Active fuel design—A way to manage the right fuel for HCCI engines

Abstract

Homogenous charge compression ignition (HCCI) engines feature high thermal efficiency and ultralow emissions compared to gasoline engines. However, unlike SI engines, HCCI combustion does not have a direct way to trigger the in-cylinder combustion. Therefore, gasoline HCCI combustion is facing challenges in the control of ignition and, combustion, and operational range extension. In this paper, an active fuel design concept was proposed to explore a potential pathway to optimize the HCCI engine combustion and broaden its operational range. The active fuel design concept was realized by real time control of dual-fuel (gasoline and n-heptane) port injection, with exhaust gas recirculation (EGR) rate and intake temperature adjusted. It was found that the cylinder-to-cylinder variation in HCCI combustion could be effectively reduced by the optimization in fuel injection proportion, and that the rapid transition process from SI to HCCI could be realized. The active fuel design technology could significantly increase the adaptability of HCCI combustion to increased EGR rate and reduced intake temperature. Active fuel design was shown to broaden the operational HCCI load to 9.3 bar indicated mean effective pressure (IMEP). HCCI operation was used by up to 70% of the SI mode load while reducing fuel consumption and nitrogen oxides emissions. Therefore, the active fuel design technology could manage the right fuel for clean engine combustion, and provide a potential pathway for engine fuel diversification and future engine concept.

Graphical abstract

Keywords

active fuel design / HCCI / gasoline / n-heptane / engine / combustion

1 Introduction

Homogenous charge compression ignition (HCCI) combustion, an advanced engine combustion strategy initially developed three decades ago [1], has been widely investigated due to its improved thermal efficiency as well as ultralow emissions compared to gasoline spark ignition (SI) engines. Major strategies to achieve gasoline HCCI combustion include intake air heating [2], increased compression ratio [3] and variable valve train control [4]. However, unlike SI engines, HCCI combustion does not have a direct way, i.e., spark ignition, to trigger the in-cylinder combustion; instead, fuel oxidation reaction kinetics are key factors determining the ignition timing, combustion process and emissions formation [5]. Therefore, at high load conditions, HCCI combustion tends to release a large amount of heat in a short period, causing rapid an increased pressure rise rate and the occurrence of combustion knocking [6,7]. On the other hand, at low load conditions, the fuel-lean charge in HCCI combustion may have increased combustion instability problems, which leads to increased cycle variance, reduced thermal efficiency and increased incomplete combustion products [8,9]. All these problems limit the real application of gasoline HCCI combustion and need different technology breakthroughs in the combustion control and operational range extension.

Intake boost has been considered as an effective strategy to broaden the load range of gasoline HCCI engines since a decade ago [10,11]. Kuboyama et al. [12] utilized the blow down supercharge (BDSC) system to elevate the in-cylinder temperature and fuel concentration at low load conditions and create diluted mixture at high load conditions. Meanwhile, a secondary air injection system was employed to inject air into an exhaust port of each cylinder to control recharged EGR gas temperature and minimize the ignition timing variance in each cylinder. In this way, the operational range of HCCI combustion was extended at both low load and high load ends. The intake pressure was further increased from 1 bar to 2 bar using an effective external supercharger, and the maximum indicated mean pressure (IMEP) of this BDSC-HCCI engine could be comparable to that of the naturally aspirated SI engine. Meanwhile, the pressure rise rate is well controlled due to the in-cylinder temperature stratification produced by the BDSC system [13].

Urushihara et al. [14] strived to extend the low load limit of a gasoline HCCI engine by fuel reformulation. A split injection strategy, including an injection event during the negative valve overlap (NVO) period for fuel reforming, and the other injection event during the intake stroke for homogenous charge formation, was proposed. The fuel injected into the high temperature exhaust gas could produce active chemical radicals, beneficial for HCCI combustion at the low load conditions. Aiming to maintain the stable HCCI operation, the fuel injection quantity during the NVO period should be increased as engine load decreased. Sjöberg et al. [15,16] investigated the effects of in-cylinder temperature stratification, realized by coolant temperature reduction and increasing in-cylinder swirl flow, on HCCI combustion. It was found that increased temperature stratification could reduce the maximum heat release rate and extend combustion duration, thus restricting the potential of knocking occurrence but increasing the heat transfer loss. Liu et al. [17] investigated the effects of in-cylinder temperature and charge stratification on HCCI combustion. It was indicated that reasonably adjusted in-cylinder swirl flow, injection strategy, and coolant temperature could influence the stratification extent, which could effectively reduce the pressure rise rate and broaden the HCCI operational range.

