Quantitative analysis on the effects of compression ratio and operating parameters on the thermodynamic performance of spark ignition liquefied methane gas engine at lean burn mode

Fuel xxx (xxxx) xxxx

Contents lists available at ScienceDirect

Fuel journal homepage: www.elsevier.com/locate/fuel

Full Length Article

Quantitative analysis on the effects of compression ratio and operating parameters on the thermodynamic performance of spark ignition liquefied methane gas engine at lean burn mode Jianqin Fua, Jun Shua, Feng Zhoua, Lianhua Zhonga, Jingping Liua, Banglin Dengb, a b

⁎

State Key Laboratory of Advanced Design and Manufacturing for Vehicle Body, Hunan University, Changsha 410082, China College of Mechatronics and Control Engineering, Shenzhen University, Shenzhen 518060, China

A R T I C LE I N FO

A B S T R A C T

Keywords: Methane Natural gas engine Compression ratio Thermodynamic cycle Thermal efficiency

In this study, a thermodynamic analysis for the liquefied methane gas (LMG) engine with the variation of compression ratio (CR) was conducted through theoretical and experimental investigations. Firstly, the equations for thermodynamic cycle efficiency were further corrected based on the previous studies, in which the losses due to heat release rate (HRR), exhaust valve opening (EVO) timing, specific heat ratio, incomplete combustion and heat transfer were considered. Then, the sweeping test of CR was conducted on an LMG engine. On this basis, the thermodynamic cycle process was studied and various kinds of energy losses were analyzed. The results show that the improvement of indicated thermal efficiency by increasing CR mainly depends on engine operating conditions, the maximum of which occurs at high load and is close to the theoretical value (4.2 percent points). The actual cycle efficiency of LMG engine is mainly influenced by the specific heat ratio of medium gas, followed by the heat transfer loss and the effective expansion ratio (EER) loss. Compared with combustion duration, the combustion phase plays a much more important role in EER loss. All these have provided theoretical basis and direction for the improvement of actual thermal efficiency of LMG engine.

1. Introduction Due to the energy crisis and the increasingly serious environmental pollution, the energy conservation and emission reduction of automotive engine makes good sense and becomes one of the research hotspots in global [1–3]. It is well known that the heat-work conversion efficiency of internal combustion (IC) engine increases with the rising of CR within a certain range [4,5]. Especially, the improvement potential of thermal efficiency by increasing the CR is more obvious as the original CR is lower. However, for the spark ignition (SI) premixed combustion engines, such as gasoline engines and most of the natural gas (NG) engines, due to the restraint of knocking limit under low speed and high load conditions [6,7], the designed CR is usually not very high (e.g., around 11). Therefore, if the fuel’s octane number is improved, the CR of SI engine can also be increased accordingly, which is a potential way to improve the fuel economy of SI engine.

As a common alternative fuel for automobile and gas turbine, NG is dominated by methane, but also contains lower proportions of ethane, propane, butane and alkanes with higher carbon molecules [8,9]. Usually, the composition of NG from different production place is quite different [5,10], and the higher proportion of high-carbon alkanes in NG signifies the greater probability of detonation (knocking combustion). With the composition of NG in different regions considered, NG engines can only be designed according to the physical properties of NG with the lowest octane value. Therefore, the CR of NG engines is mostly designed to around 11 at present [11]. If the NG can be purified to near pure methane (or if the butane and higher carbon molecules in the NG are reduced below a certain amount), the CR of the NG engine can be increased, e.g., 13 or above. Of course, how high the final CR can be is also closely related to the design of the combustion system and the organization of the working process under the operating state of engine. Due to the large differences between different engines, it can only be

Abbreviations: BDC, bottom dead center; BMEP, brake mean effective pressure; CNG, compressed natural gas; CR, compression ratio; EE, expansion efficiency; EEE, effective expansion efficiency; EER, effective expansion ratio; EOC, end of combustion; ER, expansion ratio; EVO, exhaust valve opening; HRR, heat release rate; IC, internal combustion; IMEP, indicated mean effective pressure; LMG, liquefied methane gas; LNG, liquefied natural gas; NG, natural gas; RCCI, reactivity controlled compression ignition; SI, spark ignition; SOC, starting of combustion; TDC, top dead center; VVT, variable valve timing ⁎ Corresponding author. E-mail address: [email protected] (B. Deng). https://doi.org/10.1016/j.fuel.2019.116692 Received 19 October 2019; Received in revised form 14 November 2019; Accepted 17 November 2019 0016-2361/ © 2019 Elsevier Ltd. All rights reserved.

