Effects of fuel combination and IVO timing on combustion and emissions of a dual-fuel HCCI combustion engine

Abstract

This paper experimentally and numerically studied the effects of fuel combination and intake valve opening (IVO) timing on combustion and emissions of an n-heptane and gasoline dual-fuel homogeneous charge compression ignition (HCCI) engine. By changing the gasoline fraction (GF) from 0.1 to 0.5 and the IVO timing from –15°CA ATDC to 35°CA ATDC, the in-cylinder pressure traces, heat release behaviors, and HC and CO emissions were investigated. The results showed that both the increased GF and the retarded IVO timing delay the combustion phasing, lengthen the combustion duration, and decrease the peak heat release rate and the maximum average combustion temperature, whereas the IVO timing has a more obvious influence on combustion than GF. HC and CO emissions are decreased with reduced GF, advanced IVO timing and increased operational load.

Keywords

homogeneous charge compression ignition / dual-fuel / n-heptane / gasoline / intake valve opening timing

Introduction

In conventional internal combustion (IC) engines, diesel engines have a high thermal efficiency but high emissions of nitric oxides (NO_x) and particulate matter (PM). In comparison, gasoline engines produce low NO_x and PM emissions but have a low thermal efficiency [¹]. With the requirement of thermal efficiency improvement and emissions reduction for IC engines, various advanced combustion strategies have been studied [²–⁴]. Homogeneous charge compression ignition (HCCI) combines the advantages of diesel engines and gasoline engines, and can simultaneously achieve a high thermal efficiency and low NO_x and PM emissions under certain operation conditions [⁵–⁷]. However, it is difficult to control the HCCI combustion process due to its kinetics-controlled auto-ignition characteristics.

In HCCI combustion, fuel chemical properties are the major factors determining the auto-ignition characteristics [⁸,⁹]. To accurately control the fuel auto-ignition in HCCI engines and avoid the abnormal combustion phenomena, an effective method of flexibly adjusting fuels with different auto-ignition tendencies according to the engine conditions, also named as dual-fuel technology is often adopted [¹⁰–¹⁴]. Using a gasoline/diesel dual-fuel engine, Lu et al. [¹⁵] investigated the effects of fuel fraction on the combustion and emissions characteristics. The results indicated that the maximum in-cylinder pressure and maximum heat release rate are decreased and the NO_x emissions are increased with elevated gasoline fraction. Huang et al. [¹⁶] investigated the combustion and emissions characteristics of alcohol/n-heptane dual-fuel HCCI engine. It is found that the combustion phasing is sensitive to fuel heat value and premixed ratio. Besides, NO_x and PM emissions can be reduced with alcohol addition. Ma et al. [¹⁷] studied the effects of dual-fuel injection strategies of gasoline/diesel dual fuel and found that a higher gasoline ratio could achieve lower PM emissions. Kuzuoka et al. [¹⁸] studied the dual-fuel combustion characteristics with diesel direct injection and gasoline port injection, and demonstrated that by varying gasoline fractions to an optimal fraction, low NO_x emissions and high thermal efficiency could be simultaneously achieved.

In addition to the dual-fuel strategy, air inlet control also has great influences on HCCI combustion [¹⁹,²⁰]. Cinar et al. [²¹] studied the effects of intake air temperature on the emissions of a dual-fuel HCCI engine. They indicated that NO_x emissions increased with intake air temperature, while carbon monoxide (CO) and unburned hydrocarbon (UHC) emissions first increased and then decreased after the intake air temperature exceeded a critical value. Yeom et al. [²²] studied the effects of intake valve timing on combustion and emissions of a liquefied petroleum gas (LPG) and gasoline dual-fuel HCCI engine. They pointed out that the indicated mean effective pressure (IMEP) was reduced when IVO was outside an optimal range, due to the negative work and incomplete combustion. Vos et al. [²³] experimentally and numerically studied the effects of intake valve close (IVC) timing on charge efficiency and emissions on a diesel engine. The results showed that delaying IVC enhanced the volumetric efficiency, allowing a higher exhaust gas recirculation (EGR) for NO_x emissions reduction.