In spite of the limited operational range of HCCI combustion, however, HCCI could cover most common vehicle drivable conditions. In particular, at intermediate-speed vehicle cruise conditions, HCCI mode could reduce fuel consumption rate by 20% and NO_x emissions by 90% [18,19]. Therefore, utilizing an HCCI-SI dual-mode strategy, in which the HCCI mode is realized at low to intermediate load conditions and the SI mode is applied to the high load conditions, might be a practical way to achieve the full load engine operation. In the HCCI-SI strategy, a rapid and smooth transition process between operation modes is a key factor. Hyvönen et al. [20] realized the transition process between the SI and HCCI combustion on a variable compression ratio (VCR) engine with a rapid thermal management system, but misfire was frequently observed during this process. Tian et al. [21] also studied the transition process from SI to HCCI modes, during which the valve timing and injection strategy were first changed and the throttle valve position was then adjusted. The engine after transition could be fired by compression ignition without any misfire or knocking problems. Similar SI-HCCI transition strategies were employed in general motors (GM) prototype vehicles in 2007 to cover the full operational conditions [22].

Because of the kinetics-dominant features in HCCI combustion, the fuel chemical properties play important roles in the HCCI ignition and combustion process. High cetane fuels have more compression ignition tendency, and might extend HCCI operation to fuel-lean conditions. On the other hand, at fuel-rich and high-temperature in-cylinder conditions, low cetane fuels are more suitable for the in-cylinder pressure rise rate control and may elevate the upper load limit of HCCI operation. Therefore, researchers strived to take advantage of the fuel chemical properties in HCCI engine control and optimization. Lü et al. [23–24], Hou et al. [25] and Huang et al. [26] proposed a fuel design concept for HCCI combustion, in which fuel components and physiochemical properties could be optimized according to engine operation conditions. First, two primary reference fuels (n-heptane and iso-octane) and their blends were studied on an HCCI engine. By changing the proportion of different primary reference fuels, the timing and amount of low temperature heat release in HCCI combustion was adjustable, and the system reactivity, including the duration of the negative temperature coefficient region, high temperature ignition timing, could be flexibly controlled [23,24]. In addition to simply blending mixing basic fuels, the fuel design concept was further realized in a more practical way. By adopting two independent fuel supply systems, the real-time control of dual-fuel injection was achieved to control ignition timing and combustion phasing of HCCI [25,26]. Wilhelmsson et al. [27] investigated an n-heptane/natural gas dual-fuel HCCI engine equipped with a variable geometry turbo charger. Intake pressure and combustion phasing were adopted as the major operational parameters to optimize emissions, combustion noise and efficiency. Yeom and Bae [28] evaluated effects of LPG-DME dual fuel strategy on HCCI combustion. Liquefied petroleum gas (LPG) was injected into the intake port while dimethyl ether (DME) was used as an ignition promoter by directly injecting into the cylinder during the intake stroke. Besides, the combustion, knock characteristics and exhaust emissions were investigated. Aldawood et al. [29] tried to control the combustion heat release rate by adjusting fuel reactivity, and the HCCI engine operation window could be expanded. Shibata and Ogawa [30] applied DME-ethanol binary fuel to a single cylinder HCCI engine. Considering the promotion effect of DME and inhibition effect of ethanol on low temperature heat release, the HCCI combustion phasing could be optimized by adjusting fuel proportions, and thus the engine load-speed range could be broadened. Meanwhile, EGR was utilized to extend the high temperature heat release period; therefore, the cylinder pressure rise rate could be controlled.

In this paper, an active fuel design concept was proposed to a four-cylinder HCCI engine, to explore a potential pathway to realize real-time HCCI combustion control, robust SI-HCCI operation transition, and extended operational range. The active fuel design concept was realized by real time dual-fuel (gasoline and n-heptane) injection control using two independent injection systems, with EGR rate and intake temperature adjusted.