Please cite this article as: Jianqin Fu, et al., Fuel, https://doi.org/10.1016/j.fuel.2019.116692

Fuel xxx (xxxx) xxxx

J. Fu, et al.

concepts were proposed to evaluate the heat-work conversion process, and the quantitative relations between indicated thermal efficiency and various influencing parameters were built. In this way, the improvement potential of thermal efficiency in LMG engine by increasing CR was revealed, various energy loss mechanisms in the actual cycle and the correlations between energy losses and operating parameters of LMG engine were obtained. All these have provided theoretical basis and data support to improve the heat-work conversion efficiency of NG engine.

verified through experiments. In order to improve the energy-saving potential of NG engine, lots of researches have paid attentions on its thermal efficiency and thermodynamic performance [12–14]. Liu et al. [15] investigated a compressed natural gas (CNG) engine applied with radical induced ignition method which can operate with higher CR and better thermal efficiency. Zhou et al. [16] carried out the thermal balance test of liquefied natural gas (LNG) engine and analyzed the influence factors of various forms of energy flow. The results showed that the thermal efficiency of LNG engine can be further improved by optimizing the thermodynamics process. Poorghasemi et al. [17] explored the NG/diesel reactivity controlled compression ignition (RCCI) engine by Converge CFD model, and found that the gross indicated efficiency of engine can be improved by 5% through reducing the first injection pressure. Papagiannakis et al. [18] studied the combustion characteristics of a dual fuel compression ignition engine operated with pilot diesel fuel and NG. The results implied that the total heat release of dual fuel operation is obviously higher than the conventional diesel operation. Yousefi et al. [19] investigated the influence of swirl ratio on the combustion performance and emissions of a NG/diesel dual fuel engine, and claimed that swirl motion provides better mixture preparation and higher thermodynamics efficiency but leads to higher heat losses under very high swirl ratio (> 1.5). On the other hand, the research of hydrogen addition for NG engine has also attracted a great deal of attention to improve the combustion and thermodynamics performance of NG engines [20-22], since the hydrogen presence on lean methane combustion brings lots of the beneficial effects. In particular, hydrogen addition to methane increases the laminar burning velocity [23] and the resistance of the flame to strain-induced extinction [24,25], and also enlarges the operating window of stable combustion [26]. Chen et al. [27] claimed that the increase in CR and hydrogen content can attain higher EER by improving thermodynamic process, and consequently achieve higher indicated efficiency. Wang et al. [28] investigated a direct injection engine operating on various fractions of NG-hydrogen blends. They found that the brake effective thermal efficiency increases with the increase of hydrogen fraction, and suggested that the optimum hydrogen volumetric fraction in blends is around 20% to get a balance of performance and emissions. Huang et al. [29] analyzed the sensitivity of the time intervals between the end of fuel injection and ignition timing to the thermal efficiency and emissions, and found that the NG-hydrogen engine got much lower THC, CO and CO2 emissions and higher thermal efficiency compared with normal NG engine. Dimopoulos et al. [30] optimized a state of the art passenger car NG engine for hydrogen NG mixtures and high EGR rates under part load, and achieved significant efficiency increase with substantially lower engine-out NOx through the optimal combinations of spark timing and EGR rate. Although researchers have carried out a lot of investigations on the thermodynamic performance of NG engines from various aspects and proposed lots of methods to improve the thermal efficiency of NG engines, the existing researches are mainly based on the optimization of combustion process and combustion organization mode [8,31,32]. Up to now, there is a lack of in-depth and systematic analysis on the thermodynamic process of NG engines from the perspective of thermodynamic cycle, and the in-depth understanding on various energy loss mechanisms during the cycle process is also very rare. As a result, the optimization direction and improvement potential of the thermal efficiency of NG engines are still not clear. On the other hand, although some researches have been carried out on the improvement of NG engine thermal efficiency through increasing CR, the quantitative analysis on the effect of CR on NG engine thermal efficiency is still scarce, let alone the difference of improvement potential between the theoretical and actual values. To solve the above issues, in this study the effects of CR on the thermodynamic performance of LMG engine were quantitatively discussed, the thermodynamic cycle process was disintegrated and various kinds of energy losses were analyzed. Some new