Although some studies have been conducted to identify the individual effect of fuel combination strategies, IVO timings or injection timings on HCCI combustion, the synergistic effects of fuel combination, and IVO timing on dual-fuel HCCI combustion and emissions have not yet been investigated. Based on the HCCI engine test bench proposed in Ref. [²⁴], the effects of gasoline/n-heptane dual-fuel combination and IVO timing on the combustion and emissions of the HCCI engine was studied. The in-cylinder pressure and temperature, heat release behaviors, and HC and CO emissions were characterized at different operational loads, fuel combinations, and IVO timings.

Methodology

Experimental setup

The engine test was conducted on an n-heptane/gasoline dual-fuel HCCI engine modified from a four-cylinder, naturally aspirated 1.5 L gasoline engine. The specification of the test engine is listed in Table 1. To realize dual-fuel HCCI combustion, two injectors for n-heptane and gasoline, respectively, were installed in the intake manifold. Both fuels were injected at the same time after the intake valves were closed in the previous cycle. The fuel injection mass was adjusted by changing the width of energizing pulse, and the total injection mass was controlled with an energy input of 296.8 J/cycle to maintain the engine load. The gasoline fraction is defined as the fraction of heat value provided by gasoline in Eq. (1).

(1)\mathrm{G}\mathrm{F}={\displaystyle \frac{{{\displaystyle m}}_{\mathrm{g}}\times {{\displaystyle \mathrm{L}\mathrm{H}\mathrm{V}}}_{\mathrm{g}}}{{{\displaystyle m}}_{\mathrm{g}}\times {{\displaystyle \mathrm{L}\mathrm{H}\mathrm{V}}}_{\mathrm{g}}+{{\displaystyle m}}_{\mathrm{h}}\times {{\displaystyle \mathrm{L}\mathrm{H}\mathrm{V}}}_{\mathrm{h}}}},

where m_g and LHV_g are the mass and the lower heating value of gasoline, respectively, and m_h and LHV_h are the mass and the lower heating value of n-heptane, respectively. The IVO timing can be progressively adjusted using a variable valve timing (VVT) system. In addition, the intake cam profile was modified to increase the adjustment range of IVO timing and extend the air intake duration. The modified intake and exhaust valve lift profiles are shown in Fig. 1.

The experiment was performed with different fuel combinations and IVO timings, keeping the intake air temperature, oil temperature and coolant temperature constant. Considering that the inlet temperature was set to 15°C to ensure adequate charge coefficient at naturally aspirated conditions, n-heptane instead of diesel fuel was selected due to its higher vaporization rates. Besides, a certain inlet turbulence and inlet valve heating was necessary and realized in this experiment. More information about the experimental condition is provided in Table 2. The in-cylinder pressure traces shown in this paper were collected and averaged from 50 consecutive cycles, and the in-cylinder heat release rate was calculated based on a zero-dimensional heat release model.

Computational model

The simulation was conducted using a three-dimensional CFD software CONVERGE [²⁵]. The computational models are presented in Table 3. The turbulent flow was estimated using the RNG k-ε turbulence model [²⁶] and combustion was simulated based on a multi-dimension model [²⁷]. Primary reference fuel 93 (PRF93) was selected as the gasoline model fuel, due to the same research octane number (RON= 93) as the test gasoline. A skeletal PRF mechanism with 41 species and 124 reactions developed by Liu et al. [²⁸] was used in the simulation. In addition, the heat transfer on cylinder wall was estimated using the heat transfer model proposed by Han and Reitz [²⁹].

An automatically generated orthogonal hexahedron mesh was used and the base grid was 4 mm. The grid generation at the bottom dead center is illustrated in Fig. 2. An adaptive mesh refinement (AMR) and a fixed embedding mesh were implemented to satisfy the actual computational requirements. The AMR method that adaptively refines the grids where the velocity and temperature gradients are large, was used at the intake valve bottom and the computational domain boundaries. The fixed embedding method that refines the grids at designated areas was set for 3 levels, meaning the minimum grid size was 0.5 mm. Figure 3 demonstrates the comparison between the measured and simulated in-cylinder pressures and heat release rates (HRRs) under different operational conditions, i.e., GFs and IVO timings, and a satisfactory agreement is observed.