2 Experimental apparatus and method

The engine used in the experiments is an SI, port injected, naturally aspirated in-line four-cylinder engine with 1.5 L displacement/cylinder whose specifications are presented in Table 1. In order to realize a stable operation and the transition between SI and HCCI modes, an intake heating system, an intake boost system, and an exhaust gas recirculation (EGR) system were installed. The experimental apparatus is illustrated in Fig. 1. The two fuels used in the experiment were high-octane gasoline and high-cetane n-heptane whose properties are listed in Table 2. Two identical fuel injection systems were employed for gasoline and n-heptane injection, respectively. The fuel injection pressure was 0.3 MPa. To ensure a steady HCCI combustion and study the influence of intake air temperature on the dual-fuel HCCI, the compression ratio was increased to 13. An intake air heat exchanger was installed and an insulating coating of zirconia was placed on the intake manifold. Gasoline and n-heptane were both injected into the intake port after intake valve close (IVC), but n-heptane was injected 10^oCA later than gasoline injection to avoid fuel spray collision.

To study the influence of fuel proportion on HCCI combustion, the gasoline ratio (RG) in the cyclic fuel injection quantity is defined as

\mathrm{RG}={\displaystyle \frac{{m_{\mathrm{g}}}_{}\times {\mathrm{L}\mathrm{H}\mathrm{V}}_{\mathrm{g}}{}_{}}{m_{\mathrm{g}}\times {\mathrm{L}\mathrm{H}\mathrm{V}}_{\mathrm{g}}+m_{\mathrm{n}}\times {\mathrm{L}\mathrm{H}\mathrm{V}}_{\mathrm{n}}{}_{}}},

in which m_g is the cyclic gasoline injection quantity, LHV_g is the lower heat value of gasoline, m_n is the cyclic n-heptane injection quantity, and LHV_n is the lower heat value of n-heptane.

The engine speed and torque were measured using an electric dynamometer. The in-cylinder pressure was measured using a Kistler 6118B spark-plug type cylinder pressure sensor. The cylinder pressure data of 50 consecutive cycles of the four cylinders were collected and averaged for the heat release calculation. The heat release rate, bulk gas temperature and IMEP in this study were calculated from a zero-dimensional model while a crank angle based differential form of the first law was derived by applying the energy conservation equation. In this differential equation, only the in-cylinder internal energy, cycle work and heat loss between combustion gases and cylinder wall were taken into account, and the heat and mass transfer as a result of blow-by and crevice flow was not considered. The heat loss coefficient was obtained using empirical correlations developed by Woschni [31]. The gases during the compression, combustion and expansion processes were assumed to possess the ideal-gas characteristics, so that the instantaneous bulk gas temperature can be calculated using the ideal gas state equation, based on the measured cylinder pressure and calculated cylinder volume. At each crank angle, the in-cylinder gases species were estimated by assuming complete combustion of the burned gases. The CO, HC and NO_x emissions were measured using an AVL gas analyzer.

3 Results and discussion

3.1 Spark ignition to HCCI transition

It has been proven that it is difficult to start the engine in HCCI mode. Therefore, the engine is usually started in SI mode and then transited to HCCI mode. With the dual-fuel injection system, the transition from SI to HCCI could be realized by fuel proportion adjustment. During the transition process, the intake temperature and pressure are first acquired by an electronic control unit and compared with the pre-stored data. Then the ratio of gasoline and n-heptane, as well as the cyclic injection quantity are calculated according to engine load condition. Afterwards, the throttle valve was completely opened, and the injection pulse widths of gasoline and n-heptane were adjusted to realize the fuel injection quantity determined previously. The switching process targets the same output torque and engine speeds. The end of transition period is judged by the fact whether spark ignition is required or not. During the transition process, there is one misfire cycle. However, the output torque before and after the switching is kept unchanged, and the engine speed has a variation of 40 r/min, which disappears after 50 engine cycles. Figure 2 presents the SI to HCCI transition process at a condition of 35 N∙m output torque and 1200 r/min engine speed, with the coolant temperature held at 85 ^oC. The whole transition process was in an open loop control and the engine torque before and after the transition process was held constant. The throttle opening degree increased from 5.5% in the SI mode to 99% in the HCCI mode, the gasoline injection pulse was reduced from 4.85 ms to 3.1 ms, and the n-heptane injection pulse increased from 0 ms to 3.2 ms. As is shown, the transition process was completed after about 10 engine cycles.