2. Engine bench test To improve the thermal efficiency of NG engine, an improved method was proposed in our previous study, in which the LNG is first purified into liquefied methane and then fueled to the high CR engine [5]. To verify the energy-saving potential of this method, the sweeping test of CR was conducted on a turbocharged LNG engine under universal characteristic conditions, and the effects of CR on the in-cylinder combustion and heat-work conversion process were investigated. A six-cylinder, single point injection, spark ignition LNG engine was used for this experimental investigation, which operates in lean burn mode at all loads (the excess air ratio ranges between 1.2 and 1.5). The main parameters of the tested LNG engine are listed in Table 1. Based on this engine, the bench test was performed with the sweeping of CR under various operating conditions. It should be noted that in the test process, the engine fueled with purified LNG, which can be regarded as liquefied methane since the methane volume fraction reaches up to 99.68% [5]. As it can be seen from Table 1, the original CR of this tested LNG engine is 11.6, and the sweeping tests of CR were carried out at four kinds of CR (12.6, 13.6, 14.6 and 15.6). The schematic diagram of test principle is displayed in Fig. 1. As illustrated, the liquefied methane firstly evaporates in the heat exchanger, and then the gaseous methane is injected into the intake port (the methane mass is measured by the gas flow meter). An in-cylinder pressure sensor was installed at the IC engine cylinder head, the signal of which was firstly transformed into voltage signal and then transferred to the AVL combustion analyzer. In this way, the HRR and combustion characteristic parameters were obtained. Also, the evaluation parameters of cycle performance, such as indicated mean effective pressure (IMEP) and indicated thermal efficiency, can also be obtained by analyzing the tested in-cylinder pressure [33]. The combustion efficiency can be obtained based on either the HRR or the tested results of exhaust gas analyzer (gas composition and concentration). The experimental scenario of sweeping test of CR for LMG engine is exhibited in Fig. 2, and the specifications of main measuring instruments are listed in Table 2. It is worth mentioning that all the tests were conducted under the condition without knock occurrence.

Table 1 Basic parameters of the tested methane engine.

2

Item

Content

Engine type Displacement (L) Bore (mm) Stroke (mm) Compression ratio (-) Aspiration mode Fuel injection mode Sparking mode Rated power (kW)/Speed (rpm) Maximum torque (N·m)/Speed (rpm)

In-line six-cylinder, four-stroke 9.726 126 130 11.6–15.6 Turbocharging, intake inter-cooling Port injection Spark ignition 250 /2200 1350 /1400

Fuel xxx (xxxx) xxxx

J. Fu, et al.

Aftertreatment system

Air filter

Fresh air LMG Mixing (lean mixture) Exhaust

Flow sensor

Lambda analyser Turbocharger

Shut-off valve

Heat extransfer

Electromagnetic valve

Injector

Gas flow meter Pressure sensor

Throttle valve Pressure sensor

Lean mixture

LMG

Fig. 1. Schematic diagram of test principle. Table 2 Specifications of main measuring instruments. Instruments

Type

Precision

Dynamometer

Xiangyi electronic dynamometer AVL-INDISET ADVANCED PLUS Kistler D14FR-5DD2B TOCEIL CMF025 Chengduyikong S2 ETAS Lambda Meter 4CA1.01 Pt100

Torque: ± 0.2%F.S; Speed: ± 5rpm /

K type thermocouple

± 0.18 °C

HORIBA MEXA-584L

CO: 0.01% vol HC: 1 ppm vol CO2: 0.02% vol O2: 0.01% vol NO: 1 ppm vol

Combustion analyzer Cylinder pressure sensor NG flow meter Hydrogen flow meter Lambda meter Crank angle calculator Intake temperature sensor Exhaust temperature sensor Exhaust gas analyzer

Fig. 2. Experimental scenario of sweeping test of CR for LMG engine.

