Considering the limited installation space of the buffer, the external size of the buffer should not be too large, and the diameter of the piston should be 20 mm. The piston rod is designed as a hollow rod so that the coil lead on the piston can be drawn from it. Therefore, the diameter of the piston rod should ensure that there is enough lead space. At the same time, the piston rod diameter should not be too large, otherwise it will affect the buffer force. Considering the above two factors, the diameter of the piston rod is 8 mm. The MR fluid was provided by Chongqing Materials Research Institute in Chongqing city, China. The magnetization curve is shown in Figure 5. Since the zero-field viscosity of the selected MR fluid is large (0.8 Pa\*s), it is necessary to select a larger annular gap width to achieve a smaller zero-field buffer force. However, too large an annular gap width will have an impact on the buffer force. Therefore, considering the above two factors, the annular gap width is 2.5 mm. The effective length of the piston is one of the main factors affecting the buffer force. Excessive effective length will result in a large appearance size of the MR buffer. Too small an effective length will lead to unstable output of buffer force. Therefore, considering the above two factors, the effective length is 10 mm. The main structural parameters of the MR buffer are shown in Table 2.

**Figure 5.** Magnetization characteristic curve of the MR fluid.

The finite element simulation of the magnetic field can be used to analyze the magnetic circuit of the MR buffer. Figure 6 shows the magnetic induction intensity and magnetic force line distribution of the MR buffer. Due to the axisymmetric structure of the MR buffer, only 1/2 of the finite element model of the MR buffer is established in the electromagnetic field simulation. As can be seen from Figure 6, the magnetic induction intensity at the annular gap is mainly concentrated near the effective length. The distribution of the magnetic induction intensity is uniform, and most of the magnetic force lines pass through the annular gap vertically.

The distribution curves of magnetic induction intensity at the annular gap under different coil current intensities are shown in Figure 7. As can be seen from Figure 7, the effective length of the buffer is the area where the magnetic field is mainly concentrated. The distribution of magnetic induction intensity in this region is more uniform, and the magnetic induction intensity outside the effective length is lower. With the increase of current intensity, the magnetic induction intensity at the effective length of the annular gap will also increase.

**Figure 6.** Magnetic field distribution of the MR buffer.

**Figure 7.** Magnetic induction intensity distribution curve in the annular gap.

#### **4. Multi-Physics FE Modeling and Simulation of the MR Bu**ff**er**

Figure 8 shows the multi-physical field model of the MR buffer established in COMSOL Multiphysics provided by COMSOL in Stockholm, Sweden. The model involves the coupling of the magnetic field and flow field. The moving grid module drives the piston up and down in the model to simulate the buffer force.

The Navier–Stokes equation is used in the fluid region and the Maxwell equation is used in the magnetic field region. The coupling relationship between magnetic field and flow field can be achieved by two variables, hydrodynamic viscosity η and magnetic inductance intensity B. The coupling mode is shown in Table 3. The multi-physical field simulation model takes the simulation results of the magnetic field as one of the input values of the flow field [24].

**Figure 8.** Multi-physical field model of the MR buffer.

**Table 3.** Multi-physical field equations.

When the magnetic field is applied to the buffer, the MR fluid is characterized by Bingham fluid characteristics. The fluid flow and velocity distribution of a typical Bingham fluid between two parallel plates are shown in Figure 9. The MR fluid can be divided into three regions along the radial direction. The yield flow regions are on both sides and the plug flow region is in the middle. In the upper and lower regions near the plate, the MR fluid is yielding flow. At this region, the shear stress of the MR fluid is greater than the yield stress, and the MR fluid is in the flow state. The shear stress of the MR fluid in the middle plug flow region is less than its yield stress, and the MR fluid is in a relatively static state. In Figure 9, the active region is defined as the other region except the piston winding part at the annular gap. This region is the most obvious region where the MR fluid has MR effects. *L* is the length of the region. *u* and *v* are the velocity of the MR fluid and piston, respectively. *h*1, *h*2, and *h* are the height of yield flow region, the height of the plug flow region, and the width of the annular gap, respectively.

**Figure 9.** Fluid flow and velocity distribution of the MR fluid.

The Bingham model used to describe the properties of the MR fluid can be expressed as follows:

$$\begin{cases} \quad \tau = \mu\_0 \gamma + \tau\_{y\prime} & \tau > \tau\_y \\ \quad \gamma = 0, & \tau \le \tau\_y \end{cases} \tag{2}$$

where μ0 is the zero field viscosity of the MR fluid, τ is the actual yield strength of the MR fluid, γ is the shear rate of the MR fluid, and τ*y* is the ultimate yield strength of the MR fluid.

However, due to the non-differentiability of the plug flow region in the model, it will lead to complications in the actual calculation process. One way to solve this problem is to first determine the boundary of the plug flow regions, and then describe the region as a region with a constant velocity vector. This approach, however, assumes a fully developed flow and become very difficult when dealing with complex annular gap geometries. Thus, it does not have universal applicability.

The method used in this paper is the Bingham plastic model proposed by Case et al. [25]. In this model, the continuous differentiability of yield stress is achieved by describing the plug flow region as a micro-flow region.

This model is described as:

$$
\pi = \left(\frac{\tau\_y \tanh(\xi \gamma)}{\sqrt{\varepsilon^2 + \gamma^2}} + \mu\_0\right) \gamma \tag{3}
$$

where ξ, ε are constant coefficients.

By dividing the two sides of Equation (3) by the shear stress γ at the same time, the viscosity formula of the MR fluid can be obtained as follows:

$$
\mu = \frac{\tau\_y \tanh(\xi \gamma)}{\sqrt{\varepsilon^2 + \gamma^2}} + \mu\_0 \tag{4}
$$

In the multi-physical field simulation software, the correlation between the magnetic inductance intensity B and the fluid viscosity μ at each node of the annular gap can be set up by Equation (4), so as to realize the multi-physical field coupling calculation of the magnetic field and the flow field.

In the simulation, the mesh is divided into moving meshes. The motion equation of the piston part is set to a sine with a frequency of 5 Hz and an amplitude of 4 mm. At the same time, the current in the buffer coil is set to 0 A, 1 A, and 2.5 A, respectively. Constant coefficients ξ and ε are set to 0.03 and 0.1, respectively. In the model, the integral of the total stress along the axis on the boundary of piston and piston rod is the result of the buffer force obtained by simulation.

Figures 10–12 show the velocity distribution and dynamic viscosity distribution of the MR fluid in the MR buffer. As can be seen from the figure, the dynamic viscosity of the MR fluid increases significantly in the active region of the MR buffer, and compared with other fluid regions, the flow velocity of the active region decreases significantly. The results show that the MR fluid in the active region have an MR effect and the dynamic viscosity increases due to the influence of the magnetic field.

**Figure 10.** Velocity distribution map of the MR fluid when the current is 0 A.

**Figure 11.** Velocity distribution map of the MR fluid when the current is 2.5 A.

**Figure 12.** Dynamic viscosity distribution map of the MR fluid: (**a**) I = 1 A; (**b**) I = 2.5 A.

#### **5. Validation with Experimental Results**

#### *5.1. Electromagnetic Force Characteristic Experiment of the EMLA*

This experiment mainly measures the electromagnetic force characteristic curve of the EMLA, and compares with the simulation results to verify the accuracy of the simulation model. Figure 13 shows the experimental bench for measuring the electromagnetic force characteristic of the EMLA. The experimental bench includes the EMLA, displacement sensor (500LCIT series displacement sensors), force sensor (LSR type tension pressure sensor, with a maximum range of 200 N), adjusting screw rod, and baffle. The displacement sensor and the force sensor are used to measure the displacement and electromagnetic force of the mover, respectively. The measured data are fed back to the PC for data processing. The adjusting screw rod is mainly used to adjust the position of the mover. The position of the mover is changed by 1 mm for every turn of the screw. The EMLA is the control objective.

**Figure 13.** Experimental bench for the electromagnetic force characteristic.

In this experiment, the position of the mover is changed by adjusting the screw, and current intensities of 5 A, 10 A, and 15 A is applied to the electromagnetic coil to measure the electromagnetic force characteristic curve of the EMLA. Figure 14 shows a comparison between the measured electromagnetic force characteristic curve and the simulated curve. As can be seen from the figure, the electromagnetic force characteristic curves obtained from the experiment and simulation are in good agreement. The electromagnetic force is proportional to the current, and it changes very little with the change of the mover position. This shows that the magnitude of the electromagnetic force is only slightly affected by the change of the mover position, which meets the requirements of the structural design.

**Figure 14.** Electromagnetic force characteristic curve of the EMLA.

#### *5.2. MR Bu*ff*er Force Measurement*

This experiment mainly measures the buffer force of the MR buffer, where a constant current intensity is applied in the buffer coil, and the piston reciprocates in a sinusoidal curve. Then the experiment results and simulation results are compared and analyzed. An experimental bench for measurement is shown in Figure 15. The experimental bench includes the EMLA, displacement sensor, force sensor, and MR buffer. The displacement sensor and force sensor are used to collect the displacement data and buffer force data in the experiment and feed them back to the controller and PC for data processing. The EMLA and MR buffer are the control objectives. In the experiment, the EMLA drives the buffer piston to reciprocate at frequencies of 1 Hz, 3 Hz, and 5 Hz, respectively, and the sinusoidal wave amplitude is 4 mm. At the same time, currents of 1 A and 2.5 A are applied to the buffer coil, respectively.

**Figure 15.** Experimental bench for the buffer force measurement.

Figure 16 shows a comparison between the experimental displacement–velocity curve and the target displacement–velocity curve of the MR buffer. As can be seen from Figure 16, a sudden change in the curve occurs during the start point and stroke end point. The reason for the sudden change is that the static friction will affect the displacement curve of the buffer piston at the beginning of the movement and reversal, which makes the experimental curve not follow the target curve.

**Figure 16.** Displacement–velocity curve of the MR buffer.

Figures 17–19 show comparisons between the experimental buffer force curve and the simulation curve. As can be seen from the figures, when the piston motion frequency is 5 Hz and the buffer coil current is 1 A and 2.5 A, the maximum buffer force measured is 10 N and 35 N, respectively. At the same time, the buffer force measured by different coil current intensities is different, showing that the coil current intensity is the main factor affecting the buffer force. As can be seen from the figures, the simulation results are in good agreement with the experimental results. The main deviation occurs at the piston's starting and reversing position. Due to the sudden change of displacement at this location, the measured buffer force curve will produce a certain sudden change at this location, and the mutation becomes more obvious as the frequency increases.

**Figure 17.** Buffer force curve at 1 Hz frequency.

**Figure 18.** Buffer force curve at 3 Hz frequency.

**Figure 19.** Buffer force curve at 5 Hz frequency.

#### *5.3. Seating Performance Experiment*

The main purpose of this experiment is to verify the seat buffering effect of the EMVT with the MR buffer. The experimental bench is shown in Figure 20. The experimental bench includes the EMVT with the MR buffer, displacement sensor, and valve components. Where the EMVT with the MR buffer is the control object, a displacement sensor is used to measure the displacement curve of the mover. In this experiment, the EMLA is used to drive the valve opening and closing. The buffer effect of the system is verified by applying the buffering force before the valve seating.

**Figure 20.** Valve seating test bench.

Figure 21 shows the experimental curve of the valve seat velocity and displacement when the valve lift is 8 mm. In this experiment, the coil in the buffer is energized before the valve seat to generate the buffering force, so as to reduce the valve seat velocity and realize the valve seat buffering. As can be seen from Figure 21, when the valve lift is 8 mm, the valve seat velocity decreases from 0.58 m/s to 0.14 m/s, and the valve rebound height decreases from 0.2 mm to 0.05 mm. The result shows that by applying buffering force before valve seat can reduce the valve seat velocity and restrain the valve rebound effectively. The seat buffer performance of the system is good.

**Figure 21.** Experimental result of seating performance when the valve lift is 8 mm.

#### **6. Conclusions**

In this paper, an EMVT with an MR buffer is designed. The structure scheme of the system is put forward and the main components of the EMLA and MR buffer are designed. According to the structure of the main components, the magnetic field simulation of the EMLA and the MR buffer and the multi-physical field simulation of the MR buffer are carried out, respectively, allowing an accurate estimation of the magnetic force and buffer force. Electromagnetic simulation results show that the maximum electromagnetic force that EMLA can achieve is 320.3 N. The experimental bench of the system is built, and the measured buffer can achieve a 35 N buffer force when the piston motion frequency is 5 Hz and the buffer coil current is 2.5 A. The experimental results are in good agreement with the simulation results. The seating performance of the system is verified. When the lift of the

valve is 8 mm, the seating velocity of the valve decreases from 0.58 m/s to 0.14 m/s, and the rebound height of the valve decreases from 0.2 mm to 0.05 mm. From the above results, the system can meet the electromagnetic force required when the valve is opened and the buffer force required when the valve is buffered, and it can reduce the valve seat velocity and restrain the seat rebound effectively. Therefore, the system has good seat buffering performance.

The requirements of the EMLA and MR buffer are higher for the EMVT with MR buffer. It not only requires the EMLA to have a large electromagnetic force, but also ensures that the buffering force can achieve seat buffering and quick response. Therefore, the EMVT with the MR buffer still needs further research on the buffering force response time of the MR buffer and the structure optimization of the system.

**Author Contributions:** Supervision: L.L., S.C., and Z.X.; research—original draft—editing: H.G.; system control: X.Z.

**Funding:** This work was supported by the National Natural Science Foundation of China (Grant No. 51975297, 51875290) and Shanghai Aerospace Science and Technology Innovation Fund (SAST2018-107).

**Conflicts of Interest:** The authors declare no conflict of interest.

#### **References**

© 2019 by the authors. Licensee MDPI, Basel, Switzerland. This article is an open access article distributed under the terms and conditions of the Creative Commons Attribution (CC BY) license (http://creativecommons.org/licenses/by/4.0/).