3.2 Comparison between dual-fuel HCCI and gasoline HCCI combustion

Differences are observed between dual-fuel HCCI and gasoline HCCI at the same operation condition. Although the coolant temperature and lubricant oil temperature were the same at 90°C, the intake temperatures were quite different. At the condition of 34.5 N∙m/1600 r·min^–1, the intake air temperature of dual-fuel HCCI was 50°C with a gasoline ratio of 0.53, while the intake air temperature of gasoline HCCI was 185°C. The intake pressures of gasoline HCCI and dual-fuel HCCI are 0.866 bar and 0.985 bar, respectively. As depicted in Fig. 3, the cylinder temperature of gasoline HCCI is higher for its higher intake temperature, and the higher intake temperature of gasoline HCCI also reduces the intake charge quantity, thus causing a decreased cylinder pressure. However, gasoline HCCI has a higher peak heat release rate, which compensates for the pressure difference before ignition and finally produces a similar peak cylinder pressure compared to dual-fuel HCCI.

Generally, the cycle-to-cycle variation of dual-fuel HCCI is apparently lower than that of gasoline HCCI at the same condition, as demonstrated in Fig. 4. The cycle variations of IMEP, pressure peak and CA50 in the gasoline HCCI mode are 5.72%, 4.25% and 10.9%, respectively. In the dual-fuel HCCI, however, the values are 3.20%, 3.43% and 8.00%, respectively. This probably results from the improved compression-ignition behavior caused by the high cetane number of n-heptane. The low-temperature heat release of n-heptane increases the robustness of ignition, thus leading to an increased stability in combustion phasing.

Figures 5 and 6 describe the cylinder-to-cylinder variation of IMEP and the maximum pressure rise rate, respectively. At the same condition, the cylinder-to-cylinder variation of gasoline HCCI is much higher than that of dual-fuel HCCI. At 34.5 N∙m/1600 r·min^–1, the intake temperature needed in the gasoline HCCI is 185°C.At the same load, the temperature in dual-fuel HCCI is only 50°C. The higher intake temperature results in an increased heat loss. For an in-line four-cylinder engine, the heat loss in the first and fourth cylinders located at the engine ends is more than that in the second and third cylinder, located in the middle of the engine. The extent of heat loss directly influences the intake charge temperature of each cylinder, and may result in an increased cylinder-to-cylinder variation in the combustion process. In addition, the higher reactivity of n-heptane in dual-fuel HCCI may contribute to ignition robustness, thus reducing the combustion stability.

The cylinder-to-cylinder variation has significant influences on multi-cylinder engine operation. First of all, the high cylinder-to-cylinder variation is not advantageous for load extension of a multi-cylinder HCCI engine. As illustrated in Fig. 6, at 34.5 N∙m/1600 r·min^–1, for the gasoline HCCI engine, the mean value of the maximum pressure rise rate in the first and fourth cylinders is smaller than 5 bar/CA; meanwhile, the mean value of the maximum pressure rise rate in the second and third cylinders has already exceeded 8 bar/CA. Therefore, the maximum load for the HCCI engine is constrained by the pressure rise rate in the second and third cylinder in this case. Second, the high cylinder-to-cylinder variation may affect the balance of the block and crankshaft, and decrease the engine reliability.

3.3 Operational range extension of dual-fuel HCCI engine

3.3.1 Effects of EGR rate on dual-fuel HCCI combustion and operation extension

To study the effects of EGR on dual-fuel HCCI engine combustion, the engine speed was held at 1600 r/min, the intake temperature at 150°C, and coolant and lubricant oil temperatures at 90°C. The cycle fuel energy was kept at 296.8 J. The EGR rate was varied from 0% to 45%, and the gasoline ratio was adjusted from 0.53 to 0.65. Figure 7 displays the effects of gasoline ratio and EGR rate on HCCI combustion. At a given EGR rate, a decreased gasoline ratio advances the HCCI combustion phasing, shortens the combustion duration and elevates the cylinder pressure. With a reduced gasoline ratio, the low temperature heat release amount increases and cylinder charge temperature rapidly rises. The active radicals released in the low temperature reaction also increase the combustion rate, shortening the combustion period and advancing combustion phasing.