3. Theoretical analysis of IC engine cycle efficiency 3.1. Ideal cycle efficiency For the SI NG engine, its combustion mode is much similar to that of gasoline engine. Thus, the ideal thermodynamic cycle of SI NG engine can be regarded as Otto cycle. According to the thermodynamics theory, the thermal efficiency of ideal Otto cycle (or abbreviated as ideal cycle efficiency) in SI engine can be calculated as [4]:

ηide = 1 −

1 εcγ − 1

< = ± 0.6%F.S ± 0.35% ± 0.1% ± 0.01 0.5°CA ± 0.15 °C

where ηide is the ideal cycle efficiency of SI engine; εc is the geometric CR of SI engine; γ is the specific heat ratio of medium gas. Based on Formula (1), the ideal cycle efficiency of SI engine with the change of CR is obtained and shown in Fig. 3. For the original CR of 11.6, the ideal cycle efficiency of SI engine is 62.5%. When the CR changes from 11.6 to 12.6, 13.6, 14.6 and 15.6, the ideal cycle

(1) 3

Fuel xxx (xxxx) xxxx

J. Fu, et al.

called as constant-volume combustion; (2) The exhaust valve opens at the bottom dead center (BDC) so as to make the actual expansion ratio (ER) of medium gas equal to the geometric CR of IC engine; (3) The medium gas is treated as ideal gas, thus the specific heat ratio of medium gas can be taken as a constant; (4) The energy losses due to the gas exchange process, heat transfer and incomplete combustion are ignored. However, when it comes to the actual cycle of IC engine, Formula (1) should be further revised with the above issues considered. Strictly speaking, since the combustion process requires a time interval (which is called as combustion duration), Formula (1) is only suitable for the small part of fuel–air mixture which burns at the TDC. For the fuel–air mixture burns at the other locations, Formula (1) is no longer applicable. To solve these issues, the thermodynamic cycle processes of IC engine are dispersed into an infinite number of infinitesimal parts, and the differential form of expansion work (dW ) due to the heat released at the crank angle dϕ can be expressed as [34]

Fig. 3. Influences of CR on ideal cycle efficiency.

dW =

dQE ⎛ 1 ⎞ ·⎜1 − ⎟ · dϕ dϕ ⎝ ε (ϕ)κ (ϕ) − 1 ⎠

where

dQE dϕ

(2)

is the instantaneous HRR; κ (ϕ) is the instantaneous specific

heat ratio of medium gas at crank angle ϕ , which can be obtained according to the gas compositions and temperature; ε (ϕ) is the instantaneous expansion ratio of medium gas, which is defined as the ratio of cylinder volume at the EVO timing to that at crank angle ϕ

ε (ϕ) = VEVO/ V (ϕ)

(3)

where VEVO is the cylinder volume at the EVO timing; V (ϕ) is the cylinder volume at crank angle ϕ , which can be calculated through the following formula:

V (ϕ) =

Fig. 4. Specific heat ratios of N2, O2, CO2, H2O vs. temperature (pressure = 1 bar).

1 1 πD 2 S ⎡ π π ⎛1 + ⎞ − cos ⎛ · · ϕ⎞ − 1 − λ2 ·sin2 ⎛ · ϕ⎞ ⎤ + V0 4 2⎢ 180 180 λ λ ⎝ ⎠ ⎝ ⎠ ⎝ ⎠⎥ ⎣ ⎦ (4)

efficiency increases by 1.2, 2.3, 3.3 and 4.2 percent points (absolute value), respectively. As is well known, the actual thermal efficiency of IC engine is far less than the ideal cycle efficiency due to the effects of various energy losses in the cycle process. Now, there comes a question that what is the actual improvement of thermal efficiency in LMG engine due to the increase of CR. Whether is it also affected by various energy losses during the cycle process?

where D is the cylinder bore; S is the cylinder stock; λ is the ratio of connecting rod and crank; V0 is the cylinder clearance volume. Based on Formula (2), the ideal cycle efficiency can be calculated according to the following formula when issues (1), (2) and (3) are treated as practical processes: φ

