1 Introduction

There are two representative operating modes for internal combustion engines. The first one is a spark ignition (SI) engine using low-reactivity fuels such as gasoline, ethanol, and gaseous fuels. The second mode is based on the auto-ignition of fuel, known as a compression ignition (CI) engine using high-reactivity fuels such as diesel. These two engines possess different strengths and weaknesses.

The SI engine is easily controllable and exhibits low smoke emissions in the exhaust gases. However, due to concerns regarding the knocking phenomenon, it has a relatively low compression ratio (CR), resulting in a low gross indicated thermal efficiency (GIE). GIE is related to compression and expansion strokes, including the combustion period. On the contrary, the CI engine demonstrates high GIE due to its high CR, and its maximum load capacity is much greater than that of the SI engine, particularly under turbocharged conditions. Despite its advantages, the CI engine’s control is more complex, and it generates a high level of smoke emissions due to heterogeneous mixture combustion, namely diffusive flame (Heywood, 1988).

For these reasons, efforts have been made to combine the merits of both engines. In SI engine development, direct injection (DI) technology was developed to extend the load range without causing knocking. In addition, the stratified combustion method was introduced to reduce nitrogen oxide (NOx) emissions (Feng et al., 2019). Furthermore, lean SI combustion has been developed to achieve both high-efficiency and low-NOx emissions. However, the knocking phenomenon remains a major challenge, limiting the improvement of GIE through SI combustion alone (Hosseini & Chitsaz, 2023; Zhao et al., 2022).

Conversely, efforts have focused on enhancing the premixed rate of CI combustion. The most popular method is homogeneous charge compression ignition (HCCI) combustion. While pinpointing the inception of HCCI combustion is challenging, the research of Najt and Foster (1983) is considered seminal, as they utilized fully blended gasoline in a four-stroke engine. They demonstrated nearly complete auto-ignition of gasoline under part-load conditions, although controlling the combustion phase was challenging. Subsequently, Ryan III and Callahan (1996) experimented with diesel fuel in HCCI combustion by injecting diesel into the intake port (Gray III & Ryan III, 1997). The results indicated that while HCCI combustion reduced engine-out NOx emissions, total hydrocarbon (THC) emissions increased, and diesel required elevated intake temperatures to evaporate.

As researchers recognized the complexity of implementing HCCI combustion, concepts such as “partial” HCCI (pHCCI) combustion and “premixed” charge compression ignition (PCCI) were developed. In many studies, extending the ignition delay to enhance air–fuel mixing involved increasing exhaust gas recirculation (EGR) rates and adjusting early diesel injection timings (Akihama et al., 2001; Hasegawa & Yanagihara, 2003; Lu et al., 2011). However, these pHCCI and PCCI concepts suffered from shortcomings in terms of “physical mixing time,” limiting their application to low-load conditions.

In this context, the concept of dual-fuel premixed CI (PCI) combustion addressed the mixing time challenge. Inagaki et al. (2006) proposed dual-fuel PCI by simultaneously using gasoline and diesel fuel. Gasoline was fully premixed with air through port fuel injection (PFI), while diesel fuel served as the ignition source. This allowed for load extension in PCI combustion. Subsequently, researchers from the Wisconsin ERC group expanded this concept to reactivity-controlled compression ignition combustion, widely known as “RCCIw” (Kokjohn et al., 2011). RCCI combustion is a dual-fuel CI combustion aiming for HCCI combustion based on entirely premixed participant fuels. As such, diesel injection timing was significantly advanced compared to conventional diesel injection timings under RCCI combustion (Hanson et al., 2011; Kokjohn et al., 2011; Nieman et al., 2012).

However, due to the high premixed state of RCCI combustion, a high maximum pressure rise rate (PRRmax) in the cylinder could result, leading to issues related to noise, vibration, and harshness (NVH). Therefore, a general dual-fuel CI combustion with relatively late diesel injection should also be considered. Building on previous research by Lee et al. (2017), three modes of dual-fuel CI combustion were identified based on varying reactivity stratification levels. Among these modes, the first mode with late diesel injection exhibited a heat release rate (HRR) profile similar to neat diesel combustion. However, the rear part of the HRR curve under dual-fuel combustion closely resembled flame propagation. This suggests that high-reactivity stratification in the cylinder, combined with late diesel injection, led to a mixed combustion phenomenon involving both auto-ignition and flame propagation. This phenomenon strongly relates to the differing fuel reactivity in the cylinder (Bhagatwala et al., 2015; Zhou et al., 2017).