In general, an increased EGR rate could retard HCCI combustion phasing and extend combustion duration, at a given gasoline ratio. However, with a small EGR rate (e.g. 5%), an advanced HCCI combustion phasing and an increased cylinder pressure are observed. The inhibition effects of exhaust gas on HCCI combustion become significant with an increased gasoline ratio. For example, at a gasoline ratio of 0.61, an EGR rate of slightly above 30% could result in misfire. The effects of EGR rate on HCCI combustion could be divided into reduction of mixture oxygen concentration, reduction of cylinder temperature by increased mixture specific heat, and reaction promotion by the active radicals from exhaust gases. With a low EGR rate such as 5%, the promotion of active radicals from EGR might play a dominant role, causing increased cylinder pressure and advanced combustion phasing. For example, a small amount of NO has been found to promote hydrocarbon auto ignition [32]. As the EGR rate is sufficiently high, the inhibition effects are more dominant. At this time, a further increase in EGR rate could significantly reduce the oxygen concentration and cylinder temperature, and retard combustion phasing.

Generally, at a constant gasoline ratio, IMEP first increases and then decreases with an increased EGR rate, as exhibited in Fig. 8. Besides, it can be found that the EGR rate achieving the maximum IMEP, which is named the critical EGR rate, is gradually reduced with an increased gasoline ratio. In Fig. 9, the COVs for peak cylinder pressure (COV_ppeak) and IMEP (COV_IMEP) are investigated versus the EGR rate. It is observed that the same critical EGR rate exists for the engine operation variance. Below the critical value of the EGR rate, COV_ppeak and COV_IMEP do not significantly change but are maintained at a low level. However, when the EGR rate is above the critical value, COV_ppeak and COV_IMEP rapidly increase with the EGR rate. The critical EGR rate is reduced by an increased gasoline proportion. These critical EGR rates are also found to produce the maximum IMEP. The variance in IMEP is related to the combustion phasing such as CA10 and CA50. As shown in Fig. 10, at a given gasoline ratio, CA50 first remains in a linear trend with CA10, but then rapidly rises and falls out of this linear relation. As denoted in Fig. 10, the last CA50 point keeping the linear relation with CA10 (as circled) corresponds to the critical EGR rate. Below the critical EGR rate, CA50 linearly changes with CA10, and the retarded phasing in this range elevates IMEP and reduces combustion variance. As EGR rate further increases above the critical value, CA50 and CA10 are no longer in a linear relationship, causing increased combustion instability.

Figure 11 presents the HC and CO emissions versus EGR rate. HC and CO emissions do not change significantly with the EGR rate at low EGR rates, but increase dramatically in HC while CO emissions are observed as the EGR rate further increases. HC emissions mainly originate from the near-wall regions and are influenced by the cylinder temperature. At low EGR rates, HCCI combustion is stable and the cylinder temperature does not substantially change with an increased EGR rate; therefore, HC emissions are maintained at a similar level. However, when the EGR rate increases to a high level, the in-cylinder oxygen concentration and the cylinder temperature could be remarkably reduced, causing deteriorated combustion and producing a large amount of unburned hydrocarbons. Similarly, at low EGR rates, CO emissions are not significantly changed due to sufficient oxygen concentration. However, as the EGR rate is significantly increased, the heavily retarded combustion phasing and reduced oxygen concentration reduces the combustion completeness, and CO emissions dramatically increase. In this dual-fuel HCCI combustion mode, NO_x emissions are always held at a very low level (below 5 ppm).

Figure 12 shows the ringing intensity (RI) and COV_IMEP trends versus CA50. The ringing intensity is often used to evaluate the intensity of cylinder pressure waves, and it is defined as [33]

\mathrm{RI}={\displaystyle \frac{\mathrm{1}}{\mathrm{2}\mathit{\gamma }}}{\displaystyle \frac{{\left[\mathrm{0.05}{\left({\displaystyle \frac{\mathrm{d}P}{\mathrm{d}t}}\right)}_{\max }\right]}^{\mathrm{2}}}{{p_{}}_{\mathrm{p}\mathrm{e}\mathrm{a}\mathrm{k}}}}\sqrt{\gamma {R{T}_{}}_{\max }},

in which g is the specific heat, (dP/dt)_maxis the maximum pressure rise rate, T_max is the peak cylinder temperature, and p_peak is the peak cylinder pressure. At a given gasoline ratio, ringing intensity decreases monotonically as CA50 increases. Meanwhile, a critical CA50 could be observed for the optimal COV_IMEP. As CA50 is advanced from this critical value, COV_IMEP increases due to the increasingly violent combustion; as CA50 retards from this critical value, the extent of late combustion increases, causing increased combustion instability. For automotive engines, the limits of RI and COV_IMEP are commonly considered as 5 MW/m² and 5% [34,35], respectively, which could define an adjustable range in CA50. As shown in Fig. 11, a reduced gasoline ratio could increase the CA50 range, which is beneficial for the operational range extension in the HCCI engine.