Results and discussion

Effects of fuel combination on dual-fuel HCCI combustion and emissions

Figure 4 exhibits the experimental results of the effects of GFs on in-cylinder pressure and HRR at the loads of 50 N·m and 40 N·m. As GF is increased, the combustion phasing is delayed, the combustion duration is lengthened, and the peak heat release is decreased. The reason for this is that when GF is increased, the low-temperature heat release of n-heptane is reduced, which further delays the high temperature combustion phasing, and lowers the pressure rise rate and the maximum pressure.

Figure 5 plots the effects of GFs on CA10 and CA50. The crank angles correspond to the instants when 10% and 50% of the total cumulative heat release are achieved. With increased GF, CA10 and CA50 are both delayed at the given operational loads. This is because the reduced n-heptane fraction causes less low-temperature heat release, and thus reduces fuel auto-ignition propensity and extends combustion duration. Further, with increased engine load, CA50 is advanced, as increased fuel equivalence ratio at higher loads elevates combustion rate and as such advances the combustion phasing. In contrast, CA10 is slightly affected by engine loads. This is because CA10 is closely related to fuel low-temperature heat release characteristics, and the low-temperature ignition behaviors of homogeneous fuel mixtures has been proved to be slightly affected by equivalence ratios [³⁰].

Effects of GFs on in-cylinder temperature distribution contours, in low-temperature and high-temperature combustion stages, are shown in Figs. 6 and 7, respectively. Figure 6 shows the in-cylinder temperature distribution in the low-temperature heat release stage (700 K–900 K) at IVO= –5°CA ATDC, n = 1600 r/min, and T = 50 N∙m. At the same crank angle, the temperature distribution characteristics of GF= 0.3 and GF= 0.39 are almost the same, indicating that both the low-temperature combustion phasing and the HRR are similar at different GFs. This is consistent with the experimental observation for the low-temperature heat release behaviors shown in Fig. 4. Figure 7 exhibits the in-cylinder temperature distribution in the high-temperature heat release stage (900 K–1800 K) at IVO= –5°CA ATDC, n = 1600 r/min, and T = 50 N∙m. Obviously, the heat release phasing in the high-temperature stage is delayed as GF increases from 0.3 to 0.39. Besides, the temperature contours show that the maximum average in-cylinder temperature drops from 1915 K to 1855 K as GF varies from 0.3 to 0.39, revealing that the high-temperature heat release is inhibited.

Figure 8 shows the effects of GFs on brake specific fuel consumption (BSFC). With increased GF, BSFC is first decreased and then increased. The turning GF point that corresponds to the minimum BSFC is dependent on engine load. Comparing Fig. 8 with Fig. 5, under optimal BSFC conditions, the corresponding CA50 is around 4–6°CA ATDC. The GFs corresponding to the minimum BSFCs, at 30 N∙m, 40 N∙m, and 50 N∙m operational loads, are 0.1, 0.19, and 0.39, respectively. This indicates an increase of the optimum GF with elevated engine load. This is because the increased engine load promotes the combustion rate and advances combustion phasing, and GF has to be increased to maintain the key combustion phasing at the optimal range for the minimum BSFCs.

The temperature distribution is uniform in HCCI engines, and the maximum in-cylinder temperature is around 1900 K, at which NO_x formation is not favored [³¹]. As such, the engine experimental results reveal that the NO_x emissions are below 5 × 10⁻⁶ across the range of the entire test conditions [²⁴]. In addition, the simulation results also confirm the ultralow NO_x emissions under studied test conditions. However, HC and CO are the primary emission species in HCCI engines, the concentrations of which are shown in Fig. 9. With increased GF or reduced engine load, both HC and CO emissions increase. Combustion temperature is a primary factor affecting HC and CO formation in HCCI combustion. As increased GF and reduced load decrease heat release and in-cylinder temperatures, HC oxidation [³²] and the conversion from CO to CO₂ [³³] are inhibited, elevating HC and CO emissions. Taking CO as a representative incomplete combustion product, the in-cylinder CO evolvement processes at different GFs are illustrated in Fig. 10. CO is initially formed near the wall and in the center of the combustion chamber, where a higher in-cylinder temperature is generated, as shown in Fig. 7. When GF is increased from 0.3 to 0.39, the in-cylinder temperature rises more slowly, and thus higher CO emissions are produced.