*Article*

### **Comparison of the Emissions, Noise, and Fuel Consumption Comparison of Direct and Indirect Piezoelectric and Solenoid Injectors in a Low-Compression-Ratio Diesel Engine**

### **Stefano d'Ambrosio \*, Alessandro Ferrari, Alessandro Mancarella, Salvatore Mancò and Antonio Mittica**

Energy Department, Politecnico di Torino, Corso Duca degli Abruzzi 24, 10129 Torino, Italy<EMAIL_ADDRESS>(A.F<EMAIL_ADDRESS>(A.M<EMAIL_ADDRESS>(S.M<EMAIL_ADDRESS>(A.M.)

**\*** Correspondence<EMAIL_ADDRESS>Tel.: +39-011-090-4415

Received: 19 September 2019; Accepted: 21 October 2019; Published: 23 October 2019

**Abstract:** An experimental investigation has been carried out to compare the performance and emissions of a low-compression-ratio Euro 5 diesel engine featuring high EGR rates, equipped with different injector technologies, i.e., solenoid, indirect-acting, and direct-acting piezoelectric. The comparisons, performed with reference to a state-of-the-art double fuel injection calibration, i.e., pilot-Main (*pM*), are presented in terms of engine-out exhaust emissions, combustion noise (CN), and fuel consumption, at low–medium engine speeds and loads. The differences in engine performance and emissions of the solenoidal, indirect-acting, and direct-acting piezoelectric injector setups have been found on the basis of experimental results to mainly depend on the specific features of their hydraulic circuits rather than on the considered injector driving system.

**Keywords:** solenoid injectors; indirect-acting piezoelectric injectors; direct-acting piezoelectric injectors; engine-out emissions; fuel consumption; combustion noise

#### **1. Introduction**

In the last few years, significant improvements have been made, by the automotive industry, in the development of innovative solutions that are able to comply with increasingly stringent international regulations, in terms of reduction of both CO2 and pollutant emissions [1]. One of the main research fields in the diesel sector deals with the development of the Common Rail (CR) fuel injection systems [2–4] and the related fuel injection strategies [5].

Fuel spray formation plays a fundamental role in diesel engines, as the interaction between the injected fuel and the in-cylinder intake charge affects pollutant formation to a great extent [6]. For instance, a greater air entrainment in the spray can lead to a reduction in soot emissions, but can also increase NOx as combustion temperatures increase [7]. In this regard, a major role is played by diesel injectors, which are one of the core components of CR systems [8,9]. Two main driving technologies exist to control the needle lift, and diesel injectors can thus be classified into solenoid- [10,11] or piezoelectric-driven [12,13]. In this latter category, it is possible to further distinguish between directand indirect-acting piezo injectors.

Solenoid injectors obtain the needle upstroke thanks to a magnetic field that is generated by means of an electric current circulating in a coil, while the piezoelectric ones use piezoelectric crystals that are able to elongate when subjected to an electric potential difference [14]. The solenoid technology has proved to be robust and simple, and it is the most widely adopted in diesel automotive engines. Nevertheless, it is often claimed that the piezoelectric alternatives have a wider range of potentialities

than their conventional solenoid counterparts [15]. Ueda et al. [16], applying the same fuel injection calibration (featuring a double pilot injection), stated that a 4th generation indirect-acting piezoelectric injector could achieve lower soot emissions, thanks to its steeper fuel injection rate than a 4th generation solenoid reference measurement. Moreover, a potential fuel consumption reduction of up to 2.8% was also pointed out, thanks to the fast opening and closing of the needle which creates a short injection and combustion duration. Kim et al. [17] compared direct- and indirect-acting piezo injector performance, and pointed out that direct-acting piezo injectors could reduce NOx emissions, with only slight HC penalties, in most of the analyzed tested conditions. In fact, most of the conditions considered in their study resulted in a higher heat release rate in the premixed combustion phase when an indirect piezo injector was used.

The fuel spray pattern and atomization characteristics of indirect-acting piezo injection systems may be different from those of the solenoid-driven technology, and may be a potential source of improvement for fuel consumption and reduction of exhaust pollutant emissions from the engine [18]. In this regard, not only can the particular driving technology (piezo or solenoid) play an important role, but also the hydraulic layout and the mechanical setup of each injector [19]. If a solenoid-driven injector features an innovative pressure-balanced pilot-valve layout, its static leakage can be reduced significantly, by up to 25% less than similar injectors featuring standard pilot valves [20,21]. Moreover, a further reduction, of up to 50%, may be obtained if the innovative pressure-balanced pilot-valve is coupled with an integrated Minirail [22]. The hydraulic setups commonly applied to modern indirect-acting piezo injectors may be endowed with 3-way pilot valves [23] and bypass-circuits [19], with both devices affecting the dynamics of the injector control chamber, which is placed at the rear of the needle. The 3-way pilot-valve has the potential to reduce the dynamic leakage of the injector, while the bypass-circuit can improve the dynamic response of the needle during the nozzle closure phase [23].

When the direct-acting piezoelectric technology is implemented, the injector needle is directly actuated by the piezoelectric driving system, without the need for any pilot-valve: in this way, a reduction of the injector leakage, major control of the injected quantity, flexible multiple injections and flowrate shaping capabilities are all claimed benefits of this solution [24].

The present work has the aim of comparing the performance of solenoid-driven, direct, and indirect piezo-driven injectors in terms of engine-out pollutant emissions, fuel consumption and combustion noise (CN), under different working conditions in a low-compression-ratio diesel engine for passenger cars. The considered solenoid-driven injector is equipped with a pressure-balanced pilot-valve and an integrated Minirail, while the indirect-acting piezo-driven injector is endowed with a bypass-circuit and the direct-acting one features a feedback control strategy of the injected fuel mass.

#### **2. Experimental Setup and Engine Conditions**

The experimental tests were performed on a 2.0 L, four-cylinder, four-stroke diesel engine, fueled with conventional diesel oil (EN 590), whose main technical specifications are reported in Table 1.

**Table 1.** Main technical specifications of the engine adopted for the experimental tests.

A simplified scheme of the engine is shown in Figure 1. It is homologated for Euro 5 regulations and equipped with a high-pressure CR fuel injection system, a twin-stage turbocharger regulated by means of valve actuators and wastegate valves and a short-route high-pressure EGR system endowed with an EGR cooler. The EGR valve is installed downstream of the EGR cooler.

**Figure 1.** Scheme of the tested engine installed on the dynamic test bench at the Politecnico di Torino.

The experimental tests were carried out at the dynamic test cell of the Politecnico di Torino (ICEAL—Internal Combustion Engines Advanced Laboratories). This test cell is equipped with an "ELIN AVL APA 100" AC dynamometer and an "AVL KMA 4000" fuel flowrate measuring system. The brake specific fuel consumption (*bsfc*) values reported in the "Results and discussion" Section were calculated as the ratio of this fuel flowrate measurement to the brake power of the engine.

An "AVL AMAi60" was used to measure the raw gaseous emissions from the engine. It includes a complete analyzer train endowed with devices capable of simultaneously measuring gaseous concentrations of NO, NOx, HC, CH4, CO, CO2, and O2 chemical species, while another analyzer train is only equipped with a CO2 measurement instrument, which was mounted at the intake manifold to estimate the EGR rate to the engine, defined as the ratio of the recirculated exhaust gas mass flowrate to the total mass flowrate inducted in the cylinders [25]:

$$X\_{EGR} = \frac{\dot{m}\_{EGR}}{\dot{m}\_{EGR} + \dot{m}\_a} \tag{1}$$

where . *mEGR* and . *ma* represent the EGR and fresh air mass flowrates, respectively. In the present investigation the calculation of *XEGR* was performed considering the accurate expression developed in [26], which requires the evaluation of the volumetric concentrations of all the species at the exhaust and the knowledge of the combustion air composition at the inlet. An "AVL 415S" smokemeter was used to measure the soot engine-out emissions.

The test engine was instrumented with a high-frequency piezoelectric pressure transducer (Kistler 6058 A), which was installed on the cylinder head to provide measurements of the gas in-cylinder pressure temporal traces in cylinder #2. A high-frequency piezoresistive pressure transducer (Kistler 4005 BA) was installed in the corresponding intake runner of the same cylinder to obtain the absolute in-cylinder pressure measurement, which was used as a reference. Furthermore, slow-frequency piezoresistive pressure transducers and thermocouples were installed at different positions of the gas flow path (e.g., upstream and downstream of the turbocompressor, intercooler and turbine, in the intake manifold, in the cylinder runners and along the EGR circuit) to perform pressure and temperature measurements (cf. Figure 1).

Finally, all the above-mentioned measurement devices are controlled by AVL Puma Open 1.3.2 and IndiCom automation software. Data post-elaboration was performed with AVL CONCERTO 5 software. AVL CONCERTO 5 was in particular used to calculate the reported CN values, based on the measured in-cylinder pressure signals, which were filtered by means of a low-pass filter with a cut-off frequency of 5 kHz. Each in-cylinder pressure acquisition was averaged over 100 consecutive engine cycles.

Tables 2 and 3 report the available data to establish the accuracy of the measured pollutant emission values. Previous works [27] have shown that the expanded uncertainties of pollutant emission measurements taken at this engine test facility fall within a 2–4% range. As far as the extended uncertainties pertaining to the brake specific emissions are concerned, the fuel flowrate system accuracy (0.1% over a 0.28–110 kg/h fuel flowrate measurement range) and the maximum errors of the engine speed (1.50 rpm at full scale) and torque (0.30 Nm at full scale) also have to be taken into account.

**Table 2.** Composition of the gas calibration cylinder and extended uncertainty (95% confidence interval).

#### **3. Injector Hydraulic Performance**

The three considered injector types, that is a new-generation indirect-acting solenoid (for brevity here referred to as IAS), an indirect-acting piezoelectric (IAP) and a direct-acting piezoelectric (DAP) injector, all feature the same nozzle, conical angle, and geometrical sizes of the needle. However, the IAP and IAS injectors both have a ballistic needle (i.e., it does not reach its stroke end position for all the engine working conditions) with the same key features, while the DAP injector needle is not ballistic and reaches the stroke end at 0.1 mm [12]. Consequently, the capability of the injected

fuel flowrate of the DAP injector differs from the other two solutions. In general, the injected mass results to be lower for medium to high ET and prail values, while it leads to a larger injected fuel mass for low ET and prail values, than for IAS and IAP types [22,23]. In the former case, this is due to the reduced needle lift of the DAP injector, which yields a smaller flow restriction in correspondence to the needle-seat passage. In the latter case, the DAP injector delivers more fuel than the other two solutions, because the elongation of its piezoelectric crystal is linked directly to the mechanical actuation of the needle, which is thus able to open more quickly than in the case of an indirect actuation for both IAP and IAS [22,23]. In fact, it has been verified that the injected fuel flowrate of the DAP injector increases with the highest time derivative, at least for lower prail levels than 1200 bar [23].

Figure 2 shows a schematic of the indirect-acting piezo-driven injector, endowed with the bypass-circuit [23], while Figures 3 and 4 represent the direct-acting piezo injector [12] and the solenoid-driven one, respectively; the latter highlights the pressure-balanced pilot-valve [22] and a standard pilot-valve.

**Figure 2.** Schematic of the indirect-acting piezoelectric injector.

**Figure 3.** Schematic of the direct-acting piezoelectric injector.

**Figure 4.** Solenoid injector. (**a**) Hydraulic circuit of the injector; (**b**) Pressure-balanced pilot-valve; (**c**) Standard pilot-valve.

#### **4. Results and Discussion**

The results pertaining to the experimental tests carried out at the dynamometer test cell are presented in what follows, and the performance of the engine is highlighted (in terms of engine-out pollutant emissions, *bsfc* and CN) when equipped with IAS, IAP, or DAP injectors. The engine was originally provided by the OEM with a state-of-the-art pilot-main (*pM*) fuel injection calibration referring to IAP injectors. Experimental tests were then carried out with DAP and IAS injectors instead of IAP ones, keeping the same fuel injection calibration, to perform a comparison and investigate their different performance.

Some engine operating points, which were representative of an application of the tested engine to a vehicle over the New European Driving Cycle (NEDC), were selected for the analysis. The resulting key-points (expressed in terms of speed *n* [rpm] × *bmep* [bar]) were in the 1.5 bar ≤ *bmep* ≤ 12 bar and 1000 rpm ≤ *n* ≤ 2750 rpm ranges. For conciseness reasons, only three key-points are analyzed hereafter: a low load and speed condition, namely 1500 × 2, (whose results are presented in paragraph 4.1), and two medium–high speed and load points, namely 2000 × 5 and 2500 × 8 (whose results are presented in paragraph 4.2). For each working condition, results are presented for the three different injector technologies along some EGR sweeps, to have a wider comparison than that based on the single baseline points. The aim is to compare the performance of IAP, DAP and IAS injectors when the same fuel injection calibration is implemented and to highlight the impact of the particular injector technology on CN, *bsfc*, and engine-out pollutant emissions.

#### *4.1. Low Load and Speed (1500* × *2)*

Starting from the baseline double *pM* calibration (whose main parameters are reported in Table 4) EGR trade-offs were performed at a low speed and load engine point (i.e., 1500 × 2), for the three different injector typologies under investigation. The results of this comparison, in terms of CN, fuel consumption and engine-out pollutant emissions, are shown in Figure 5. Different colors and symbols refer to different injector typologies, while the contoured symbols refer to the baseline calibration results. The *x*-axis reports the NOx brake specific emissions (in g/kWh) for each diagram, with the lowest values corresponding to the highest applied EGR fractions (XEGR).

**Table 4.** Main parameters of the reference *pM* calibration at 1500 × 2, with IAP injectors. The main quantity is set directly by the test bench to obtain the desired *bmep*.