Consequently, evaluating the effect of the overall equivalence ratio on combustion characteristics for dual-fuel combustion optimization is necessary. In general, for novel CI combustion concepts including RCCI, creating a lean mixture condition is preferable by increasing intake pressure to suppress NOx formation. In addition, utilizing a high EGR rate is beneficial for reducing NOx emissions. However, since flame propagation occurs in dual-fuel CI combustion, adopting a lean mixture condition and using EGR may adversely affect combustion efficiency.

Hence, this study investigates the effect of varying overall equivalence ratios on dual-fuel combustion through experimental evaluation in a 0.4-L single-cylinder CI engine. To modify the overall equivalence ratio, intake pressure and EGR rate were independently controlled under part-load conditions (1500 rpm/gIMEP 0.6 MPa). The study primarily examines combustion characteristics such as HRR, engine-out emissions, and efficiencies, including combustion loss related to unburned emissions such as carbon monoxide (CO) and THC emissions in the exhaust gas.

2 Experimental Setup and Condition

2.1 Experimental Setup

A single-cylinder CI engine with 0.4 L displacement was used for these experiments. A solenoid diesel injector was equipped with a common rail system. Two solenoid gasoline port fuel injectors were equipped on the intake port with a fuel pressure of 0.5 MPa. The ratio between diesel and gasoline was calculated by the mass of each. More detailed specifications of the engine are introduced in Table 1. To control the engine, a 37-kW DC dynamometer was adopted. To measure the fuel flow rates, a mass burette type flowmeter (ONO SOKKI, FX-203P) for diesel and a mass flow meter (AVL, 7030 flow meter) for gasoline were used. The concentrations of NOx, THC, CO, CO2, and O2 were measured using an exhaust gas analyzer (Horiba, MEXA 7100-DEGR), and the smoke emission was measured using a smokemeter (AVL, 415 S). EGR and air were compressed simultaneously by a supercharger that was controlled separately from the engine system. Thus, this entire EGR system was similar to the LP (low pressure) EGR system. The EGR and air mixture was then cooled by an intercooler system, with the EGR rate being controlled by an EGR valve. The EGR rate based on the volumetric values was calculated from the CO2 fraction in the exhaust and intake gases. To measure the pressures, an absolute pressure transducer (Kistler, 4045A5) was used, and a relative pressure transducer (Kistler, 6055Bsp) was adopted for the in-cylinder pressure. Signals from the pressure transducers were recorded using a scale of one crank angle per 100 cycles for each case using a data acquisition (Kistler, KiBox

Table 1 Engine specifications
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To Go 2893) system. The experimental setup is depicted in Fig. 1.

Fig. 1
figure 1

Schematic diagram of experimental setup

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GIE and combustion losses were calculated using the below Eqs. (1) and (2). The low heating value (LHV) was 42.5 MJ/kg for diesel and 42.8 MJ/kg for gasoline.

$${\text{GIE}}=\frac{{W}_{\rm gross}}{{m}_{\rm gasoline}\times {Q}_{\rm LHV \,of\, gasoline}+{m}_{\rm diesel}\times {Q}_{\rm LHV \, of\, diesel}}$$
(1)
$$\mathrm{Combustion}\, {\rm loss }=\frac{{m}_{\rm THC \,of\, each\, cycle}\times {Q}_{\rm LHV\, of\, fuel}+{m}_{\rm CO \,of\, each \,cycle}\times {Q}_{\rm LHV \,of\, CO}}{{m}_{\rm total\, fuel}\times {Q}_{\rm LHV \,of\, fuel}}$$
(2)

2.2 Experimental Conditions

All the experiments were conducted at 1500 rpm. The total input energy, based on the LHV, was fixed at 570 J/str, resulting in a variation of the gross indicated mean effective pressure (gIMEP) close to 0.6 MPa. Two variables were manipulated to investigate the effect of overall equivalence ratios. The first variable was the intake pressure, which was altered within the range of 0.10–0.16 MPa, with intervals of 0.02 MPa, without employing external EGR. As the intake pressure varied, adjustments were made to the diesel injection timing to maintain a consistent mass fraction burned at 50% (MFB50) position. To aid in understanding, the variations in intake pressure were also applied to neat diesel combustion, specifically the diesel 100% condition.

The second experiment involved varying EGR rates in the context of dual-fuel combustion. With a fixed intake pressure of 0.10 MPa, EGR rates were altered across four distinct cases (0, 26, 32, and 37%). Similar to the previous scenario, the diesel injection timing was adjusted to retain the same MFB50 positions for achieving the timing of maximum brake torque (MBT). Since the influence of varying EGR on neat diesel combustion was already widely recognized, the varying EGR test was exclusively performed within the dual-fuel combustion context (if necessary, the general trend of neat diesel combustion with different EGR rates will be explained using previous research).