ηcor _1 3.2. Analysis of actual thermodynamic cycle

Wi = = ΔQfuel

EOC ∫φSOC

dQE dϕ

(1 −

1 ε (φ)κ (φ) − 1

) dφ

Δmfuel ·Hu

(5)

where ηcor _1 is the corrected cycle efficiency with the losses of heat release process (or HRR), EVO timing, specific heat ratio and incomplete combustion considered; Wi is the indicated work in each cycle; ΔQfuel is the fuel energy in each cycle; Δmfuel is the mass of fuel injected in each cycle; Hu is the low heating value of fuel; ϕSOC and ϕEOC are the crank angle at the starting of combustion (SOC) and the end of combustion (EOC), respectively. Through introducing the normalized HRR, Formula (5) can be rewritten as

As it can be seen from Formula (1), the ideal cycle efficiency of SI engine is only related to the geometric CR of IC engine and specific heat ratio of medium gas. For the ideal gas, the specific heat ratio is a constant and only depends on the gas compositions. However, for the actual gas, it is related to not only the gas compositions but also the temperature. Fig. 4 shows the correlations between the specific heat ratios of N2, O2, CO2, H2O (which are the compositions of combustion gas in cylinder) and temperature. As illustrated, the specific heat ratios of all the compositions monotonically decrease with the temperature increasing, especially in CO2 and H2O the downtrend is more obvious. It indicates that when the combustion process occurs, the instantaneous heat-work conversion efficiency will be reduced due to the decrease of specific heat ratio of medium gas. Thus, the variation of specific heat ratio is one of the significant factors for the deviation of actual cycle efficiency from the ideal cycle efficiency. Actually, Formula (1) is based on several assumptions, including: (1) The fuel–air mixture burns completely in an instant at the top dead center (TDC) with the combustion duration ignored, which is also

ηcor _1 =

∫φ

φEOC

SOC

1 ⎞ HRR (φ)·⎜⎛1 − ⎟ dφ ε (φ)κ (φ) − 1 ⎠ ⎝

(6)

where HRR (ϕ) is the normalized HRR. From Formula (6), it can be seen that the corrected cycle efficiency is related to the normalized HRR, instantaneous expansion ratio and specific heat ratio of medium gas. Actually, in the previous studies [34,35], the concept of EER was proposed to evaluate the combined effects of HRR and EVO timing on the heat-work conversion process of IC engine, which is defined as follows. 4

Fuel xxx (xxxx) xxxx

J. Fu, et al.

(a) 1000 rpm

(b) 1600 rpm

(c) 1000 rpm

(d) 1600 rpm

(e) 1000 rpm, 10 bar

(f) 1600 rpm, 10 bar

Fig. 5. Influences of CR on indicated thermal efficiency and cycle efficiency. φ

EER =

EOC HRR (φ)·ε (φ)·dφ ∫φSOC

φEOC φSOC

∫

HRR (φ)·dϕ

ηcor _2 =

(7)

As illustrated, EER has considered the combined effects of heat release process (or HRR), combustion efficiency and EVO timing on the thermodynamic cycle, thus it is an important parameter to evaluate the heat-work conversion performance of IC engine. In the real process, the combustion of fuel is usually incomplete, and thus the accumulated normalized HRR is lower than 1.0, as shown in the denominator of Formula (7). As mentioned above, Formulas (5) and (6) have already considered the effect of incomplete combustion (or combustion loss) on the thermodynamic cycle. To predict the heat-work conversion efficiency of the heat released by combustion, the effect of incomplete combustion (or combustion loss) is removed, and thus Formula (6) is revised as

(

φEOC HRR (φ)· 1 − ∫φSOC φEOC φSOC

∫

1 ε (φ)κ (φ) − 1

HRR (φ) dφ

) dφ (8)

where ηcor _2 is the corrected cycle efficiency with the losses of heat release process (or HRR), EVO timing and specific heat ratio considered, which is also called as effective expansion efficiency (EEE). As a matter of fact, it indicates the ratio of ideal cycle expansion work to the heat released through combustion. Obviously, the difference between Formula (6) and Formula (8) gives the effect of combustion loss on the cycle efficiency. Another important factor for the difference between ideal cycle efficiency and actual cycle efficiency of IC engine is the heat transfer loss. To evaluate the effect of in-cylinder heat transfer loss on the actual cycle efficiency of IC engine, the concept of adiabatic efficiency is introduced, as shown below. 5