The engine-out soot and NOx emissions (Figure 5a) do not show the typical trade-off behavior of the conventional diesel combustion mode, for any of the three examined injector configurations. A PCCI-like trend (which means a simultaneous reduction of soot and NOx emissions when *X*EGR grows) is clearly achieved in each case, thanks to the realization of a high degree of premixed combustion. The high *X*EGR, in addition to the somewhat delayed fuel injection timings (which make combustion occur mainly during the piston expansion stroke) and the reduced engine compression ratio, are all able to enhance the local mixing of the injected fuel plumes with the inducted charge, producing a highly premixed stratified charge and relatively low peak combustion flame temperatures [28]. Unfortunately, PCCI-like strategies lead to high incomplete combustion species (HC emissions are reported as a reference in Figure 5d, but significant levels were also obtained for CO emissions, which are not presented here for conciseness reasons) and fuel consumption (Figure 5c), which show rapidly worsening trends as the EGR fraction increases. If all the graphs in Figure 5 are compared, the DAP injector provides the worst results, in terms of CN, *bsfc*, and HC emissions, because it has the highest tendency toward premixed combustion of the three analyzed solutions.

**Figure 5.** Engine-out soot emissions (**a**) CN; (**b**), *bsfc*; (**c**) and engine-out HC emissions; (**d**) vs. NOx for the three injector typologies, at 1500 × 2.

The IAS and DAP injectors are both able to deliver fuel with an improved atomization compared to the IAP solution, especially when relativelylow prail and ET values areimplemented, asin this case. Theimproved atomization for the IAS injector can be mainly ascribed to the presence of the integrated Minirail in its hydraulic layout. The Minirail, in fact, is able to maintain a more stable (and thus, higher) injection pressure during both the pilot injection and the first part of the main shot, thereby counterbalancing the pressure reduction due to the injection in the injector delivery chamber, and allowing a faster needle opening [22]. The DAP injector features a faster needle upstroke because of its direct mechanical actuation on the needle, which gives rise to a steeper fuel injected flowrate, at least for prail values below 1200 bar [23]. This improved fuel atomization (with respect to the IAP solution) creates leaner equivalence ratios locally, increases the local oxygen concentration and improves the tendency toward low temperature premixed combustion through an increased autoignition delay and more delayed combustion developments. Unlike high temperature diesel ignition, which mostly arises through the chain-branching H + O2 = O + OH reaction (which means that a decrease in the local oxygen concentration enlarges the autoignition delay and slows down the reaction process), at lower temperatures typical of PCCI-like combustion fuel decomposition and ignition mainly proceed along a different chain-branching reaction (alkyl → alkyl peroxyl → hydroperoxyl alkyl → hydroperoxyl peroxyl → keto-hydroperoxide) [29,30]. In the latter case, unlike the former, the leaner local equivalence ratios (i.e., lower fuel concentrations) achievable with IAS and DAP injectors, thanks to the improved atomization, can reduce the concentrations of fuel radicals inside the air-fuel mixture, thus effectively slowing down the ignition process and increasing the premixed stage of combustion.

As already mentioned, the highest premixed degree is achieved using DAP injectors, thus justifying their worse HC, *bsfc*, and CN levels in Figure 5. The higher the premixed degree is, the more likely the occurrence of fuel over-mixing and flame quenching phenomena. These phenomena are thought to be the main mechanisms responsible for the emission of incomplete combustion species, i.e., unburned HC and CO, under low speed and low load conditions, together with wall wetting phenomena [31]. Moreover, these phenomena may also negatively affect the fuel economy.

The CN vs. NOx (or the CN vs. EGR rate, as NOx emissions have strong correlations with EGR fractions) trends reported in Figure 5b do not show any trade-off behavior. On the one hand, when the EGR rate increases, the combustion develops more in the premixed stage, primarily thanks to the dilution effect of the recirculated gases, which increase the autoignition delay [25], and this leads to higher HRR peaks, higher cylinder pressure derivatives during the main combustion and deteriorated CN. On the other hand, the reduced combustion velocity, which is induced by a decrease in the oxygen fraction [32], dampens the intensity of the premixed combustion. Furthermore, if the autoignition delay grows, the combustion process tends to develop further during the piston expansion stroke (a late PCCI mode is implemented) and is thus influenced more by the expansion cooling effect, which reduces the cylinder pressure rise. The latter effects tend to prevail over the dilution effect (cf. Figure 6), thus making CN decrease as the EGR rate increases.

**Figure 6.** Comparison of the HRR traces featuring different EGR rates, at 1500 × 2, for the three injector typologies under investigation: (**a**) DAP; (**b**) IAP; (**c**) IAS.

θ

The DAP injector provides a higher CN (and the worst CN-NOx trade-off) than the other two investigated solutions, if a similar EGR rate (i.e., similar intake oxygen concentrations) is applied, while the solenoid injector results to be the best one. Figure 7 shows the HRR traces related to the three injector typologies for almost the same intake oxygen concentration (around 17.5%, corresponding to an XEGR of around 44%). The highest HRR peak pertaining to the DAP injector is likely due to the combination of a longer autoignition delay (as already explained), which increases the premixed stage of combustion, and a larger fuel injected quantity at the start of combustion, thanks to the faster upstroke of its needle than that of the IAP and IAS solutions. The IAS injector features a lower CN level than the IAP one, and its combustion process develops further during the piston expansion stroke with consequent lower HRR peaks. This is mostly due to a better atomization of the IAS injector, which is reached because of the presence of the Minirail [22].

θ θ θ

**Figure 7.** HRR traces featuring similar EGR rates, at 1500 × 2, for the three injector typologies under investigation.

### *4.2. Medium–High Speed and Load (2000* × *5, 2500* × *8)*

Figures 8 and 9 report the soot, *bsfc*, CN and HC for the EGR trade-offs performed at two medium speeds and loads (engine points 2000 × 5 and 2500 × 8) for the three different injector types under investigation, in a similar way to the previous section. As has already been seen for the low load and speed engine conditions, the contoured symbols in each graph refer to the baseline tests obtained, for each injector type, when implementing the same baseline double *pM* calibration (whose main parameters are reported in Table 5) provided by the OEM and originally applied to the engine equipped with the IAP injectors.

**Table 5.** Main parameters of the reference *pM* calibration at 2000 × 5 and 2500 × 8, with IAP injectors.

The engine-out soot and NOx emissions at 2000 × 5 (Figure 8a) show a clear trade-off behavior, which is typical of a conventional diesel combustion mode, for each of the three examined injector typologies. Nevertheless, a low-NOx combustion can also be achieved for these higher load and speed conditions, through the application of a relatively delayed fuel injection pattern and a high EGR rate (cf. the contoured symbols in Figure 8, which represent the results pertaining to the reference calibration for each injector type). The worst soot-NOx trade-off can be observed for the IAP injector (cf. triangle symbols in Figure 8a), which features more than doubled soot emissions compared to the other two injector solutions for the highest EGR rates (lowest NOx values). This is because, as already mentioned, the IAP injector gives the worst atomization and the lowest premixed degree, thus entraining the least air inside the injected fuel plumes and locally creating the richest mixture portions, which are prone to forming soot particles.

**Figure 8.** Engine-out soot emissions (**a**) CN; (**b**) *bsfc*; (**c**) and engine-out HC emissions; (**d**) vs. NOx for the three injector typologies, at 2000 × 5.

Differences in fuel consumption (Figure 8c) have been highlighted between the three injectors under investigation. A mean penalty in *bsfc* of about 3–4% may be observed for the IAP solution, while DAP and IAS show almost the same behavior. However, the differences in fuel leakage between the IAP, DAP and IAS injectors should only play a negligible role at this engine operating point: the IAP and IAS injectors show comparable injector leakages, while the DAP injector, which features the lowest leakage values, does not show any appreciable *bsfc* reduction. The 3–4% reduction in *bsfc*, when IAS and DAP injectors are installed, is likely due to the enhancement in the thermal efficiency of the engine deriving from the higher premixed degree, which makes the combustion approach constant volume rather than constant pressure conditions and reduces its overall duration. Unlike the low load case, this gain in thermal efficiency is not offset by any severe CO (here not reported) or

HC penalties (Figure 8d). When load increases, concerns about incomplete combustion species are generally mitigated.

The CN vs. NOx trends (cf. Figure 8b) at 2000 × 5 show a slight trade-off behavior, i.e., a slight increase in CN when NOx emissions are reduced as a result of higher EGR rates, unlike what is highlighted at 1500 × 2. In this case, the larger premixed combustion phase prevails over the reduction in combustion velocity due to increased EGR and the damping effect resulting from the delayed combustion that develops along the expansion stroke. This leads to higher HRR peaks, higher cylinder pressure derivatives and increased CN as the EGR fractions grow. The lowest CN level of three injector types is achieved by the IAP injector, with improvement of up to 1–1.5 dBA with respect to the other solutions. The three HRR traces pertaining to the baseline tests, for each injector, are plotted in Figure 9 versus the crankshaft angle. It is clear that the IAP injector determines the lower ignition delay and, thus, less fuel to burn in the premixed stage, and this gives rise to the lowest HRR development and consequently to the lowest CN.

**Figure 9.** HRR traces of the reference calibrations, at 2000 × 5, for the three injector typologies under investigation.

The worst soot-NOx trade-off at 2500 × 8 may be observed for the DAP injector (cf. the circle symbol curve in Figure 10a), while the IAS injector still gives the best results (cf. the square symbol curve), even though the differences between the three solutions are not big. In the case of the IAS injector, the lowest soot production at each EGR level is still due to its higher tendency toward premixed combustion, as a consequence of the presence of the Minirail in the hydraulic layout, which determines a faster needle upstroke and, thus, a better fuel atomization in the first part of the injection. The fact that the DAP injector does not give better results than the IAP, as observed at lower loads, is due to the higher rail pressure (as can be seen in Table 5, the rail pressure at 2500 × 8 is equal to 1200 bar). The needle velocity of the DAP injector for prail ≥ 1200 bar during the upstroke tends to become similar to that of the IAP one, as the increased rail pressure level generates a higher pressure force that counteracts the needle lift.

Once again, the IAP injector shows an increased *bsfc* (about 3–4% more than the other injector typologies, cf. Figure 10c) compared to the other two injectors. The previously highlighted differences in the premixed combustion portion reflect on the fuel consumptions of the injectors, as occurred at 2000 × 5. Moreover, the considerations on the negligible levels of CO (here not reported) and HC emissions (Figure 10d) for all the considered injector types are even more valid at this higher considered load.

Finally, the CN (Figure 10b) has been confirmed to be higher for the DAP injector, even though it does not highlight the maximum HRR peak (cf. Figure 11) as it does at 2000 × 5. In fact, when double injection is performed, CN not only correlates with the heat release peak of the main injection (HRRpk2), but also with that of the pilot shot (HRRpk1) and with the timings of the two HRR bumps [33]. However, smaller differences are present at 2500 × 8 for the three injectors, than for the previously analyzed points.

Figure 11 reports the three HRR traces versus the crankshaft angle pertaining to the baseline tests at 2500 × 8 for all the injector types. Again, in this case, the IAP injector determines the lowest autoignition delay, even though the differences are smaller than in the previous cases.

**Figure 10.** Engine-out soot emissions (**a**) CN; (**b**) *bsfc*; (**c**) and engine-out HC emissions; (**d**) vs. NOx for the three injector typologies, at 2500 × 8.

θ **Figure 11.** HRR traces of the reference calibrations, at 2500 × 8, for the three injector typologies under investigation.

#### **5. Conclusions**

The experimental investigation has concerned a comparison of the performance of indirect-acting solenoid-driven, direct, and indirect piezo-driven injectors in terms of engine-out pollutant emissions, *bsfc* and CN, under different working conditions in a low temperature combustion diesel engine for passenger cars (a pilot-Main *pM* fuel injection strategy was adopted in all of the tested calibrations).

It is possible to avoid the typical NOx-soot trade-off of a conventional diesel combustion mode at low loads and speeds, i.e., at 1500 × 2, and thus to simultaneously obtain low-NOx and low-soot PCCI-combustion for all of the three investigated injector types, even though the PCCI behavior is more pronounced for the DAP injector. In fact, the DAP injector provides penalties of about 2–3 dBA, in terms of CN, and the worst *bsfc* and HC trade-offs. In general, both the IAS injector (due to the presence of the Minirail integrated in its hydraulic layout) and DAP injector (due to its direct mechanical actuation on the needle) are able to deliver the fuel with an improved atomization compared to the IAP solution, and this leads to a more pronounced premixed combustion stage.

The engine equipped with each of the analyzed injector solutions shows a conventional diesel combustion mode at medium loads and speeds, i.e., at 2000 × 5 and at 2500 × 8. This can be inferred from the obvious presence of a NOx-soot trade-off. The worst soot-NOx trade-off at 2000 × 5 is shown for the IAP injector, due to its poorer fuel atomization and the lower premixed degree than the other two injector solutions. Instead, it is the DAP injector that gives the highest soot emissions at 2500 × 8, for any EGR rate applied. In fact, the needle velocity of the DAP injector tends to be slower during the nozzle opening phase at higher speed and load, as the accordingly increased rail pressure generates a higher pressure force that counteracts the needle lift.

A mean penalty in *bsfc* of about 3–4% was observed for the IAP injector, at both 2000 × 5 and 2500 × 8, whereas CN was confirmed to be higher for the DAP injector, although smaller differences between the injector types were present at 2500 × 8 than at the lower load points.

In general, the performance differences between the analyzed injector technologies can mainly be ascribed to the different layout solutions present in their internal hydraulic circuits (such as the bypass, the pressure-balanced pilot-valve and the Minirail), rather than to the injector driving system (solenoidal, indirect, or direct piezoelectric).

**Author Contributions:** The authors contributed equally to the preparation of the paper. Conceptualization, S.d., A.F.; Methodology, S.d., A.F.; Data Curation, S.d., A.F.; Writing-Original Draft Preparation, A.M. (Alessandro Mancarella); Writing-Review and Editing, S.d., A.F., A.M. (Alessandro Mancarella); Supervision, S.M., A.M. (Antonio Mittica).

