a Biogas Fueled HCCI Engine for Power Generation

[+] Author and Article Information
Iván D. Bedoya

Grupo de Ciencia y Tecnología del Gas y Uso Racional de la Energía, Department of Mechanical Engineering,  University of Antioquia, Calle 67 No. 53-108, Medellín, Colombiaibedoyac@udea.edu.co

Samveg Saxena

Combustion Analysis Laboratory,  University of California, Berkeley, California, CA 94720samveg@berkeley.edu

Francisco J. Cadavid

Grupo de Ciencia y Tecnología del Gas y Uso Racional de la Energía, Department of Mechanical Engineering,  University of Antioquia, Calle 67 No. 53-108, Medellín, Colombiafcadavid@udea.edu.co

Robert W. Dibble

Combustion Analysis Laboratory,  University of California, Berkeley, California, CA 94720rdibble@berkeley.edu

J. Eng. Gas Turbines Power 134(7), 072806 (May 23, 2012) (9 pages) doi:10.1115/1.4006075 History: Received November 18, 2011; Revised November 29, 2011; Published May 23, 2012; Online May 23, 2012

This paper evaluates strategies for reducing the intake temperature requirement for igniting biogas in homogeneous charge compression ignition (HCCI) engines. The HCCI combustion is a promising technology for stationary power generation using renewable fuels in combustion engines. Combustion of biogas in HCCI engines allows high thermal efficiency similar to diesel engines, with low net CO2 and low NOx emissions. However, in order to ensure the occurrence of autoignition in purely biogas fueled HCCI engines, a high inlet temperature is needed. This paper presents experimental and numerical results. First, the experimental analysis on a 4 cylinder, 1.9 L Volkswagen TDI diesel engine running with biogas in the HCCI mode shows high gross indicated mean effective pressure (close to 8 bar), high gross indicated efficiency (close to 45%) and NOx emissions below the 2010 US limit (0.27 g/kWh). Stable HCCI operation is experimentally demonstrated with a biogas composition of 60% CH4 and 40% CO2 on a volumetric basis, inlet pressures of 2–2.2 bar (absolute), and inlet temperatures of 200–210 °C for equivalence ratios between 0.19–0.29. At lower equivalence ratios, slight changes in the inlet pressure and temperature caused large changes in cycle-to-cycle variations, while at higher equivalence ratios these same small pressure and temperature variations caused large changes to the ringing intensity. Second, numerical simulations have been carried out to evaluate the effectiveness of high boost pressures and high compression ratios for reducing the inlet temperature requirements while attaining safe operation and high power output. The one zone model in Chemkin was used to evaluate the ignition timing and peak cylinder pressures with variations in temperatures at intake valve close (IVC) from 373 to 473 K. In-cylinder temperature profiles between IVC and ignition were computed using Fluent 6.3 and fed into the multizone model in Chemkin to study combustion parameters. According to the numerical results, the use of both higher boost pressures and higher compression ratios permit lower inlet temperatures within the safe limits experimentally observed and allow higher power output. However, the range of inlet temperatures allowing safe and efficient operation using these strategies is very narrow, and precise inlet temperature control is needed to ensure the best results.

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Copyright © 2012 by American Society of Mechanical Engineers
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References

Figures

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Figure 1

Gross indicated mean effective pressure (IMEPg ) for different combustion phasing (CA50), intake pressures, and intake temperatures

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Figure 2

Gross indicated efficiency for different combustion phasing (CA50), intake pressures, and intake temperatures

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Figure 3

Normalized H2 O2 molar fraction, and normalized cumulative heat release (CHR) for 0.29 PHI and 0.19 PHI. PIVC  = 2 bar, TIVC for 0.29 PHI  = 473 K, and TIVC for 0.19 PHI  = 463 K.

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Figure 4

Simulated and experimental motored pressure traces comparison. PIVC  = 2 bar, TIVC  = 473 K, and PHI  = 0.29.

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Figure 5

Temperature distribution in 20 zones at 3 CAD bTDC. PIVC  = 2 bar, TIVC  = 473 K, and PHI  = 0.29.

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Figure 6

Averaged temperature and mass fraction distribution in 20 zones at 3 CAD bTDC. PIVC  = 2 bar, TIVC  = 473 K, and PHI = 0.29.

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Figure 7

Temperature profile for 20 zones related with the crank angle. PIVC  = 2 bar, TIVC  = 473 K, and PHI  = 0.29.

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Figure 8

Comparison between simulated pressure traces using 40 zones and experimental pressure traces for three different equivalence ratios. PIVC  = 2 bar, TIVC for 0.29 PHI  = 473 K, TIVC for 0.21 PHI  = 467 K, and TIVC for 0.19 PHI  = 463 K.

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Figure 9

Simulated-over-experimental results ratio related with the equivalence ratio. The experimental values are extracted from Table 4. The simulated results are achieved with the 40-zone model. The HC emissions and CO emissions are located at EVO.

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Figure 10

(a) Ignition timing predicted with the 1-zone model related with the TIVC for different PIVC and different CR. (b) In-cylinder peak pressures (Pmax ) predicted with the 1-zone model and related with the ignition timing for different PIVC and different CR. PHI  = 0.29.

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Figure 11

(a) Ignition timing predicted with the 1-zone model related with the TIVC for different PIVC . (b) In-cylinder peak pressures (Pmax ) predicted with the 1-zone model and related with the ignition timing for different PIVC . CR = 16.86, and PHI = 0.29.

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Figure 12

Simulated pressure traces achieved with the 40-zone model for Cases 2 and 3, allowing lower temperatures at IVC and for the experimental base case (1). PHI  = 0.29.

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Figure 13

In-cylinder zone average temperature achieved with the 40-zone model for Cases 2 and 3, allowing lower temperatures at IVC and for the experimental base case (1). PHI  = 0.29.

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Figure 14

Comparison of combustion parameters and emissions for Cases 2 and 3 achieved with the 40-zone model. The results have been normalized by the results for Case 1. PHI  = 0.29.

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