In both experiments involving dual-fuel combustion, the gasoline ratio to the total input energy differed. The common objective was to provide the highest possible fraction of gasoline while maintaining stable combustion (with a coefficient of variation of gIMEP lower than 3%). Consequently, for the varying intake pressure conditions, the gasoline fraction was set at 80% based on the LHV value. However, for the varying EGR test, this fraction was reduced to 40%. Detailed specifications are provided in Table 2.

Table 2 Experimental conditions
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3 Results and Discussion

3.1 1st Experiment: Varying Intake Pressures

3.1.1 Combustion Characteristics

In Fig. 2, HRR and in-cylinder pressure traces are presented, showcasing the effect of varying intake pressures under conditions of dual-fuel and neat diesel combustion. Due to the substantial 80% gasoline fraction, the HRR profile of dual-fuel combustion displayed dual peaks. The first peak was associated with the auto-ignition of diesel, a high-reactivity fuel, as well as a portion of gasoline. The second peak resulted from the auto-ignition of remaining fuels (Lee et al., 2017, 2021). With an increase in intake pressure, the HRR peak decreased, and the duration of HRR expanded.

Fig. 2
figure 2

HRR and in-cylinder pressure traces as varying intake pressures under dual-fuel combustion (a) and neat diesel combustion conditions (b) (ND: neat diesel, DF: dual-fuel)

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On the contrary, Fig. 2b depicts the scenario of neat diesel combustion, exhibiting a sole peak attributed to the premixed auto-ignition of diesel fuel. Moreover, as the mixture became leaner and intake pressure increased, the peak diminished, and the duration of combustion was extended. Hence, whether concerning auto-ignition or flame propagation, the trend remained consistent in both the conditions of neat dual-fuel and neat diesel combustion. A leaner mixture condition contributed to a prolonged HRR profile and a reduction in the peak, regardless of the specific combustion mode.

However, there were some differences in terms of the ignition delay, which is defined as the duration between the start of diesel injection (SOI) and MFB10, as well as the ratio between MFB10-50 and MFB50-90, as shown in Fig. 3. Concerning the ignition delay, it was observed to shorten with higher intake pressure under neat diesel combustion, whereas it became prolonged under dual-fuel combustion. The increase in intake pressure led to higher mixture density in the cylinder at the end of the compression stroke, resulting in a shorter ignition delay under the same reactivity conditions (Heywood, 1988; Siebers, 1999).

Fig. 3
figure 3

Ignition delay and combustion durations (MFB10-50, MFB50-90) as varying intake pressures under neat diesel and dual-fuel combustion conditions (ND: neat diesel, DF: dual-fuel)

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Conversely, under dual-fuel combustion, the dominance of low-reactivity fuel and the decreased reactivity gradient due to the advancement of diesel injection timing with increasing boost pressure contributed to the delay in the start of combustion (SOC) as the mixture became leaner and intake pressure increased. This phenomenon was influenced not only by “flame propagation” but also by the “self-ignition property” of fuels (Bhagatwala et al., 2015; Lee et al., 2017; Zhou et al., 2017).

What is intriguing is that while the total main combustion duration (MFB10-90) did not show significant differences between neat diesel and dual-fuel combustion, the front half combustion duration (MFB10-50) was notably longer in dual-fuel combustion compared to neat diesel combustion due to the low reactivity of gasoline (Kang et al., 2018; Lee et al., 2018). The front half duration extended with increasing boost pressure, irrespective of combustion types.

Conversely, the rear half combustion duration (MFB 50-90) shortened with increasing intake pressure. In addition, it is evident that MFB10-90 decreased under neat diesel combustion as boost pressure increased. However, under dual-fuel combustion, there was no substantial alteration in the total main combustion duration with varying intake pressure, because the extended MFB10-50 compensated for the shortened MFB50-90 duration.

In Fig. 4, the maximum in-cylinder pressure rise rate (PRRmax) results are depicted. Basically, PRRmax of neat diesel combustion was always higher than that of dual-fuel combustion. Also, for the both combustion conditions, PRRmax became decreased as increasing boost pressure. If this PRRmax result was considered in conjunction with the combustion duration in Fig. 3, total main combustion duration did not influence on the PRRmax. The duration of front half combustion was rather closely related with PRRmax (Lee et al., 2019).