**Funding:** This research received no external funding.

**Conflicts of Interest:** The authors declare no conflict of interest.

#### **Abbreviations**

θ Rotational angle of the crankshaft

#### **References**

© 2019 by the authors. Licensee MDPI, Basel, Switzerland. This article is an open access article distributed under the terms and conditions of the Creative Commons Attribution (CC BY) license (http://creativecommons.org/licenses/by/4.0/).

### *Article* **Combustion Analysis of Homogeneous Charge Compression Ignition in a Motorcycle Engine Using a Dual-Fuel with Exhaust Gas Recirculation**

**Yuh-Yih Wu 1,\* and Ching-Tzan Jang 1,2**

Received: 25 January 2019; Accepted: 25 February 2019; Published: 4 March 2019

**Abstract:** Exhaust emissions from the large population of motorcycles are a major issue in Asian countries. The regulation of exhaust emissions is therefore becoming increasingly stringent, with those relating to nitrogen oxides (NOx) the most difficult to pass. The homogeneous charge compression ignition (HCCI) has special combustion characteristics and hence produces low NOx emissions and exhibits high thermal efficiency. This study developed an HCCI system for a 150 cc motorcycle engine. The target engine was modified using a dual-fuel of dimethyl ether (DME) and gasoline with exhaust gas recirculation (EGR). It was tested at 2000–4000 rpm and the analysis was focused on 2000 rpm. The DME was supplied continuously at an injection pressure of 1.5 kg/cm2. The gasoline injection rate was adjusted at a pressure of 2.5 kg/cm2. A brake-specific fuel consumption of <250 g/kW·h was achieved under a condition of air–fuel equivalence ratio (λ)<2 and an EGR of 25%. The nitric oxide concentration was too low to measure. The brake mean effective pressure (BMEP) increased by 65.8% from 2.93 to 4.86 bar when the EGR was 0% to 25%. The combustion efficiency was close to 100% when the BMEP was >3 bar.

**Keywords:** homogeneous charge compression ignition (HCCI); exhaust gas recirculation (EGR); dual-fuel; dimethyl ether (DME); exhaust emission

#### **1. Introduction**

The number of motorcycles in Asia is extremely large. This is because motorcycles are low in cost and small in size, allowing freedom of movement in crowded areas, easy parking, and high mobility. For example, the density of motorcycles in Taiwan in 2018 was 384/km2 [1], the highest in the world. Motorcycles contribute to air pollution more than other vehicles [2,3]. They not only damage human health but also contribute to global warming. Consequently, regulations on their exhaust emissions and fuel consumption are becoming more stringent.

For instance, the Economic Commission of Europe (ECE) established long-term stage-by-stage emission regulations for motorcycles from 1999, as shown in Figure 1. The Taiwan Environmental Protection Administration (EPA) implemented motorcycle emission regulations from 1988, in a stage-by-stage approach. Taiwan's regulations were harmonized with ECE regulations from their fifth stage, which is similar to EURO 3 (i.e., the third stage of the ECE). The double-arrow red lines in Figure 1 link the corresponding stages between Taiwan and ECE. The implementation of Taiwanese regulations occurred approximately 1 year later than ECE regulations.

**Figure 1.** Evolution of motorcycle emission regulations by the ECE and Taiwan (WMTC: worldwide motorcycle test cycle; OBD: on-board diagnostics).

Table 1 shows the current and next-stage emission standards for motorcycles in Taiwan [4]. This encompasses carbon monoxide (CO), hydrocarbons (HC), non-methane hydrocarbons (NMHC), nitrogen oxides (NOx), and particulate matter (PM). The emission standards of the sixth and seventh stages in Table 1 are the same as those in EURO 4 and 5, respectively. Numerous Asian countries follow the ECE regulations. To comply with emission standards, all the motorcycles produced in Taiwan now use an electronic fuel injection system. The NOx standard is the most difficult to pass. In-use motorcycles emission controlled by Taiwan's EPA have even been recalled because of high NOx emissions. Although lean-burn can improve fuel economy, motorcycle manufacturers do not use this approach. This is because NOx emissions are difficult to reduce in a lean-burn system. The seventh-stage standard will be implemented from 1 January 2021. This will be extremely strict, which means that new technologies must be developed.

**Table 1.** Emission standards of motorcycles in Taiwan.

1 Particulate matter (PM) is for gasoline direct injection engines only.

In a spark ignition (SI) engine, a three-way catalytic converter efficiently reduces HC, CO, and NOx, although the mixture must be maintained within a narrow air-fuel-ratio window. Other technologies for reducing NOx include exhaust gas recirculation (EGR) [5], selective catalytic reduction (SCR) [6], a NOx trap [7], a plasma reactor [8], water injection [9], water/fuel emulsions [10], and homogeneous charge compression ignition (HCCI) [11–13]. SCR and lean NOx-trap have been applied in diesel vehicles to meet standard emission limits [14]. However, these systems increase production costs and cannot be used in motorcycles. An HCCI engine has special characteristics, as it constitutes a combination of an SI engine and a diesel engine. The homogeneous charge of fuel–air mixture is similar to that in an SI engine, whereas the compression ignition is similar to that in a diesel engine. Such combustion characteristics produce very low amounts of NOx emissions and exhibit high thermal efficiency.

HCCI is an advanced low-temperature combustion concept that has attracted global attention in recent years [12]. However, the operational range of HCCI combustion is restricted due to the absence of direct control of the ignition timing and heat release rate [15]. In short, the key issues affecting HCCI combustion are thermal effects and chemical kinetics.

The factors that cause thermal effects include negative valve overlap (NVO), intake heating, glow plug, spark-assisted combustion, and compression ratio. NVO engines retain hot residual gas in the cylinder to heat the intake charge, extending the engine load range [16]. The intake heating improves autoignition, advances the combustion phase, and shortens combustion duration [15]. Spark-assisted HCCI reduces combustion noise under high-load conditions [17,18] and extends the operating load up to 750 kPa indicated mean effective pressure (IMEP) [18]. The glow plug can be used to control the heat release rate, and the temperature distribution is broadened to heat the charge unevenly [19]. A suitable compression ratio is necessary, which is from 12 to 18 according to most studies on HCCI engines.

The factors that cause the chemical kinetic effect include fuel composition, fuel properties, dual-fuel, external EGR, EGR stratification, and fuel stratification. One of the vital properties of fuel used for HCCI is its autoignition temperature. Fuel with a low autoignition temperature is termed diesel-like fuel and fuel with a high autoignition temperature is called gasoline-like fuel. To ensure suitable ignition timing and combustion phasing, an additive fuel is used. Diesel-like fuel can be used as an additive to enhance autoignition in a gasoline-like fuel system. Gasoline-like fuel can also be used as an additive to inhibit autoignition in a diesel-like fuel system.

Ji et al. [20] used a diesel-like fuel of 2-ethylhexyl nitrate to enhance the autoignition of an E10 gasoline HCCI engine. A maximum indicated thermal efficiency of 50.1% was found at 1800 rpm and an intake pressure of 180 kPa. Wang et al. [21] developed a DME-diesel blend fuel system using the gasoline-like fuel of liquefied petroleum gas as an ignition inhibitor. Pedersen et al. [22] used the port fuel injection (PFI) of methanol in the direct injection (DI) of a DME engine. The added methanol increased the BMEP and slowed the combustion so that it was after top dead center (aTDC). Mohebbi et al. [14] applied a PFI of ethanol and diethyl-ether blend fuel in a DI diesel engine: this resulted in a 14% increase in IMEP and a 33% reduction in the maximum rate of pressure rise (MRPR). Li et al. [23] investigated a blend of n-butanol and n-heptane in HCCI combustion and found that the knock tendency decreased as the n-butanol volume fraction increased. Khandal et al. [24] used a PFI of hydrogen in the DI of a biodiesel engine, which resulted in 65%–67% less smoke and 98%–99% lower NOx emissions. Finally, Zheng et al. [25] developed a dual-fuel system using the DI of biodiesel and PFI of several gasoline-like fuels. Among the gasoline-like fuels, the biodiesel-ethanol produced low NOx and soot emissions (soot < 0.3 FSN and NOx < 1.5 g/kW·h).

Regarding the mechanism of the effect of EGR on combustion, the addition of EGR slows the decomposition of hydrogen peroxide (H2O2) by reducing the rate of hydroxyl radicals (OH) [26]. This results in the suppression of autoignition and delay of combustion phasing, thus allowing high load operation. Researchers [27,28] have reported that in-cylinder EGR stratification reduced the MRPR of HCCI engines. Other researchers [13,29] have reported that the effects of external EGR on HCCI engines were a delay in combustion phasing, decrease in maximum temperature and maximum HRR (heat release rate), prolonged high-temperature heat release, and decreased MRPR. Lee et al. [30] investigated the optimization of EGR (0–25%) and two-stage injection with compression ratios of 15 and 17.8.

Superior HCCI operation can be achieved using a combination of several strategies, such as the stratification of external EGR, fuel stratification using PFI and DI injectors including asymmetric injection, open valve injection using a PFI injector, and NVO injection [31]. Researchers [16,32] have developed an HCCI engine with intake boosting, NVO, and external EGR. The engine load was extended to an IMEP of 8 bar. This type of system can achieve a thermal efficiency of 47%, NOx of ≤0.1 g/kW·h, and combustion efficiency of ≥96.5% [16].

The best elements of previous research have been developed for automobile engines and cannot be used for motorcycles. Most motorcycle engines are small-scale, with a displacement ≤150 cc, while the surface-volume ratio is large compared to automobile engines and hence causes high heat loss [33]. This condition makes HCCI operation in motorcycles difficult. Additionally, the combustion characteristics of a small-scale air-cooled HCCI engine are insufficient. Therefore, a combination of DME-gasoline dual-fuel and external EGR with a suitable compression ratio for HCCI operation was developed. This is useful for low-cost motorcycle engines with only minor modifications. The purpose of this study was to therefore investigate the combustion characteristics of this HCCI system.

#### **2. Experimental Methodology**

#### *2.1. Experimental Setup*

A commercial motorcycle engine (SYM, Taipei, Taiwan) with an electronic fuel injection system was retrofitted for HCCI operation. The engine was a 150 cc air-cooled SI engine. Detailed engine specifications are listed in Table 2. The compression ratio was increased from 10.5 to 12.4 by replacing the cylinder head with a smaller clearance volume. The increased compression ratio raises the temperature of the compressed mixture, which easily achieves compression ignition.

**Table 2.** Engine specifications.

1 Valve timing is defined at 1 mm of valve lift. aTDC: after top dead center; bTDC: before top dead center; aBDC: after bottom dead center; bBDC: before bottom dead center.

The fuel for HCCI operation in an SI engine should have a low autoignition temperature. Previous research [34] has concluded that DME, the properties of which are listed in Table 3, is a favorable choice for this. The cetane number of DME is 60, higher than that of diesel fuel (i.e., 40–60). It can autoignite in the target engine. Therefore, DME was selected as the main fuel. Increasing the fuel amount of DME to increase engine load will cause high MRPR or engine knocking.

1 Autoignition temperatures were obtained from Material Safety Data Sheets.

Gasoline with a research octane number (RON) 92 and external EGR were added to extend the operating range of the engine. The autoignition temperature of gasoline is much higher than that of DME, as shown in Table 3. The addition of gasoline can therefore increase the engine load without knocking.

The experimental setup for the proposed HCCI engine is shown in Figure 2. The external EGR system was built on the target engine with a small EGR pump and a control valve, which adjusted the EGR ratio. A surge tank was installed between the air flow meter and the throttle; it was used for attenuating the pressure pulsation in the intake system of the engine. So the intake air flow rate can be measured stably. The volume of the surge tank is 40 L, which is larger than the minimum requirement calculated according to SAE J244. A dual-fuel supplying system was then built into the target engine (Figure 3). The original fuel and ignition systems of the target engine were retained to start the engine. The spark plug was used only for starting the engine and the spark timing was the same as the original engine. The original gasoline injector was also used for the addition of gasoline in the HCCI engine. The additional DME supplying system, which includes a DME tank, pressure regulator, filter, and a flow meter, was attached to the target engine with a DME supply tube installed near the intake port.

**Figure 2.** Experimental setup for the proposed dual-fuel HCCI engine with external EGR.

**Figure 3.** Experimental setup for the fuel supply system.

For the engine test, an eddy-current engine dynamometer (FE150-S, Borghi & Saveri S.R.L., Bologna, Italy) was used to measure the engine speed and brake torque. The gasoline flow rate was measured using a mass burette flow detector (FX-1110, Ono Sokki, Yokohama, Japan). The DME flow rate was measured using a thermal mass flow controller (NM-2100, Tokyo Keiso, Tokyo, Japan). The exhaust emissions of CO, HC, NO, carbon dioxide (CO2), oxygen, and the lambda were measured

using an emission analyzer (MEXA-584L, Horiba, Kyoto, Japan). The lambda is calculated by the carbon balance method in standard configuration using the MEXA-584L emission analyzer. The hydrogen/carbon ratio and oxygen/carbon ratio of the fuel must be input to this analyzer. Additionally, another emission analyzer (MEXA-584L, Horiba) was installed in the intake system to measure the CO2 concentration and thus calculate the EGR ratio. K-type thermocouples were installed on the engine to measure the temperature of the intake gas, exhaust gas, cylinder head, and lubricating oil, as shown in Figure 2.

A piezoelectric pressure transducer (Kistler 6051B, Winterthur, Switzerland) coupled to a charge amplifier (Kistler 5018A) was used to record in-cylinder gas pressure. A shaft encoder (BEI H25, Thousand Oaks, CA, USA) was used to detect the crank angle (CA). The pressure signal was transmitted to a data acquisition system (IndiCom 619, AVL, Graz, Austria) for every 1 ◦CA of 100 continuous cycles. The number of cylinder pressure data of 100 cycles is good enough for the combustion analysis. The pressure data were used to analyze engine combustion parameters, such as the in-cylinder gas temperature, coefficient of variation (COV), and HRR.