Fuel xxx (xxxx) xxxx

J. Fu, et al.

low speed condition and the other is high speed condition. The indicated thermal efficiency is calculated via the following formula:

ηind =

ΔQloss ΔQfuel

(9)

where ΔQloss is the accumulated heat transfer loss from the SOC to the end of expansion process. Finally, the actual cycle efficiency of IC engine can be obtained by considering the combined effects of various energy losses, including the losses due to heat release process (or abbreviated as EER loss), heat transfer loss and combustion loss, as shown below.

ηact =

(

ϕEOC HRR (φ)· 1 − ∫ϕSOC φEOC φSOC

∫

1 ε (φ)κ (φ) − 1

HRR (φ) dφ

)

(11)

where ηind is the indicated thermal efficiency, Vh is the cylinder displacement. As it can be observed from Fig. 5(a) and (b), the indicated thermal efficiency of LMG engine increases slightly with the increase of CR, but the increase rate is not obvious. In contrast, the influence of engine speed and load is obviously greater than that of CR. Taking the CR = 15.6 as an example, when the brake mean effective pressure (BMEP) increases from 2 bar to 14 bar at 1600 rpm, the indicated thermal efficiency of LMG engine increases from 29.3% to 50.9% (increase rate reaches 21.6 percent points). However, the variation range of indicated thermal efficiency is within 4 percent points in all the operating conditions due to the change of CR (from 12.6 to 15.6), the upper limit value of which is close to the theoretical value mentioned in Fig. 3. Therefore, it should give priority to choose a suitable operating condition to pursue a high indicated thermal efficiency for the engine, and then the higher CR can be considered. Fig. 5(c) and (d) show the high-pressure cycle efficiency of LMG engine, which is obtained based on the tested in-cylinder pressure (from −180 oCA to 180 o CA).

Fig. 6. Specific heat ratio of in-cylinder mixture in LMG engine at different speeds.

ηadi = 1 -

IMEP·Vh Δmfuel ·Hu

ηind,HP =

IMEPHP ·Vh Δm fuel ·Hu

(12)

o

180 CA

IMEPHP =

∫−180oCA pdV Vh

(13)

where ηind,HP is the high-pressure cycle efficiency. In general, the variation trend of high-pressure cycle efficiency is similar to that of indicated thermal efficiency in this LMG engine. However, differences are observed in the change rate especially at low load since the pumping loss is very large at low load but small at high load [36,37]. Compared with the indicated thermal efficiency, the high-pressure cycle efficiency is less sensitive to the IC engine load. This is because the influence of gas exchange process (or pumping loss) is removed from the highpressure cycle. Therefore, the high-pressure cycle efficiency is more useful to analyze the in-cylinder heat-work conversion process. Nevertheless, the high-pressure cycle efficiency is still sensitive to the IC engine speed. Meanwhile, Fig. 5(e) and (f) clearly show the highpressure cycle efficiency of LMG engine under two typical operating conditions (1000 rpm and 1600 rpm, 10 bar BMEP), which represents the biggest improvement of cycle efficiency through increasing the CR. At the fixed load of 10 bar BMEP, as the CR increases from 11.6 to 15.6, the high-pressure cycle efficiency of LMG engine increases by 4.2 percent points at 1000 rpm but 3.9 percent points at 1600 rpm, both of

dφ ·ηadi ·ηcom (10)

where ηcom is the combustion efficiency. 4. Results and discussions 4.1. Influence of CR on thermodynamic cycle efficiency of LMG engine In this section, the influence of CR on indicated thermal efficiency and high-pressure cycle efficiency is discussed. It should be noted that both the two kinds of efficiency are obtained based on the tested incylinder pressure, and the heat transfer effect in cylinder has already been considered through the Woschni model. Fig. 5(a) and (b) show the tested results of indicated thermal efficiency of LMG engine at 1000 rpm and 1600 rpm, respectively. Obviously, one represents the

Fig. 7. Influences of ideal gas and real gas on the cycle efficiency. 6

Fuel xxx (xxxx) xxxx

J. Fu, et al.

(a) CR=12.6, 1600 rpm, 3bar

(b) CR=12.6, 1600 rpm, 7bar

(c) CR=12.6, 1600 rpm, 11bar

(d) CR=12.6, 1600 rpm, 15bar

Fig. 8. Nondimensional HRR and cycle efficiencies.