3.3.2 Effects of intake temperature on dual-fuel HCCI combustion and operation extension

To compare the effects of intake temperature on dual-fuel HCCI combustion, the intake temperature was adjusted from 55 °C to 170 °C. Meanwhile, the engine speed was held at 1600 r/min, and the EGR valve was always closed. The coolant and lubricant oil temperatures were held at 90 °C, cycle fuel energy was kept at 296.8 Joule, and the gasoline ratio was varied from 0.57 to 0.65.

Figure 13 describes the cylinder pressure and heat release rate versus intake temperature, with gasoline ratios of 0.57 and 0.61. As is shown in Fig. 13, the decreased intake temperature causes a retarded HCCI combustion phasing, reduced heat release rate and peak cylinder pressure, as well as extended combustion duration. At the compression stroke, the cylinder pressure slightly increases with a reduced intake temperature. Meanwhile, the increased gasoline ratio leads to a narrow adjustable range for intake temperature. For a gasoline ratio of 0.57, the HCCI combustion could be stable even though the intake temperature is as low as 60°C, while for the gasoline ratio of 0.61, misfire is observed as long as the intake temperature is below 90°C. A decreased intake temperature is also found to reduce the peak cylinder pressure as well as the peak pressure rise rate, as shown in Fig. 14. The influences of intake temperature on HCCI combustion can be elucidated in two aspects: first, the reduced intake temperature retards combustion phasing; and second, the reduced intake temperature raises the charge coefficient. Thus, at the compression stroke, the cylinder pressure slightly increases with the reduced intake temperature. In addition, the increased intake charge reduces the in-cylinder equivalence ratio, causing an extended combustion duration, a reduced peak cylinder pressure, and a pressure rise rate. With an increased gasoline ratio, the fuel compression ignition tendency is reduced and the low temperature heat release process is inhibited, leading to a retarded combustion phasing and reduced combustion stability. Therefore, as the gasoline ratio increases, the adjustable range for intake temperature narrows down.

Figure 15 shows IMEP and COV_IMEP versus intake temperature. In general, IMEP first increases and then decreases with the reduced intake temperature, while COV_IMEP presents an opposite trend. IMEP is mainly influenced by the combustion phasing. At a high intake temperature, the combustion phasing is overly advanced, which increases the negative work done at the compression stroke and thus reduces IMEP. On the contrary, the dramatically reduced intake temperature can retard the combustion phasing, decrease combustion stability and even cause misfire. Therefore, a reasonable adjustment of intake temperature is beneficial for the HCCI combustion phasing control and optimization, elevating IMEP and reducing cycle variance. Besides, at a low gasoline ratio (e.g., 0.57),the maximum IMEP could be achieved within an intake temperature window from 80 °C to 120 °C. A critical intake temperature exists for COV_IMEP. As the intake temperature is above the critical value, COV_IMEP is maintained at a low level and not significantly changed with the intake temperature, but as the intake temperature is lower than the critical value, the further reduced intake temperature deteriorates HCCI combustion and causes increased combustion instability. The critical intake temperature for COV_IMEP increases with the gasoline ratio. Hence, the reduced gasoline ratio could avoid the strong reliance on the increased intake temperature in HCCI combustion, beneficial for the combustion optimization, the thermal efficiency improvement, as well as the combustion instability and noise control.

Figure 16 describes the effects of intake temperature on HC and CO emissions. Generally, HC and CO emissions increase with a reduced intake temperature. The reduced intake temperature retards HCCI combustion phasing and increases the excessive air coefficient, which slows down the heat release process. With a reduced in-cylinder temperature, the oxidation of HC and CO emissions is weakened, leading to more HC and CO emissions.