The engine control was replaced by a controller (MotoHawk ECU 555-80, Woodward, CO, USA) which controlled the spark timing (for SI engine starting), gasoline injector, DME flow control valve, and EGR control valve.

Temperature is an important factor in HCCI engines [35]. Reference [36] has shown that, for stable HCCI operation in a small engine the cylinder head and oil temperatures should be maintained above 120 ◦C and 65 ◦C, respectively. The intake charge temperature of all tests in this research ranges from 21.2 ◦C to 23.2 ◦C, its mean value is 22.4 ◦C, and the standard deviation is 0.6 ◦C. The range of intake temperature is not large, so the influence of intake temperature on combustion can be neglected.

The engine was started using the original ignition and fuel systems. When the cylinder head and oil temperatures reached 120 ◦C and 65 ◦C, respectively, the controller switched the engine operation to HCCI mode by interrupting the ignition system. Simultaneously, the dual-fuel of DME and gasoline was injected and the throttle was opened further. The throttle was used only for SI operation of the original engine. For HCCI operation, it was fully opened. The fuel supply was adjusted for stable HCCI operation; the ratio of these two fuels was therefore not fixed. The engine was tested at 2000–4000 rpm and with a wide-open throttle. Most of the analyses were focused on 2000 rpm because its operating load range was large. The DME fuel was supplied continuously; its injection pressure was adjusted before engine test for stable combustion and it was found to be 1.5 kg/cm2. The gasoline was injected at a pressure of 2.5 kg/cm2, which was the same as the original engine. The gasoline fuel injection pulse width was adjusted to set the fuel flow rate. The EGR ratio was adjusted from 0% to 30%.

The engine speed, brake torque, DME flow rate, gasoline flow rate, intake air mass flow rate, exhaust emissions, and EGR ratio were measured and recorded during the engine tests. The EGR ratio was calculated using the CO2 percentages measured in the intake and exhaust systems, as formulated in Equation (1) [37]:

$$EGR = \frac{[CO\_2]\_{in}}{[CO\_2]\_{exh}} \times 100\% \tag{1}$$

where [*CO*2] is the *CO*2 concentration.

The maximum error of the experimental data was calculated with an engine test at 2000 rpm and repeated five times using the Kline and McClintock method [38]. The results are listed in Table 4.

**Table 4.** Maximum error in the experimental results of the engine test.

#### *2.2. Calculation of Combustion Parameters*

Combustion parameters, such as IMEP, COV, combustion efficiency, cylinder gas temperature, HRR, and MFB, are used to study the combustion characteristics of an engine.

IMEP is defined as the work divided by the engine displacement volume, which can be calculated during compression and expansion strokes without pumping work. Because this study focused on combustion effect, the elimination of intake and exhaust strokes is better for analyzing. This is expressed as:

$$\begin{aligned} \stackrel{180 \text{CAD}}{\int} P dV\\ IMEP = \frac{-180 \text{CAD}}{V\_d} \end{aligned} \tag{2}$$

where *CAD* denotes CA degrees.

The *COV* is expressed as:

$$COV = \frac{IMEP\_{std}}{IMEP\_{avg}}'\tag{3}$$

The combustion efficiency *ηc* can be defined as the fraction of carbon emitted as *CO*2 in relation to the total carbon emitted (*CO*2, *CO*, HC and PM). The modified combustion efficiency is defined as the ratio between *CO*2 and *CO*2 plus *CO*. This study used the modified combustion efficiency for convenient calculation. This is expressed as:

$$\eta\_{\varepsilon} = \frac{[\text{CO}\_{2}]}{[\text{CO}\_{2}] + [\text{CO}]} \,\prime \tag{4}$$

where square bracket [ ] represents exhaust species concentration as a percentage.

The in-cylinder gas temperature is obtained using the state equation of ideal gas. The HRR equation can be derived from the first law of thermodynamics and is thus expressed as:

$$\frac{dQ\_{lr}}{d\theta} = \frac{\gamma}{\gamma - 1} p \frac{dV}{d\theta} + \frac{1}{\gamma - 1} V \frac{dp}{d\theta} + \frac{dQ\_{lt}}{d\theta},\tag{5}$$

where *dQhr*/*dθ* is the HRR, *dQht*/*dθ* is the heat transfer rate between cylinder gas and the wall. The specific heat ratio *γ* is a function of temperature, which can be obtained using the empirical equation presented in [39].

The heat transfer rate *dQht*/*dθ* is expressed as:

$$\frac{dQ\_{ht}}{d\theta} = hA\left(T\_{\S} - T\_w\right),\tag{6}$$

where *h* is the heat transfer coefficient, as shown in Equation (7):

$$h = S\_t \rho\_{\mathcal{S}} c\_p(0.5 \text{C}\_m) \text{ and } \tag{7}$$

$$S\_1 = 0.718 \exp(-0.145 \mathcal{C}\_{\text{m}}),\tag{8}$$

where *St* is the Stanton number and *Cm* the average piston speed. Equations (7) and (8) were proposed by [33] for small engines.

The MFB at any CA is calculated from HRR and is expressed as:

$$MFB = \frac{\int \left(\frac{dQ\_{\rm hr}}{d\theta}\right) d\theta}{m\_f Q\_{LHV}}.\tag{9}$$

#### **3. Results and Discussion**

The experimental results are presented in terms of the performances, efficiencies, exhaust emissions, and combustion characteristics of the target engine. The data in Sections 3.1–3.4 are all at 2000 rpm of engine speed.

#### *3.1. Engine Performances*

The BMEP, IMEP, and brake-specific fuel consumption (BSFC) are discussed in this section. HCCI engines are always run with a lean mixture. Generally, the output of a HCCI engine increases with a richer mixture but is limited by engine knocking or MRPR [40]. The EGR is a kind of diluent in air-fuel mixture which suppresses the combustion rate. Previous research [41] has reported that autoignition timing is delayed and burn duration prolonged by applying EGR in a HCCI engine, which suppresses engine knocking. The test results for engine output with various EGR ratios are shown in Figure 4.

The BMEP and IMEP indicate the work output per cycle divided by the engine displacement volume. As shown in Figure 4, the engine output increases when the air–fuel equivalence ratio (lambda, λ) decreases. Each curve in Figure 4 indicates the operational range of each EGR ratio. The maximum engine output on each curve increases in line with the EGR. However, the EGR of 30% seems too high for this HCCI system because the operational range is very small and the engine output cannot be extended. In this study, an EGR of 25% yields the best engine output. When the EGR was increased from 0% to 25%, the maximum BMEP increased by 65.8% from 2.93 to 4.86 bar. In Figure 4, the first and last point of each EGR line were determined by the limit of stable operation, which is COV < 10% for low load and MRPR < 6 bar/deg for high load.

**Figure 4.** Effect of lambda on engine output with various EGR ratios: (**a**) BMEP; (**b**) IMEP.

Gasoline and DME were used in the dual-fuel system. DME was selected as the main fuel and was supplied at an almost constant flow rate. Increasing the gasoline mass ratio in the dual-fuel causes λ to decrease, as shown in Figure 5a. The BMEP increases in line with the amount of gasoline, as shown in Figure 5b. At an EGR of 25%, BMEP increases by 77.4% from 2.74 to 4.86 bar when the gasoline ratio rises from 0.13 to 0.35.

**Figure 5.** Effect of gasoline mass ratio in the dual-fuel on: (**a**) Lambda; (**b**) BMEP.

The fuel consumed contains gasoline and DME; therefore, an equivalent fuel mass was used. The equivalent fuel flow rate was calculated based on the heating value of the fuels, as shown in Equation (10), and the *BSFC* was calculated using Equation (11):

$$
\dot{m}\_c = \frac{\dot{m}\_{DME} Q\_{LHV\ DME} + \dot{m}\_{\text{gasoline}} Q\_{LHV}}{Q\_{LHV}} \text{ and} \tag{10}
$$

$$BSFC = \frac{\dot{m}\_e}{bhp'} \tag{11}$$

where . *me* is the fuel flow rate equivalent to gasoline, . *mDME* is DME flow rate, . *mgasoline* is gasoline flow rate, *QLHV DME* is the low heating value of DME, *QLHV* is the low heating value of gasoline, and bhp is the brake horsepower.

Figure 6 shows the BSFC of all test points. Most of the operation conditions have a BSFC ranging from 230 to 260 g/kW·h. The BSFC of a HCCI engine is much lower than that of a conventional motorcycle engine, which is approximately 350 g/kW·h [42]. When λ is >2 or BMEP is <3 bar, the BSFC clearly increases. This might be caused by the low engine output and low combustion efficiency. A BSFC of ≤250 g/kW·h was achieved under conditions of λ < 2 and an EGR of 25%.

**Figure 6.** Brake specific fuel consumption with respect to: (**a**) Lambda; (**b**) BMEP.

#### *3.2. Engine Efficiencies*

The combustion efficiency, *ηc*, is the ratio of the fuel mass burned and the fuel mass delivered into the engine. It can be calculated using Equation (4). The fuel efficiency *ηf* is the ratio of the power developed by the engine to the rate of fuel energy in the engine. It is calculated using Equation (12):

$$\eta\_f = \frac{bhp}{\dot{m}\_f Q\_{LHV}} = \frac{3600(s/h)}{BSFC \text{(g/kW} \cdot h) \cdot Q\_{LHV}(kl/g)},\tag{12}$$

where bhp is the brake horsepower, . *mf* is the fuel flow rate, and *QLHV* the lower heating value of fuel. The brake thermal efficiency *ηth* is calculated using Equation (13):

$$
\eta\_{th} = \frac{bhp}{\eta\_c \dot{m}\_f Q\_{LHV}} = \frac{\eta\_f}{\eta\_c} \,\tag{13}
$$

Figure 7 shows that the combustion efficiency decreases with an increase in λ or decrease in engine load (BMEP) because a high λ or low BMEP leads to a leaner mixture and incomplete combustion. In Figure 7a, the effect of λ on combustion efficiency for each EGR ratio is separated each other and does not coincide together. However, the effect of BMEP on combustion efficiency is evident (Figure 7b), irrespective of the amount of EGR. The combustion efficiency was close to 100% when BMEP > 3 bar.

**Figure 7.** Combustion efficiency with respect to: (**a**) Lambda; (**b**) BMEP.

Figure 8 shows that the fuel efficiency decreases with increasing λ or decreasing engine load (BMEP) because high λ or low BMEP leads to incomplete combustion. When λ is >2 or BMEP is <3 bar, the fuel efficiency clearly drops.

**Figure 8.** Fuel efficiency with respect to: (**a**) Lambda; (**b**) BMEP.

The brake thermal efficiency is shown in Figure 9. Most of the brake thermal efficiencies are clustered within the range of 30%–35%. They have no clear relationship with λ or engine load (BMEP). The thermal efficiency is not very high because the target engine is small in size and heat loss is relatively high [33]. Furthermore, the HCCI engines do not require as much cooling as a conventional engine due to the low combustion temperature of HCCI. Therefore, the modification of a cooling system to reduce cooling capacity will be better for HCCI operation and will improve thermal efficiency.

**Figure 9.** Brake thermal efficiency with respect to: (**a**) Lambda; (**b**) BMEP.

#### *3.3. Exhaust Emissions*

Emission data for NO is lacking because the concentration was too low to be measured during the engine test. In general, when the combustion temperature is lower than 1800 K, much less NO will be emitted [43]. The maximum cylinder temperatures of all test points in this study were less than 1600 K, which is much lower than 1800 K, so the NO emission was close to zero ppm.

The exhaust emissions of CO and HC are depicted through the brake-specific emissions as BSCO and BSHC. The BSCO increases with increasing λ or decreasing exhaust temperature, as shown in Figure 10. The leaner mixture causes higher CO emission in the HCCI engine (Figure 10a), which is the opposite of that in a conventional SI engine. The results obtained are in agreement with previous research [44,45]. This is because the combustion temperature of HCCI is lower and, again, will decrease with increasing λ. Sjöberg and Dec [44] reported that CO oxidation does not reach completion with a peak temperature below 1500 K because the OH level becomes too low.

**Figure 10.** Brake specific CO emission with respect to: (**a**) Lambda; (**b**) Exhaust temperature.

In Figure 10a, the effect of λ on BSCO for each EGR ratio is separated each other and does not coincide together. However, the effect of exhaust temperature on BSCO is evident (Figure 10b), irrespective of the amount of EGR. The BSCO is close to 0 g/kW·h when the exhaust temperature is > 550 K.

The BSHC increases with increasing λ or decreasing exhaust temperature, as shown in Figure 11. A higher λ indicates a leaner mixture and leads to a lower combustion temperature, which causes an increase in HC emissions. Both BSCO and BSHC show good correlation with the exhaust temperature, but they do not correlate with the maximum cylinder gas temperature (the Figure for which is not included here). Thus, the high temperature of the exhaust gas results in the oxidation of CO and HC.

**Figure 11.** Brake specific HC emission with respect to: (**a**) Lambda; (**b**) Exhaust temperature.

The exhaust temperature increases in line with BMEP. The curve fitting of Figure 12a shows a good correlation between exhaust temperature and BMEP, with an R2 of 0.9713. The oxygen concentration in the exhaust gas increases with λ as a quadratic equation, the R2 of which is 0.9899 (Figure 12b).

**Figure 12.** Curve fitting of exhaust properties: (**a**) Exhaust gas temperature versus BMEP; (**b**) Oxygen concentration in exhaust gas versus lambda.

#### *3.4. Combustion Characteristics*

The combustion characteristics were calculated from cylinder pressure values. The cylinder pressure and temperature under a condition of 25% EGR and 20% EGR with various λ values are illustrated in Figure 13. A lower λ means a large amount of gasoline is added, which delays the combustion phase and slows the onset of maximum cylinder pressure (Figure 13a,c). The cylinder temperature is very low (Figure 13b,d) compared with an SI engine because HCCI involves low-temperature combustion [12].