(a) 1000 rpm

(b) 1400 rpm Fig. 9. Influences of CR on EER.

mean value of the specific heat ratio of in-cylinder mixture gas in a working cycle (see Fig. 6). As it can be seen from Fig. 7(a), the cycle efficiency of the real gas decreases by about 7.5 percent points compared with that of the ideal gas, which indicates the important effect of specific heat ratio on the heat-work conversion efficiency in this LMG engine. It can also be found that the difference of cycle efficiency in the two kinds of medium gas has little correlation with the CR, which almost keeps unchanged regardless of the change of CR. Fig. 7(b) compares the calculated results of cycle efficiency in both ideal gas and real gas and the tested results of high-pressure cycle efficiency (based on the tested in-cylinder pressure) in this LMG engine. As a matter of fact, the former has already been illustrated in Fig. 7(a), and it is only used to compare with the tested high-pressure cycle efficiency in Fig. 7(b). In general, the tested high-pressure cycle efficiency is about 7 percent points lower than the calculated results in real gas, and the difference

which are close to the theoretical value. As mentioned above, the specific heat ratio is an important influence factor for the heat-work conversion efficiency of IC engine. For the ideal gas with diatomic molecule, the specific heat ratio can be considered to be a constant of 1.4. But for the real gas, it is related to both the gas components and temperature. Fig. 6 shows the changing process of specific heat ratio of in-cylinder mixture gas in this LMG engine. As shown, the specific heat ratio changes dramatically in a cycle especially in the expansion process around TDC due to the rapid increase of mixture gas temperature (see the circle marked area), which has a great effect on the instantaneous heat-work conversion efficiency. Furthermore, Fig. 7(a) and (b) exhibit the influences of ideal gas and real gas on the cycle efficiency of LMG engine. It is worth noting that the results in Fig. 7(a) are calculated by Formula (1). As mentioned above, the specific heat ratio of ideal gas is set to 1.4 while for the real gas, it is the 7

Fuel xxx (xxxx) xxxx

J. Fu, et al.

(c) 1000 rpm

(d) 1400 rpm Fig. 10. Influences of CR on EEE.

(a) 1000 rpm

(b) 1400 rpm

Fig. 11. Correlations between EER and EEE.

the real process. By comparing instantaneous EE with corrected EE, it can be found that there is an obvious decrease after TDC in the corrected EE (see the circle marked area). As a matter of fact, this is due to the decrease of specific heat ratio (as shown in Fig. 6). That is, the corrected EE can directly reflect the effect of instantaneous specific heat ratio on the instantaneous heat-work conversion efficiency. Different from HRR, the effective HRR and corrected HRR have considered the actual result of released heat (or the capacity to do work). That is, the effective HRR means the heat released can be converted into expansion work in theory, because the specific heat ratio in the definition is taken as the constant of ideal gas. Nevertheless, in the corrected HRR the specific heat ratio is treated as the value of real gas. Thus, the corrected HRR is closer to the real process, and its value indicates the heat released can be converted into expansion work in real process. As it can be seen, the peak of corrected HRR in Fig. 8(a) is obviously lower than that in Fig. 8(b)-(d), and this phenomenon can well explain the problem why the cycle efficiency increases obviously at low load but almost keeps unchanged at medium–high load with the engine load (or BMEP) rising. Fig. 9(a) and (b) exhibit the influences of CR on the EER of LMG engine at 1000 rpm and 1400 rpm, respectively. As it can be seen, the EER almost increases linearly with the CR rising. This is because the high CR gives high cylinder temperature, and high temperature brings high flame speed and then short combustion duration [38,39]. This brings the large EER and then high thermal efficiency due to the increased HRR. Compared with the CR, the EER is more sensitive to the engine load, and this is because the load has a more important effect on the cylinder temperature. Taking the case of 1000 rpm and 4 bar BMEP as an example, when the CR increases from 12.6 to 15.6, the EER