Ringing intensity and COV_IMEP are used to define an adjustable range for CA50, as shown in Fig. 17. The limits for ringing intensity and COV_IMEP are set to 5 MW/m² and 5%, respectively. As the gasoline ratio is 0.61, CA50 varies within a range of 5 ^oCA, while for a gasoline ratio of 0.57, the range of CA50 increases to 6.9 ^oCA.

3.4 Optimization of dual-fuel HCCI engine operation

At a natural aspirated condition, the operational ranges of the dual-fuel HCCI engine and gasoline HCCI engine are compared in Fig. 18. The operational range of the dual-fuel HCCI mode is significantly broadened compared to the gasoline HCCI mode, with a maximum IMEP and output torque of up to 6.8 bar and 65 N∙m, respectively. The operational limit at the low load end is also dramatically broadened and the engine could be stably operated at idle conditions.

Further, engine parameters such as intake boost, EGR, intake temperature were adjusted to maximize the operational range of the dual-fuel HCCI combustion. As shown in Fig. 19, the upper load and torque limits of this dual-fuel HCCI combustion can achieve 9.3 bar in IMEP and 86 N∙m in output torque, equivalent to 70% of the SI mode at the same speed.

Different combustion strategies are then compared at different operational conditions, with fuel consumption and emissions analyzed. The specific operation parameters are listed in Table 3.

Figure€20€compares€the€ISFCsof€different€engine€operation modes. At a condition of 28.5 N∙m/1600 r·min⁻¹, compared to the gasoline SI mode, the ISFCs of gasoline HCCI, boosted gasoline HCCI, dual-fuel HCCI and boosted dual-fuel HCCI modes can be reduced by 30%, 39%, 39% and 42%, respectively. On the other hand, at a relatively high load condition, the ISFCs of boosted gasoline HCCI, dual-fuel HCCI and boosted dual-fuel HCCI are reduced by 22%, 20% and 26%, respectively. Therefore, it is proved that the fuel efficiency of the dual-fuel HCCI combustion mode can be further improved by intake boost, EGR and temperature control optimization.

Figures 21–23 compare the HC, CO and NO_x emissions for different operation modes. Under low load conditions, HC emissions of the SI engine are below 0.3 g/(kw·h), which is far less than those from the gasoline HCCI engine. The HC emissions from the dual-fuel boost HCCI engine could be dramatically reduced, but still much higher than the SI mode. However, under a relatively high load condition, the dual-fuel boost HCCI combustion does not significantly improve HC emissions, as compared to gasoline boost HCCI combustion. In spite of the changed load conditions, SI combustion always produces the highest CO emissions of all combustion modes. All the HCCI combustion modes could dramatically reduce CO emissions. Compared to the SI engine, NO_x emissions are reduced by over 99% from all HCCI engines, of which the intake boost dual-fuel HCCI combustion could realize almost zero NO_x emissions.

4 Conclusions

In this paper, an active fuel design concept was proposed to a four-cylinder HCCI engine, to explore a potential pathway to optimize the HCCI engine combustion and broaden the operational range. The active fuel design concept was realized by real time dual-fuel injection of gasoline and n-heptane using two independent injection systems, with EGR rate and intake temperature adjusted. The major conclusions are as follows:

1) Rapid transition from SI to HCCI could be achieved by the active fuel design, which might provide a pathway to solve the start problem in the HCCI combustion mode.

2) The active fuel design concept could significantly increase the adaptability of HCCI combustion to reduced intake temperature and increased EGR rate. Critical values for EGR rate and intake temperature exist to maximize IMEP, improve thermal efficiency, reduce combustion instability, and minimize HC and CO emissions in this dual-fuel HCCI mode. With a decreased gasoline proportion, the critical value for EGR rate increases and the critical value for intake temperature decreases. Hence, an optimized gasoline ratio could substantially improve EGR tolerance and avoid the strong reliance on the increased intake temperature in HCCI combustion.

3) The operational range of the dual-fuel HCCI engine could be substantially extended and the fuel consumption and engine-out emissions further improved by optimizing EGR, intake boost and intake temperature. After the optimization of the operational parameters, the dual-fuel HCCI mode achieves IMEP and an output torque of up to 9.3 bar and 86 N∙m, equivalent to the 70% of the maximum torque of the SI engine at the same engine speed. Meanwhile, the fuel consumption and NO_x emissions are greatly reduced.

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