Fig. 4
figure 4

PRRmax results as varying intake pressures under neat diesel and dual-fuel combustion conditions (ND: neat diesel, DF: dual-fuel)

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3.1.2 Emissions and Efficiency Characteristics

In this chapter, the discussion primarily revolves around engine-out emissions and efficiencies, taking into consideration the mentioned combustion characteristics. Figure 5 displays the results of gross indicated NOx (gISNOx) and smoke emissions. In the context of neat diesel combustion, gISNOx was consistently maintained at levels between 15 and 16 g/kWh. Although the concentration of NOx decreased as the mixture became leaner, the intake air volume increased, resulting in minimal change to the overall engine-out NOx emissions.

Fig. 5
figure 5

gISNOx (a) and smoke (b) emissions results as varying intake pressures under neat diesel and dual-fuel combustion conditions (ND: neat diesel, DF: dual-fuel)

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Conversely, under dual-fuel combustion, gISNOx underwent a significant reduction as boost pressure increased. Given that the majority of the fuels were already premixed due to the gasoline port fuel injection (PFI) condition, a leaner mixture contributed to a reduction in the overall formation of NOx emissions throughout the cylinder. On the contrary, in the context of neat diesel combustion, diffusive flames still originated from the diesel spray (Heywood, 1988). In addition, the decrease in the average in-cylinder temperature, indicative of combustion inefficiency, could potentially play a role in reducing NOx emissions.

Regarding smoke emissions in Fig. 5b, dual-fuel combustion exhibited minimal smoke emissions. Conversely, for neat diesel combustion, the concentration of smoke continuously decreased as the mixture condition became leaner.

Figure 6 is presented to facilitate the discussion on combustion inefficiency, specifically losses. Within the context of neat diesel combustion, CO emissions were observed to decrease as the overall equivalence ratio diminished, influenced by the function of lambda, or the equivalence ratio (Heywood, 1988; Koci et al., 2009). Conversely, gISTHC exhibited a slight increase as the mixture condition grew leaner, possibly due to overmixing or the presence of lean pockets (Kim et al., 2009; Maiboom et al., 2008). Consequently, the combustion loss in neat diesel combustion consistently remained below 1% of the total input LHV value. Notably, the combustion efficiency improved with a leaner mixture condition.

Fig. 6
figure 6

gISCO (a), gISTHC (b) and combustion loss (c) results as varying intake pressures under neat diesel and dual-fuel combustion conditions (ND: neat diesel, DF: dual-fuel)

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Conversely, in the context of dual-fuel combustion, the simultaneous increase in CO and THC emissions was observed as boost pressure increased. The highly diluted mixture condition hindered the local flame propagation of gasoline, thus exacerbating combustion inefficiency (Bhagatwala et al., 2015; Lee et al., 2017). If all gasoline fuels underwent auto-ignition, combustion efficiency would ideally improve with higher intake pressure, stemming from the increased density of the in-cylinder mixture. Therefore,

This trend could serve as evidence of local flame propagation within the dual-fuel combustion process.

Figure 7 illustrates that the GIE of neat diesel combustion consistently exceeded that of dual-fuel combustion. The GIE of neat diesel combustion improved with a leaner mixture condition, resulting from enhanced combustion efficiency and reduced average in-cylinder temperature. This led to a decrease in overall thermal losses, including exhaust and heat transfer losses (other thermal losses were calculated by subtracting GIE and combustion loss from 100%).

Fig. 7
figure 7

Energy budget in the aspect of GIE, combustion loss, and other thermal losses as varying intake pressures under neat diesel and dual-fuel combustion conditions (ND: neat diesel, DF: dual-fuel)

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Conversely, the combination of increasing intake pressure and adopting a leaner mixture condition led to a decline in the GIE of dual-fuel combustion due to a reduction in combustion efficiency. Hence, excessively lean mixtures are unsuitable for dual-fuel combustion.

3.2 2nd Experiment: Varying EGR Rates

In Fig. 8, we introduce the HRR and in-cylinder pressure traces, showcasing the variation in (external) EGR rates. Since SOI was meticulously controlled to maintain a consistent MFB50, there were no substantial changes observed in the HRR shape. Notably, in certain instances, the HRR peak increased with the supplementation of EGR. This phenomenon stemmed from the enhanced premixed condition brought about by the prolonged ignition delay due to the presence of the inert gas, i.e., EGR, as depicted in Fig. 9. The main combustion duration for cases involving the addition of EGR proved lengthier compared to scenarios without EGR. However, there was no statistically significant difference in the main combustion duration (MFB10-90) with variations in EGR.