**Figure 13.** Cylinder pressure and temperature: (**a**) Cylinder pressure with EGR 25%; (**b**) Cylinder temperature with EGR 25%; (**c**) Cylinder pressure with EGR 20%; (**d**) Cylinder temperature with EGR 20%.

HRR and MFB under a condition of 25% EGR and 20% EGR with various λ values are illustrated in Figure 14. It is clear that the addition of more gasoline (less λ) causes higher HRR values and delays HRR and MFB. The beginning of the increase in HRR indicates the start of combustion. Figure 14a shows that the start of combustion is delayed by adding gasoline.

**Figure 14.** *Cont*.

**Figure 14.** Heat release rate and mass fraction burned: (**a**) Heat release rate with EGR 25%; (**b**) Mass fraction burned with EGR 25%; (**c**) Heat release rate with EGR 20%; (**d**) Mass fraction burned with EGR 20%.

Figure 15 shows that COV decreases and MRPR increases with an increase in BMEP. The COV increases rapidly as BMEP is <3 bar (Figure 15a). In general, a low load causes unstable combustion and a high load causes a high rate of pressure rise or knocking [11,26]. Almost all the MRPR values are less than 6 bar/deg in Figure 15b, which is a criterion for avoiding engine damage [40]. This is because the addition of gasoline and EGR delays the ignition and suppresses the combustion reaction.

**Figure 15.** The coefficient of variation of IMEP (COV) and the maximum rate of pressure rise (MRPR) with respect to: (**a**) COV; (**b**) MRPR.

In Figure 16, the maximum HRR1 (first-stage HRR) increases with increasing λ, whereas maximum HRR2 (second-stage HRR) decreases with increasing λ. The first stage of combustion (Figure 16a) expresses the ignition property, whereas the second stage of combustion (Figure 16b) expresses the combustion property. The first stage of combustion is the result of cool-combustion chemistry and negative temperature coefficient behavior [46]. In the first stage of combustion, the reaction rate decreases after the temperature reaches a certain value. Therefore, the maximum HRR1 depends on λ only (Figure 16a). By contrast, the maximum HRR2 depends on both the λ value and EGR ratio (Figure 16b).

**Figure 16.** Effects of lambda on maximum heat release rate: (**a**) First stage combustion, HRR1; (**b**) Second stage combustion, HRR2.

CA10, CA50, and CA90 are the CA values when MFB is 10%, 50%, and 90%, respectively. These indicate the combustion phase. Both CA10 and CA50 are delayed when λ is lower, as shown in Figure 17. This is caused by the effect of dual-fuel on combustion. A richer mixture (i.e., less λ) has much more gasoline (Figure 5a), which delays combustion. CA50 is a combination of first- and second-stage combustion, whereas CA10 represents first-stage combustion only. Therefore, the correlation between CA10 and λ (Figure 17a) is clearer than that of CA50 (Figure 17b).

**Figure 17.** Effects of lambda on combustion phasing: (**a**) CA10; (**b**) CA50.

Figure 18a shows that the burn duration (period between CA10 and CA90) increases with an increase in λ, because a leaner mixture causes slower combustion. Figure 18b shows that the burn duration increases rapidly when BMEP is <3 bar.

**Figure 18.** Burn duration with respect to: (**a**) Lambda; (**b**) BMEP.

The combustion efficiency is close to 100% at a burn duration of <10 deg CA (Figure 19a), which equates to a BMEP of >3 bar (Figure 18b). The burn duration indicates the combustion rate, whereas CA10 indicates the start of combustion. Therefore, the influence of burn duration on the combustion efficiency is more obvious (Figure 19a) than in CA10 (Figure 19b).

**Figure 19.** Combustion efficiency with respect to: (**a**) burn duration; (**b**) CA10.

#### *3.5. Comparison between HCCI and SI*

The engine speeds operated with HCCI contained 2000, 2600, 3000, 3300, 3500, and 4000 rpm. The experimental results of HCCI are compared with that of original SI engine as shown in Figures 20 and 21.

Figure 20a shows that the operating range of HCCI engine is much smaller than that of the original SI engine, so the proposed HCCI engine cannot be used in a conventional motorcycle. However, it could be used as a range extender for an electric motorcycle. The BSFC of HCCI engine is much better than that of original SI engine as shown in Figure 20b. The lower BSFC of HCCI engine might be caused by several reasons: (1) lower heat transfer loss from cylinder gas to the wall, (2) short combustion duration, and (3) lean mixture. The combustion duration of HCCI engine is much shorter than that of conventional SI engine. When the load increases (i.e., λ decreases in Figure 14a), the combustion phasing has to be retarded to avoid engine knocking as shown in Figure 14a.

**Figure 20.** Comparison of engine characteristics between proposed HCCI and original SI engine: (**a**) BMEP; (**b**) BSFC.

The CO emission of HCCI engine is smaller than that of original SI engine (Figure 21a) due to the lean combustion. However, the HC emission is higher (Figure 21b) because of the low combustion temperature.

**Figure 21.** Comparison of exhaust emissions between proposed HCCI and original SI engine: (**a**) BSCO; (**b**) BSHC.

#### **4. Conclusions**

The proposed HCCI engine was operated with DME-gasoline dual-fuel in a conventional motorcycle engine. The engine test results and combustion analysis led to the following conclusions:

To further improve the thermal efficiency of the proposed HCCI engine, future studies should modify the cooling system to reduce the cooling capacity.

**Author Contributions:** Y.-Y.W. designed the experiments; C.-T.J. performed the experiments and data reduction; Y.-Y.W. and C.-T.J. analyzed the data; Y.-Y.W. wrote the paper.

**Funding:** This research was funded by Ministry of Science and Technology (MOST, Taipei, Taiwan), grant number MOST 106-3113-E-027-002-CC2.X.

**Acknowledgments:** The authors would like to thank the Ministry of Science and Technology (MOST, Taipei, Taiwan) for their financial support. This manuscript was edited by Wallace Academic Editing.

**Conflicts of Interest:** The authors declare no conflict of interest. The founding sponsors had no role in the design of the study; in the collection, analyses, or interpretation of data; in the writing of the manuscript, and in the decision to publish the results.

#### **Nomenclature**

#### **References**

© 2019 by the authors. Licensee MDPI, Basel, Switzerland. This article is an open access article distributed under the terms and conditions of the Creative Commons Attribution (CC BY) license (http://creativecommons.org/licenses/by/4.0/).

### **Comparative Analysis of the Combustion Stability of Diesel-Methanol and Diesel-Ethanol in a Dual Fuel Engine**

#### **Arkadiusz Jamrozik \*, Wojciech Tutak, Renata Gnatowska and Łukasz Nowak**

Faculty of Mechanical Engineering and Computer Science, Czestochowa University of Technology, 42-201 Czestochowa, Poland<EMAIL_ADDRESS>(W.T<EMAIL_ADDRESS>(R.G<EMAIL_ADDRESS>(Ł.N.)

**\*** Correspondence<EMAIL_ADDRESS>
Received: 1 February 2019; Accepted: 6 March 2019; Published: 13 March 2019

**Abstract:** The co-combustion of diesel with alcohol fuels in a compression ignition dual fuel engine is one of the ways of using alternative fuels to power combustion engines. Scientific explorations in this respect should not only concern the combustion process in one engine cycle, which is most often not representative for a longer engine life, but should also include an analysis of multiple cycles, which would allow for indicating reliable parameters of engine operation and its stability. This paper presents experimental examinations of a CI engine with a dual fuel system, in which co-combustion was performed for diesel and two alcohol fuels (methanol and ethanol) with energy contents of 20%, 30%, 40% and 50%. The research included the analysis of the combustion process and the analysis of cycle-by-cycle variation of the 200 subsequent engine operation cycles. It was shown that the presence and increase in the share of methanol and ethanol used for co-combustion with diesel fuel causes an increase in ignition delay and increases the heat release rate and maximum combustion pressure values. A larger ignition delay is observed for co-combustion with methanol. Based on changes in the coefficient of variation of the indicated mean effective pressure (COVIMEP) and the function of probability density of the indicated mean effective pressure (f(IMEP)), prepared for a series of engine operation cycles, it can be stated that the increase in the percentage of alcohol fuel used for co-combustion with diesel fuel does not impair combustion stability. For the highest percentage of alcohol fuel (50%), the co-combustion of diesel with methanol shows a better stability.

**Keywords:** co-combustion; dual fuel; combustion stability; coefficient of variation of IMEP; probability density of IMEP

#### **1. Introduction**

Compression ignition engines (CI) are commonly used in transport, industrial machines and in the agricultural economy due to their long life time durability. The CI engines are a subject of great criticism due to their emissions output. The biggest problem in these engines is the simultaneous reduction of NOx and soot emissions [1,2]. Emissions of soot and NO are opposed to each other. In paper [3] authors presented results of an investigation of the emissions of dual fuel CI engine. Authors stated that in all analyzed cases, conditions were achieved in which soot emission starts to increase again with lower NO emissions. Emissions of NOx and soot should be reduced because these are harmful to human health and environment as well [4]. Another motivation for these activities is the European Union Directive 2009/28/EC obliging the use of a 20% share of renewable biofuels in overall transport and diesel fuel consumption by 2020 [5]. One of the reasons for using biofuels is their smaller negative impact on the environment through lower greenhouse gas emissions, while another is in order to develop diversification, which can increase energy independence. Increased consumption of biofuels causes reduced consumer reliance on imported fossil fuels and the depletion of crude oil

reserves [6,7]. Fuelling of compression ignition internal combustion engine with unconventional fuels seems to be an interesting alternative that allows for increasing the use of biofuels in the transport and energy sectors. Biofuels such as alcohols and biodiesel have been proposed as alternatives to fossil fuels for internal combustion engines [8]. Combined with diesel fuel or biodiesel, alcohol fuels may represent a valuable fuel for compression ignition engines. Alternative liquid fuels are used for powering compression ignition engines in two modes. The first one is powering engines by using a diesel-alternative fuel blend and second one is a co-combustion of fuels using dual fuel technology [9–12]. The main disadvantage of using a fuel mixture is the limited share of alcohol due to the phase separation [13,14]. In a dual-fuel engine, additional fuel is supplied to the intake system by low-pressure injection. Such produced air-alcohol mixture enters into cylinders during the inlet valve opening period. The engine in this mode of operation can burn alcohol to its 90% energy share [15,16]. Recently, there has been a growing interest in using organic oxygen compounds in engines of this type, especially alcohols [17]. Some disadvantages of alcohols compared to diesel fuel include a relatively high compression ignition temperature and a tendency for the decreasing of charge in-cylinder temperature due to the high value of heat of vaporization. Furthermore, the heating value of ethanol and methanol is lower compared to diesel. Thus, to provide the same energy of delivered fuel to the engine cycle, higher amounts of alcohol are needed [15,16]. The reduction of exhaust emissions of the compression ignition engine can be obtained by controlling the combustion process and by using alternative fuels [18]. Therefore, the substitution of diesel fuel with alternative fuels such as alcohol fuels represents a good approach to solve the environmental problems. With its properties, alcohol can also have a positive effect on NOx and particulate matter emissions [19]. The diesel fuel consists of various hydrocarbons, the ratio of carbon to hydrogen (C/H) is high, and there is a tendency to form soot under fuel rich combustion conditions. The particulate matter is created in the internal combustion engine (mainly in the self-ignition engine), when combustion causes local or temporary lack of air, and more specifically oxygen. The additional oxygen delivered in the alcohol fuel promotes the burning of coal and reduces the formation of PM. Due to their organic nature, alcohols contain up to 50% oxygen in their structure, allowing for widespread oxidation of fuel, resulting in more complete combustion. One of the alcohols that can be used for co-combustion with diesel is ethanol. Ethanol can be produced from raw materials of plant origin and therefore can be considered a fully renewable fuel [17]. Methanol can be produced from such resources as natural gas, coal and various renewable biomasses. The main disadvantages of using ethanol in diesel engines include a low cetane number and problems with solubility in the petroleum diesel [20]. Among these alternative fuels, methanol also represents a promising alternative fuel [21]. Since methanol has a higher latent heat of vaporization than ethanol, methanol has the potential to reduce NOx due to a cooling effect on the charge, leading to lower in-cylinder temperatures [20]. Another advantage of using alcohol is an oxygen content in methanol structure which reduces exhaust fume emissions. Ethanol or ethanol mixed with diesel leads to the occurrence of phase separation between these fuels. Phase separation occurs each time because diesel fuel and methanol or ethanol are immiscible [22]. Therefore, a better solution is to use alcohols in dual fuel mode. In the available literature, there are many publications about the co-combustion of ethanol or methanol with diesel fuel in internal combustion engine. The papers [14,23] presented the results of examinations of compression ignition engine powered by blends of diesel fuel with various alcohols. The authors demonstrated that diesel-ethanol blends could be used in the compression ignition engines with quite a large fraction. Using ethanol fuel in a blend with diesel fuel is supplied additional oxygen for the combustion, which can cause improvements in the overall combustion process. Another statement was that the combustion process of alcohol-diesel fuels blend in IC diesel engine takes place in a shorter time than in the diesel engine fuelled by pure diesel fuel. With an increase of the ethanol fuel fraction, the ignition delay increased. With the increase in the proportion of ethanol, the combustion duration decreases [14,20,24]. The paper [25] presented results of investigation of the combustion process in compression ignition engine powered by a diesel-methanol blend. They stated that up to a 30% methanol fraction, the combustion process

