Investigation on HCCI Engine Operated with Dual Fuel System for Normal-heptane and Methanol

• Main Authors: Bancha Thanapiyawanit, 班超
• Other Authors: Jau-Huai Lu
• Format: Others
• Language: en_US
• Published: 2013
• Online Access: http://ndltd.ncl.edu.tw/handle/56806004490743407861

博士 === 國立中興大學 === 機械工程學系所 === 101 === HCCI (Homogenous charge compression ignition) engines have a potential to raise the efficiency of reciprocating engines during partial load operation. However, the performance of the HCCI engine at high loads is restricted by severe knocking. It is observed by the excessive pressure rise rate. This is due to the rapid combustion process occurring inside the cylinder, which does not follow the flame propagation that is seen in conventional engines. In this study, a low compression ratio of 9.5:1 gasoline engine was converted to operate in HCCI mode with the goal being to expand the stable operating region at high loads. Initially, pure n-heptane was used as the fuel and could be run steadily at equivalence ratios of 0.30 to 0.50 with elevated intake charge temperatures between 150oC and 90oC, respectively. The n-heptane HCCI engine could reach a highest performance at an IMEP (indicated mean effective pressure) of 0.38 MPa, which was greater than the performance found in the literature. To reach an even higher performance, a dual-fuel system was exploited. Iso-octane, methanol, and hydrous methanol as an anti-detonation additive, was introduced into the intake stream using an injection strategy in this study. A dual fuel system between n-heptane and iso-octane could expand operating load of the HCCI engine from equivalent ratio of 0.30 to 0.60 or IMEP of 0.38 to 0.42 MPa without compensation for thermal efficiency and emissions. Required intake charge temperature range could be reduced from 75˚C to 25˚C-width, which is helpful for control issue to regulate a desired temperature for a particular load operation. Another dual fuel system employed was n-heptane and methanol. The maximum IMEP is comparable with the previous case. Indicated thermal efficiency among the operating maintained at about 34% because combustion timing among operating range were identical intentionally for ease comparison. Introduction of 90% and 95% (vol/vol) hydrous methanol showed a similar trend but a lower thermal conversion efficiency and IMEP value. Therefore, to gain extra load by injection of secondary fuel could achieve and maintain high thermal conversion efficiency across a wide load. It enhances a 10.5% larger load compared to a pure n-heptane-fuelled HCCI engine. The hydrocarbon (HC) and carbon monoxide (CO) emissions were lower than 800 ppm and 0.10%, respectively for all the conditions tested in this study. They were less at high loads because of higher fuel concentration. The nitrogen oxides (NOx) emissions were below 12 ppm and were found to increase sharply at higher loads to a maximum of 23 ppm. In parallel, a single zone model to predict the temperature and the pressure histories in an HCCI engine is developed. Combustion phase was described by double-Wiebe function. The single zone model coupled with an double-Wiebe function were performed to simulate pressure and temperature between the period of IVC (Inlet valve close) and EVO (Exhaust valve open). A reduced kinetic detail mechanism of n-heptane and methanol was also used to construct the single zone combustion model. The phenomenon of two-stage combustion in an HCCI combustion mode was simulated. The n-heptane mechanisms presented by Tanaka and Donovan and methanol mechanisms obtained from GRI-Mech (www.me.berkeley.edu, 29 June 2013) were implemented in the model to evaluate their performance in comparison to experimental data. The initiation of the first stage of combustion and the time duration between the first and second stage of combustion were validated by adjusting the heat transfer coefficients. The modified model correctly predicted trends in combustion, including the required intake charge temperature and the onset of two-stage combustion. However, the peak combustion pressures were overestimated by approximately 11%. This overestimation was due to certain effects that were not considered in the model, including inhomogeneities in the mixture and leakage in the piston ring.