Effects of IVO timings on dual-fuel HCCI combustion and emissions

Figure 11 shows the effects of IVO timings on in-cylinder pressure and HRR at different GFs. At a given GF, retarded IVO timing lowers down the peak in-cylinder pressure and the peak HRR, and the maximum pressure rise rate is reduced. The crank angles corresponding to the peak in-cylinder pressure and the peak HRR are also delayed. As shown in Fig. 11, as IVO timing is delayed, the in-cylinder pressure prior to ignition decreases noticeably. Since the HCCI combustion is highly dependent on the in-cylinder thermodynamic state prior to ignition, the combustion phasing of the low-temperature heat release stage, as well as the overall combustion phasing is thus delayed.

Figure 12 shows the effects of IVO timings on CA10 and CA50. Under the given condition (n = 1600 r/min, and T = 50 N∙m), CA10 and CA50 are both delayed with the retarded IVO timing because of the decreased charge pressure and temperature prior to ignition, which reduce the auto-ignition tendency and extend the overall combustion duration. Therefore, CA10 and CA50 are both postponed. In addition, at high GFs, the effects of IVO timings on CA10 are less evident. This is because CA10 is influenced by both the charge condition and fuel auto-ignition tendency. The ignition delay is greatly extended when GF rises, which makes the charge condition a secondary influencing factor.

The effects of IVO timings on temperature distribution contours are shown in Figs. 13 and 14. Figure 13 shows the in-cylinder temperature distribution in the low-temperature heat release stage (700 K–900 K) at GF= 0.3, n = 1600 r/min, and T = 50 N∙m. When the IVO timing varies from –5°CA ATDC to 15°CA ATDC, the combustion phasing of the low-temperature stage is slightly delayed, which is consistent with the experimental observation for the low-temperature heat release as shown in Fig. 11. Figure 14 shows the in-cylinder temperature distribution in the high-temperature heat release stage (900 K–1800 K), at GF= 0.3, n = 1600 r/min, and T = 50 N∙m. When the IVO timing varies from –5°CA ATDC to 15°CA ATDC, the heat release phasing is obviously delayed by over 3°CA ATDC and the maximum average in-cylinder temperature drops from 1852 K to 1826 K, revealing a reduced heat release as similarly observed in Fig. 11.

Figure 15 shows the effects of IVO timings on HC and CO emissions. With the retarded IVO timing, both HC and CO emissions increase. The retarded IVO timing have the same effects as increased GF. This is because the delayed IVO timing leads to lower in-cylinder combustion temperatures, which enhances incomplete fuel combustion and inhibits the CO to CO₂ oxidation reactions. Moreover, at an increased GF, CO emissions rise more rapidly as the IVO timing is delayed. This is because both increased GF and retarded IVO timing could decrease the combustion temperature, and thus the combined impact of high GF and late IVO timing makes the CO emissions increase more rapidly. Figure 16 demonstrates the CO evolvement process in the combustion chamber at different IVO timings. When the IVO timing is delayed from –5°CA ATDC to 15°CA ATDC, CO formation and consumption are delayed, and thus higher engine-out CO emissions may be produced due to the lower combustion temperatures.

Conclusions

An experimental and numerical study of an n-heptane/gasoline dual-fuel HCCI engine was conducted to investigate the effects of gasoline fraction and IVO timing on engine combustion and emissions. The in-cylinder pressure and temperature, heat release behaviors, and HC and CO emissions were analyzed. Based on the findings, the following conclusion can be reached.

Gasoline fraction mainly affects the high-temperature stage combustion. With increased gasoline fraction, the combustion phasing is delayed, the combustion duration is lengthened, and the maximum temperature is decreased. When the engine load increases, the gasoline fraction should be increased to achieve the optimal fuel consumption.

Delaying the IVO timing retards the overall combustion phasing, lengthens the combustion duration, and decreases the HRR and maximum temperature. At high gasoline fractions, the IVO timing has less influences on the combustion phasing.

Reduced gasoline fraction or advanced IVO timing decreases HC and CO emissions from the dual-fuel HCCI engine, because the in-cylinder temperature is increased with reduced gasoline fraction and advanced IVO timing, and as such the HC oxidation and CO conversion to CO₂ are promoted.

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