occurred as it would in engines powered by pure diesel. A further increase in the alcohol content led to a significant decline in cylinder pressure and the abnormal combustion process. For mixtures up to 40% of methanol fraction, it did not exceed the value of COVIMEP equal to 10% [25]. In another paper [24] results are presented which show that the use of dual fuel technology gave a satisfactory performance in a standard diesel engine without any modification in the engine. In addition, emissions of carbon monoxide, unburned hydrocarbon and smoke were lower, whereas nitrogen oxides emissions were almost equivalent with diesel. In paper [24], results are presented of varying the ethanol ratio on engine performance and emissions of the dual fuel combustion with various load conditions. The test engine was a single-cylinder diesel engine equipped with two direct injectors operated with constant speed. The authors stated, inter alia, that NOx and PM emissions decreased with increasing ethanol fraction. Additionally, they stated that for low and high loads conditions, the ethanol fraction is limited due to insufficient ignition energy at a low load and at full load due to a high in-cylinder pressure rise. The engine can also burn so-called hydrated ethanol [26]. It seems that this is a good solution because the removal of water from fermented products of ethanol consumes lots of energy. In the paper [27], results are presented of hydrous ethanol combustion in a dual fuel engine, including impacts on the combustion and emissions. The thermal efficiency reduced significantly at the ethanol purity of 60%. The reduction of ethanol purity can reduce NOx emissions but CO and THC emissions increase. In paper [4] results are presented of investigation of cylinder-to-cylinder pressure variation in a diesel engine fuelled with diesel/methanol using dual fuel technology. Research was carried out on the 4-cylinder turbocharged direct injection diesel engine. In the experiment, methanol was injected into the inlet duct by two low-pressure injectors, and directly mixed with intake air before entering the engine. The diesel fuel was directly injected into cylinder. The experimental results showed that the unevenness degree of engine increases with the increase of methanol percent under various engine loads. The uniqueness of pressure in subsequent cycles increases with the methanol share under a low engine load, and the COVpp (coefficient of variation of peak pressure) curves vary very little under the low methanol percentage under 75% and 100% engine loads, but a significant rise appeared when the methanol substitution percent further increased [4]. Co-combustion of alternative fuels such as alcohols causes lowering in cylinder temperature which then affects the value of the ignition delay due to higher value of heat of evaporation of alcohols. In the paper [28] results are presented of investigation of in-cycle pressure fluctuation intensity under long ignition delay conditions. Authors stated that under low temperature charge conditions, the difference in peak pressure between the average cycle and cycles showed low fluctuation intensity. The average intensity of the pressure fluctuations showed an increase with increasing amounts of the premixed combustion phase. This resulted in higher cycle-to-cycle variations under these conditions [28]. Other works presented comparative analysis of the impact of fuel on the uniqueness of engine cycles. In the paper [29], results are presented of investigation of cycle-to-cycle variations of peak pressure in compression ignition engine powered by diesel and biodiesel fuel. The results show that at a lower load, in-cylinder pressure variations for biodiesel were lower compared to mineral diesel fuel. On the other hand, at medium and high loads, biodiesel dominated the peak n-cylinder pressure variations [29]. The cycle of unrepeatability influences emissions variation as well. In the paper [30], results are presented of assessment of cyclic variations on NO emissions of a direct injection diesel engine. Ono of statements was that the intensity of the pressure fluctuations increases with increasing ignition delay. Additionally, the authors stated that in cases with a significant diffusion combustion phase, cycles with pressure fluctuations showed higher-than average NO concentrations, with the intensity of fluctuation determining the increase in NO. Cycle-to-cycle variations in compression ignition engines are undesirable due to negative impacts on efficiency and higher emission. This undesirable phenomenon causes power output limitations.

This work aims to increase knowledge of influence of type of alternative fuel co-combusted with mineral diesel fuel on operation parameters and cycle-to-cycle variations of the CI engine. The results of experiments of co-combustion of diesel fuel with alcohols in the compression ignition engine working in the system of dual fuel are presented in these studies. This paper presents experimental examinations of a CI engine with dual fuel system, in which co-combustion was performed for diesel and two alcohol fuels (methanol and ethanol) with energy contents of 20%, 30%, 40% and 50%. The research concerned the analysis of the combustion process and the analysis of non-repeatability for 200 subsequent engine operation cycles. In the available literature there is not enough information about the influence of the share of alcohol on the stability of the combustion process. However, a lot of information is provided about the deterioration of the combustion by the share of alcohol. In this work, we showed the effect of alcohol on combustion process stability using indicators such as COVIMEP and the probability density of IMEP.

#### **2. Materials and Methods**

#### *2.1. Cycle-to-Cycle Variability Determination*

During the observation of the course of the indicator pressures in successive engine operation cycles, it is easy to notice differences in the values of maximum pressure and indicated mean effective pressure in subsequent cycles. Cycle variability is an undesirable feature of combustion in internal combustion engines. Cyclic variations in diesel engines are undesirable since they are understood to lead to lower efficiency and higher emissions and to power output limitations [28]. They are caused by varied conditions in the cylinder from one combustion cycle to the other, under nominally identical operating conditions. Due to air pollution and increasingly stringent emission standards around the world, mitigation of the effects of cycle variability can reduce fuel consumption and minimize the emissions of harmful compounds in IC engines. Furthermore, cycle variability limitation allows for better engine optimization, i.e., it is possible to achieve efficiency close to optimal conditions. The main sources of the non-repeatability of engine cycles include:

The assessment of the of non-repeatability degree of the maximum pressure and indicated mean effective pressure of the thermal cycle occurring in the cylinder of the internal combustion engine can be conducted by analysing several dozen or several hundred consecutive engine cycles.

The obtained set of maximum pressures can be divided into classes k, from the interval:

$$\mathbf{k} \in \langle \mathbf{p}\_{\text{max} \, \mathbf{A}'} \, \mathbf{p}\_{\text{max} \, \mathbf{B}} \rangle\_{\prime} \tag{1}$$

where pmax A is the smallest value of maximum pressure in set of cycles, pmax B is the maximum value of the maximum pressure in set of cycles.

The average value of maximum pressure pmax, determined on the set of pressure:

$$\overline{\mathbf{p}}\_{\text{max}} = \frac{1}{\mathbf{N}} \sum\_{i=1}^{N} \mathbf{p}\_{\text{max i}} \tag{2}$$

where N is the cycle index, pmax I is maximum pressure in individual cycles.

One of the most popular criteria for the evaluation of the correct operation of combustion engines is non-repeatability of the work cycles. The coefficient of non-repeatability of maximum pressure COVpp, expressed in percentages and calculated as a ratio of standard deviation of the maximal pressure to its mean value from multiple engine operation cycles, was adopted as a measure of non-repeatability [4,31].

The coefficient of variation of peak pressure (COVpp) is defined as:

$$\text{COV}\_{\text{PP}} = \frac{\sigma\_{\text{pmax}}}{\overline{\text{P}\_{\text{max}}}} \cdot 100\%\_{\text{\textdegree}} \tag{3}$$

The standard deviation of peak pressure:

$$
\sigma\_{\text{pmax}} = \sqrt{\frac{1}{N} \sum\_{i=1}^{N} \left( \mathbf{p}\_{\text{max}} \,\mathrm{i} - \overline{\mathbf{p}\_{\text{max}}} \right)^{2}},
\tag{4}
$$

where pmax is the mean value of peak pressure of N cycle pmax i.

The indicated mean effective pressure is evaluated based on the recorded changes in the cylinder pressure and represents one of the indices that characterize operation of combustion engines in terms of the opportunities to ensure high and expected functional performance.

Indicated mean effective pressure for a single engine cycle:

$$\text{IMEP}\_{\text{i}} = \frac{1}{\text{V}\_{\text{d}}} \int\_{0}^{720} \text{pdV}\_{\text{i}} \tag{5}$$

The average value of IMEP:

$$\overline{\text{MEP}} = \frac{1}{\text{N}} \sum\_{i=1}^{\text{N}} \text{IMEP}\_{i\text{\textquotedblleft}} \tag{6}$$

where IMEPi is indicated mean effective pressure in individual cycles.

The uniqueness of IMEP can be determined in the same way as the uniqueness of the maximum pressure value.

$$\text{COV}\_{\text{IMEP}} = \frac{\sigma\_{\text{IMEP}}}{\overline{\text{IMEP}}} \cdot 100\%\_{\text{\textdegree}\_{\text{\textdegree}}} \tag{7}$$

The standard deviation of IMEP:

$$
\sigma\_{\text{IMEP}} = \sqrt{\frac{1}{N} \sum\_{i=1}^{N} \left( \text{IMEP}\_i - \overline{\text{IMEP}} \right)^2},
\tag{8}
$$

The probability density function allows for expressing the probability of obtaining or occurrence of a concrete value of the analyzed parameter. Distribution of probability density of the indicated mean effective pressure represents a repeatability index (index of probability of occurrence) of individual IMEP values obtained for multiple analyzed cycles of the test engine operation. This function can also be used as an indicator to assess the stability of operation of the internal combustion engine. Among other things, it indicates the frequency of occurrence of the most frequently repeated IMEP value, which is close to the mean value, indicating the repeatability of subsequent engine operation cycles. The IMEP probability density function has the profile of a normal distribution (Gaussian distribution), in which the density function is symmetrical in relation to the mean value of the distribution [32].

The probability density of the indicated the mean effective pressure:

$$\text{f(IMEP)} = \frac{1}{\sigma\_{\text{IMEP}}\sqrt{2\pi}} \exp\left(\frac{-\left(\text{IMEP}\_{\text{i}} - \overline{\text{IMEP}}\right)^{2}}{2\sigma\_{\text{IMEP}}^{2}}\right),\tag{9}$$

#### *2.2. Test Stand*

The experimental examinations of the process of co-combustion of diesel with alcohol fuels used the test engine based on a single-cylinder, four-stroke diesel engine 1CA90 manufactured by Andoria (Poland), with compression rate of 17 and engine displacement of 573 cm3. This engine is a stationary two-valve

unit with a vertical arrangement of cylinders and an air cooling system with an axial fan. The engine worked at constant rotational speed of 1500 rpm and constant injection angle of 17◦ crank angle before TDC. The engine was with the one single injection and was running at a steady-state full load. In order to change and ensure the optimal position of the initial injection angle, necessary changes were made in the design, which included changing the design of the camshaft. The engine was equipped in a high-pressure installation of fuel supply with direct injection of diesel fuel to the cylinder. The test engine was modernized and equipped with an additional installation of alcohol fuel supply to the intake manifold. The research stand, with its main component being the engine 1CA90, was equipped in the necessary control and measurement apparatus that allowed e.g., for indication and recording of the necessary parameters of the engine work. Engine indication included measurements and digital recording and analysis of fast-changing pressure in the combustion chamber of the research engine. The engine indication system included Kistler 6061B piezoelectric pressure sensor, charge amplifier Kistler 5011B, crank angle sensor and the module for data acquisition with the A/D converter (Measurement Computing USB-1608HS) (Figure 1). Table 1 presents the most basic technical data for this engine.

**Figure 1.** Diagram of the experimental setup (**a**) and view from the top of the test engine (**b**).

**Table 1.** Main engine parameters.

The test bench consisted of:

data acquisition module, crank angle encoder, resolution of 360 pulses/rev, Measurement Computing USB-1608HS–16 bits resolution of sampling frequency 20 kHz,

#### *2.3. Fuels*

Diesel fuel and technical alcohols (methanol, ethanol), were used in the study. Due to the low value of the cetane number and difficulties with compression ignition, alcoholic fuels, i.e., methanol CH3OH and ethanol C2H5OH, viewed as unconventional engine fuels, cannot be used as individual fuels in CI engines and, therefore, have been used only as fuels for SI engines. Nowadays, although methanol and ethanol belong to strongly oxygenated alcohols (with high content of oxygen in a fuel particle) and are characterized by a lower heating value compared to gasoline or diesel fuel, a continuous increase in interest in these alcohols as potential replacement fuels is being observed compared to conventional fuels, as well as in compression-ignition engines. Mass percentage of oxygen in structure of these alcohols is ca. 50% for methanol and ca. 35% for ethanol. High content of oxygen can have a positive effect on combustion in the engine's combustion chamber since this oxygen is characterized by higher activity compared to the molecular oxygen contained in air. It is conducive to intensification of combustion and contributes to the reduction in exhaust gas smokiness. Similar to ethanol, methanol is characterized by insignificant heating value compared to conventional fuels. The heating value of pure methyl alcohol (19.5 MJ/kg) is only 45%, whereas for ethyl alcohol (26.9 MJ/kg), this means 60% of the heating value of the diesel fuel or gasoline. Both alcohols are characterized by a high heat of vaporization, equal to 1100 kJ/kg for methanol and 840 kJ/kg for ethanol (Table 2). In compression ignition engines, this property reduces mixture temperature and improves the coefficient of cylinder filling, whereas in engines with compression ignition, it contributes to the ignition delay and impacts on the increased rate of pressure rise and can magnify the so-called hard operation of the engine. Hard operation is defined as engine work at a large pressure increase. For diesel engines, a limit of 1 MPa/deg is usually assumed. Alcohol fuels offer opportunities to reduce the use of fossil fuels in CI engines and to increase percentage of biofuels in the transport and energy sectors, where combustion engines are often used. Methyl and ethyl alcohols, due to relatively low costs and easy production, belong to the most popular fuels for combustion engines. Methanol can be manufactured on the industrial scale from broadly available resources, including both non-renewable (natural gas, coal) and renewable (gas generated from biomass or municipal waste) sources. Ethanol is typically obtained during alcohol fermentation of sugars, with sugar sources being mainly cereals (wheat, maize, barley) and potatoes.

Figure 2a shows the two most important features of the analyzed fuels from the standpoint of combustion engine operation. These are the calorific value, which determines engine efficiency, and heat of vaporization, which, in the case of compression ignition engines, impacts on the processes that initiate the ignition of the combustible mixture, causing a change in compression ignition delay and determining engine performance and levels of emissions. Figure 2b shows the molecular composition of diesel, methyl and ethyl alcohol. Hydrogen content in all the fuels is comparable. The main difference between them results from oxygen and carbon content. With its molecular structure, diesel fuel does not contain oxygen, whereas the most oxygenated alcohol (methanol) has a 50% oxygen content (Table 2).