Fig. 8
figure 8

HRR and in-cylinder pressure traces as varying EGR rates under dual-fuel combustion

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Fig. 9
figure 9

Ignition delay and main combustion duration (MFB10-90) as varying EGR rates under dual-fuel combustion

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In Fig. 10, we present the engine-out emissions from dual-fuel combustion, highlighting the variation in EGR rates. A trade-off relationship was evident between NOx and smoke emissions, mirroring the dynamics seen in conventional diesel combustion. To ensure combustion stability under high EGR rate conditions, a gasoline fraction of 40% was utilized, resulting in general combustion characteristics, NOx, and smoke emissions akin to those observed in conventional diesel combustion (Heywood, 1988; Ladommatos et al., 1997).

Fig. 10
figure 10

Emissions including gISNOx, smoke, gISTHC, and gISCO results as varying EGR rates under dual-fuel combustion

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However, in terms of unburned emissions such as CO and THC, gISTHC emissions displayed a decrease with increasing EGR rates. It appears that the impact of achieving a richer overall equivalence ratio through elevated EGR rates, leading to improved local flame propagation and the mitigation of localized over-mixing regions, was more substantial than the effect of reducing the average in-cylinder temperature due to EGR (Kim et al., 2009; Koci et al., 2009). At the highest EGR case, gISCO emissions increased due to the richer mixture condition.

In Fig. 11, there was no big difference in GIE as varying EGR rates. The highest GIE was found at the highest EGR rate condition. There is a possibility to improve some of combustion efficiency using appropriate EGR rate.

Fig. 11
figure 11

Energy budget in the aspect of GIE, combustion loss, and other thermal losses as varying EGR rates under dual-fuel combustion

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Lastly, Fig. 12 presents a comparison of emissions and efficiencies between dual-fuel combustion with and without the use of EGR under the same gISNOx level. While there was an advantage of achieving near-zero smoke emissions from highly lean dual-fuel combustion without EGR usage, the combustion loss was significantly higher, resulting in a noteworthy concern regarding unburned CO and THC emissions in the exhaust gas. Nonetheless, even with a smoke concentration of 14.77 mg/m3 from the dual-fuel combustion with EGR usage, it remained remarkably low, and GIE exhibited superiority. Therefore, the judicious utilization of EGR in dual-fuel combustion proved to be more appropriate than pursuing an extremely lean dual-fuel combustion approach (Lee et al., 2022).

Fig. 12
figure 12

Comparison of emissions and efficiencies between dual-fuel combustion with and without EGR usage under the same gISNOx level (DF: dual-fuel combustion)

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4 Conclusion

In this study, we investigated the impact of varying overall equivalence ratios on dual-fuel combustion by manipulating intake pressures and exhaust gas recirculation (EGR) rates, all under low-speed and low-load conditions. Below is a summary of the key findings from this study:

  1. (1)

    Combustion: With increasing intake pressure, the main combustion duration decreased in both cases. However, there were differences between dual-fuel and neat diesel combustion. In dual-fuel combustion, ignition delay lengthened, while neat diesel combustion exhibited a shortened ignition delay. The PRRmax of dual-fuel combustion consistently remained lower than that of neat diesel combustion due to the lower reactivity of gasoline and localized flame propagation. PRRmax decreased as boost pressure increased. Although there were no significant differences in combustion with varying EGR rates, certain scenarios showed higher peaks in the HRR under dual-fuel combustion due to improved premixed conditions.

  2. (2)

    Engine-out emissions: Boost pressure increases resulted in reduced gISNOx emissions under dual-fuel combustion, while they remained consistent under neat diesel combustion due to the diffusive nature of diesel combustion. However, the combustion loss in dual-fuel combustion, calculated using gISCO and gISTHC values, worsened due to the lower reactivity and slower local flame propagation. Notably, smoke concentration under dual-fuel combustion remained close to zero, irrespective of intake pressure. When exploring EGR rate variations, a clear NOx-smoke trade-off relationship emerged. On the contrary, gISTHC emissions decreased with increasing EGR rate, indicating that the richer mixture effect was more influential than the effect of inert gas.

  3. (3)

    Efficiencies: With an increase in intake pressure, the GIE of dual-fuel combustion decreased due to a decline in combustion efficiency, while GIE of neat diesel combustion improved. However, when EGR was introduced and maintained at the same MFB50 position, there was no significant alteration in GIE for dual-fuel combustion. In addition, combustion loss was mitigated when comparing conditions with and without EGR usage.

  4. (4)

    General conclusion: In general, a too lean mixture combined with high intake pressure hindered the combustion efficiency of dual-fuel conditions due to the presence of low reactivity fuels. Consequently, we recommend proper EGR utilization in dual-fuel combustion as opposed to pursuing an “ultra-lean” dual-fuel approach, given its positive impact on GIE and the reduction of unburned emissions such as CO and THC.