The study examined the processes of co-combustion of diesel with methanol and diesel with ethanol. The percentages of alcohol fuels were 20%, 30%, 40% and 50%. The following symbols were used during the analysis: for pure diesel fuel: D100, for co-combustion of diesel fuel with methanol: DM20, DM30, DM40, DM50 and, for co-combustion of diesel with ethanol: DE20, DE30, DE40, DE50. The numbers in the symbols concern the percentage of alcohol energy dose in total fuel energy (alcohol and diesel fuel) supplied to the engine.

**Table 2.** Fuels properties [33–36].

**Figure 2.** Lower heating value and heat of evaporation of fuels (**a**) and fuels component mass fraction (**b**).

#### **3. Results and Discussion**

#### *3.1. Combustion Characteristics*

The research was based on the process of indicating the engine, which consists of measuring the pressure in the engine cylinder vs. crank angle. The examinations were started from indication of the engine fuelled with pure diesel as a reference. Figure 3 shows the pressure profiles in the cylinder for the engine fuelled by pure diesel and the engine for co-combustion of diesel with methanol and ethanol. The pressure profiles were used to determine the heat release rate (HRR), which characterizes the combustion process in the engine and allows for evaluation of the operation of the test engine.

The presence and the increase in the percentage of methanol and ethanol used for co-combustion with diesel lead to an increase in ignition delay of the charge caused by a higher value of latent heat of vaporization of the alcohol fuel compared to diesel [37]. Figure 3 shows that a greater ignition delay is observed for co-combustion with methanol. The increase in ignition delay leads to the rise in heat release rate and the maximum combustion pressure values. The paper [38] showed that with an increasing methanol substitution ratio in diesel engine with diesel/methanol compound combustion mode, the combustion starting point is delayed, both for the peak heat release rate and the peak cylinder pressure increase. The increase in ignition delay with methanol fraction results in more fuel burnt in the premixed combustion (kinetic combustion phase), which, coupled with the combustion of diesel fuel in a richer methanol–air mixture, leads to the increase of the maximum cylinder pressure and heat release rate. The objective of paper [39] was to investigate the thermal performance, exhaust emissions and combustion behaviour of small capacity compression ignition dual-fuel engine using fumigated ethanol. It was found that with the induction of ethanol in the CI engine, an increased rate of heat release and increased peak pressure occurs, thus enhancing the thermal efficiency. Cycle to cycle variations were very small with the diesel fuel alone. Distinct engine cycle variations as indicated by higher value of COV of IMEP were evident with various ethanol flow rates. Figure 3 was supplemented with the computed values of the indicated mean effective pressure, which show that co-combustion of diesel with alcohol fuel does not cause significant changes in the values of this parameter. Similar values of IMEP were obtained for both pure diesel oil and for all analyzed percentages of alcohol, at the level of ca. 0.66 MPa. Figure 3 presents the value of IMEP measurement error. The IMEP measurement error was ±3.1%.

**Figure 3.** In cylinder pressure and heat release rate (HRR) traces of the engine for co-combustion of diesel fuel with methanol and diesel fuel with ethanol for 20% (**a**), 30% (**b**), 40% (**c**) and 50% (**d**) alcohol share.

In the conventional compression ignition engine, the combustion process is performed in two stages. The first, initial stage, is a period of kinetic combustion, which relates to combustion of the combustible mixture around the jet of injected fuel, whereas the second stage is diffusion combustion which characterizes the process of gradually vaporized fuel combustion. Depending on the conditions of work, the contribution of these stages changes, thus affecting engine performance and emissions. The stage of the kinetic combustion, which characterizes high values of temperature, is conducive to creation of the nitrogen oxides, whereas the stage of the diffusion combustion may be the source of

non-burned hydrocarbons and soot particles. Co-combustion of fuels in the dual fuel engine differs from classical fuel combustion in the conventional engine. However, in most conditions, it is found that the combustion characteristics and heat release rate curves of dual-fuel engines can also be divided into two phases, and its shape is very similar to the typical curves [40]. In the dual fuel engine, the injection of additional fuel to the intake manifold, a homogeneous mixture is formed during the inlet stroke, which is then compressed in the cylinder during the compression stroke. With this prepared mixture, the injection of the dose of the diesel fuel occurs at the end of the compression stoke. This dose evaporates, ignites and leads to the ignition of the homogeneous mixture in the cylinder. Due to the reduced content of diesel fuel in total dose of the fuel supplied to the engine, its time of vaporization and combustion is shorter. This leads to shortening of the diffusion combustion stage and increases in the content of the stage of kinetic combustion. Co-combustion of fuels using the idea of dual fuel engine with two independent fuel supply systems, allows for control of individual stages of the combustion process through control of the moment of compression ignition through injection of the dose that initiates ignition and offers opportunities for smooth changes in the contents of individual fuels, adjusted to varied conditions of engine work.

Furthermore, Figure 4 shows the effect of alcohol energy percentage on the kinetic and diffusive combustion periods during co-combustion of diesel with methanol and ethanol. The division into two combustion phases was made arbitrarily on the basis of HRR analysis. The kinetic phase was defined as the period from the beginning of combustion to the local minimum on the HRR curve before the diffusion phase. The diffusion phase was defined as the period from the end of the kinetic phase to the end of combustion. 10% of the heat release was considered as the start of combustion and 90% of heat release was considered as the end of combustion. As the share of alcohol increased, the kinetic combustion period was extended due to the increased heat release rate and intensification of combustion, whereas the diffusion combustion period was reduced. Studies [40] showed that the increased in the percentage ratio of methanol energy to the total energy, from 0% to about 50%, enlarges the kinetic combustion phase, called the rapid burning duration. It is also shown that the proportion of rapid burning heat release in the whole combustion heat release increases dramatically with the increase of the alcohol energy in total fuel energy (alcohol and diesel fuel) supplied to the engine. In the case of DM50, the kinetic combustion stage accounted for 56% of the entire period of combustion of the load in the engine cylinder and was 31% longer than in the case of pure diesel. In the case of DE50, the kinetic combustion phase was 43% and was 18% longer than the kinetic phase accompanying the combustion of pure diesel.

For high energy contents, methanol, compared to ethyl alcohol, turned out to be a fuel that increased the rate of charge combustion, increasing the contribution of kinetic combustion and decreasing the contribution of diffusion combustion period. For 50% of alcohol, the use of methanol compared to ethanol increased the kinetic combustion period and led to a reduction in the diffusion combustion period by 13%.

**Figure 4.** Kinetic (**a**) and diffusion (**b**) phase of the co-combustion process of diesel oil with methanol and ethanol for the analyzed energy shares of an alcohol fuel.

#### *3.2. Cycle-to-Cycle Variation of the Engine*

Combustion engine operation is characterized by the non-repeatability of subsequent operation cycles caused by the continuous changes in the conditions present in the engine cylinder, both during filling and combustion processes. Instantaneous conditions in the combustion chamber depend, among other things, on the wave phenomena in the intake channels during the suction stroke, instability of the ignition system and the ignition misfiring (SI engines) and the change in ignition delay (CI engines) associated with the heterogeneity of the combustible mixture. Furthermore, the increased non-repeatability of engine operation is affected by the instability of the rotational speed and engine temperature. The profiles of pressure changes in the cylinder obtained during engine indication for several hundred cycles differ from each other and always deviate to a greater or lesser extent from a single profile usually adopted as a representative profile. Figure 5 illustrates in cylinder pressure profiles and heat release rate during combustion of D100 and DM50 for 200 individual test engine operation cycles.

**Figure 5.** In cylinder pressure and heat release rate during combustion of D100 (**a**) and DM50 (**b**) for 200 individual cycles of a test engine.

The indicator diagrams can be used to determine characteristic values that are representative for many individual engine operation cycles, such as maximum combustion pressure or indicated mean effective pressure. Changes in these values can be computed using statistical analysis methods and presented in the form of the coefficient of variation of peak pressure (COVpp) or of the coefficient of variation of the indicated mean effective pressure (COVIMEP). This parameter can be used to evaluate stability of the internal combustion engine related to the non-repeatability of subsequent engine operation cycles and ignition misfiring. In paper [41], the COV of IMEP and pmax were both used to evaluate cyclic variations in multi-cylinder diesel engine fueled with diesel methanol dual fuel. Three different methanol injection positions were chosen to investigate their effect on cylinder-to-cylinder variation. According to literature, the threshold of correct operation of combustion engine, expressed by the maximal value of the coefficient of variation of the indicated mean effective pressure is 2–5% [35,42].

Figure 6 shows non-repeatability of the value of indicated mean effective pressure and maximum pressure, for a conventional diesel engine and a dual fuel engine for co-combustion of diesel with methanol and ethanol. It can be observed that in the case of both alcohols, the increase in their energy content leads to a reduction in the COVIMEP coefficient, which suggests an improvement in engine operation stability. This may result from an increase in the intensification and speed of the combustion process that improved process repeatability due to the presence of active oxygen supplied in the alcohol fuel. The highest oxygen content is observed in methanol, with the lowest COVIMEP = 1.63% values obtained for DM50. In paper [43] COVIMEP coefficient was analyzed for a dual-fuel engine supplied with diesel fuel and methanol, at 60% load. The COVIMEP of the dual-fuel engine remained smaller than 2% in all test conditions. This means the combustion of dual-fuel is very stable without large cycle-by-cycle variations. In paper [44] results are presented of investigation of the diesel/methanol dual-fuel combustion mode as a new combustion mode for NOx and PM reduction from the diesel engine. The aim of this research was to investigate the effect of diesel fuel injection pressure on the performance and emissions characteristics in a 6-cylinder common-rail diesel engine. It showed that the COVIMEP of the dual-fuel combustion mode remains under 2.1% among all the tests. Experimental investigations [31] of the engine with the ethanol-diesel dual-fuel combustion mode showed that, in the dual-fuel mode, more premixed combustion yielded higher levels of COVIMEP than the conventional diesel combustion mode case for higher engine loads. The increase in COVIMEP was proportional to the ethanol energy fraction. In addition, the dual-fuel operation resulted in higher COVpp than the conventional diesel combustion mode at all engine loads except a lower load. Nevertheless, the COVIMEP and COVpp could be controlled between 1.0% and 3.0%. In the case of the analysis of non-repeatability of the maximal combustion pressure, no unambiguous changes in the COVpp values were obtained. A low percentage of alcohol, both in the case of methanol and ethanol, led to a decrease of the COVpp coefficient, which indicated the improved engine stability. After exceeding the level of 20% alcohol, an insignificant increase was observed in non-repeatability of engine operation, eventually reaching the value comparable to D100 for DM50 and DE50. It should be emphasized that both the engine powered by pure diesel and the engine for co-combustion of diesel with alcohol fuels were characterized by stable and even operation, for which the coefficient of non-repeatability was much lower than the acceptable limits for the reciprocating combustion engine (2–5%).

**Figure 6.** Coefficient of variation of peak pressure (**a**) and coefficient of variation of the indicated mean effective pressure (**b**) for a conventional engine and a dual fuel engine co-combustion diesel with ethanol and methanol.

One of the indicators that allow for the assessment of the operation stability of an internal combustion engine is probability density for the occurrence of selected parameters of engine operation. Figure 7 presents the probability density for the occurrence of IMEP in a conventional engine and a dual fuel engine for co-combustion of diesel with 20%, 30%, 40% and 50% of ethanol and methanol. The distribution of probability density of the indicated mean effective pressure, based on normal distribution, represents a repeatability index for individual IMEP values obtained for 200 analyzed cycles of the test engine operation. The mean values of IMEP presented in the drawings show insignificant differences in performance between combustion of pure diesel and co-combustion of diesel with methanol and ethanol. The largest differences were found for the lowest percentage of alcohol fuel, equal to 20% (Figure 7a).

The probability density curves of the indicated mean effective pressure show, among other things, the repeatability of the mean IMEP value, which may be considered a determinant of the stability of operation of the test engine. The probability density for IMEP only during combustion with the lowest percentage of alcohol fuel (20%) was lower compared to combustion of pure diesel (Figure 7a). In the case of higher percentage of both methanol and ethanol, higher values of f(IMEP) were recorded

compared to D100. For the highest percentage of alcoholic fuel (50%) (Figure 7d), this value was 37.5% for methanol and 36.2% for ethanol, respectively.

Analysis of the profile of the probability density function in the case of DM50 for 200 engine operation cycles reveals that the individual IMEP values for the largest number of cycles were close to the mean value of IMEP. This suggests the best repeatability of subsequent cycles of the test engine operation for co-combustion of diesel with 50% energy from methyl alcohol. These findings are consistent with previous results obtained for the analysis of the stability of a dual fuel engine operation based on the changes in the coefficient of variation of the indicated mean effective pressure.

**Figure 7.** The probability density of the indicated mean effective pressure for co-combustion of diesel fuel with methanol and diesel fuel with ethanol for 20% (**a**), 30% (**b**), 40% (**c**) and 50% (**d**) alcohol share.

The correlation between the maximum combustion pressure and the indicated mean effective pressure for a conventional engine and a dual fuel engine for co-combustion of diesel with ethanol and methanol is presented in Figure 8. The figure illustrates the scattering of the pmax and IMEP values obtained for 200 engine cycles and the mean values of these parameters adopted as representative. The analysis of the scattering in subsequent cycles reveals that the increase in the indicated mean effective pressure was associated with the increase in the maximum combustion pressure for each case. Due to the fact that the tests were performed at constant engine load, insignificant changes in IMEP were obtained for each case, which did not differ much from the value of 0.66 MPa. Larger differences are observed in pmax values. Analysis of this parameter shows that for the lowest alcohol percentage (20%), the differences between co-combustion of methanol and ethanol are the greatest. The increase in the alcohol content to over 20% yielded pmax values similar for DM and DE while the differences were increased compared to D100. It is possible to obtain the expected IMEP in an engine for co-combustion of diesel with alcohol, even though the engine has a wider range of pmax changes compared to the combustion of pure diesel.