gives the energy losses in cycle process. It should be noted that the tested high-pressure cycle efficiency is the maximum value under all the operating conditions (the whole universal characteristic conditions). As a result, the energy losses in Fig. 7(b) correspond to the minimum level in all the operating conditions. To better analyze the in-cylinder heat-work conversion process, some new concepts are proposed based on the previous theoretical analysis, including effective HRR, corrected HRR, instantaneous expansion efficiency (EE) and corrected EE, the mathematical expressions of which are given as follows:

1 ⎞ HRR (φ)Eff . = HRR (φ)·⎜⎛1 − ⎟ ε (φ)κ − 1 ⎠ ⎝

(14)

1 ⎞⎟ HRR (φ)Cor . = HRR (φ)·⎜⎛1 − κ (φ) − 1 ε φ ( ) ⎠ ⎝

(15)

EE (φ)Ins . = 1 −

1 ε (φ)κ − 1

(16)

EE (φ)Cor . = 1 −

1 ε (φ)κ (φ) − 1

(17)

Fig. 8(a)-(d) show the analysis results of in-cylinder heat-work conversion process in LMG engine at 1600 rpm but different loads. According to the above definitions, one can find that the instantaneous EE is only related to the cylinder volume and EVO, which has nothing to do with the operating parameters and engine conditions, and this is the reason why the instantaneous EE is the same under different conditions. For the corrected EE, since the instantaneous specific heat ratio of mixture gas has been considered, it further approaches to 8

Fuel xxx (xxxx) xxxx

J. Fu, et al.

Geometric CR Geometric CR, ideal gas

Geometric CR, real gas

Max. ER, real gas

Max. ER

(a)

(b)

(c) Fig. 12. In-cylinder thermodynamic process parameters (CR = 12.6).

under the universal characteristic conditions are concerned. As it can be seen from Fig. 12(a), the EER is sensitive to the IC engine operating conditions. In general, the EER becomes larger with the increase of IC engine speed and load especially under low speed and low load conditions. When the IC engine speed rises to a certain level, e.g., 2000 rpm, the EER no longer increases as the IC engine speed turns larger. Meanwhile, when the load reaches a certain level, the EER also no longer increases with the load rising. In Fig. 12(a), both the geometric CR and the maximum ER in this LMG engine are displayed, which are 12.6 and 10.96, respectively. Theoretically, the maximum ER can be equal to the geometric CR (12.6). However, due to the advance opening of exhaust valve (which is before the BDC), the maximum expansion ratio occurs at the EVO rather than BDC. For this reason, the maximum expansion ratio is determined by the EVO and it is smaller than the geometric CR. Since the combustion cannot finish at TDC, the EER is lower than the maximum expansion ratio (10.96), and the difference between the maximum expansion ratio and EER is the root cause for the EER loss, which means the loss of expansion work. Accordingly, the EEE of LMG engine under the universal characteristic conditions is depicted in Fig. 12(b). As illustrated, the variation trend of EEE is similar to that of EER, which is in accordance with the above analysis. In Fig. 12(b), both the black line and the green line are obtained according to Formula (1) with the geometric CR of 12.6, while the former is based on ideal gas and the latter is real gas. As a result, the difference between the two lines (about 5 percent points) gives the effect of specific heat ratio on the EEE (or cycle efficiency). In addition, the blue line is obtained according to Formula (1) with the maximum

increases from 8.0 to 9.5 (the increase rate is 1.5). At the CR of 15.6, when the BMEP increases from 4 bar to 8 bar, the EER increases by 2.5, but it only has a very slight increase when the load further increases. On the other hand, the EER is also affected by the IC engine speed, since the high engine speed brings high turbulence which accelerates the combustion. As a result, the HRR increases and it results in the larger EER. Meanwhile, Fig. 10(a) and (b) illustrate the influences of CR on EEE of LMG engine at 1000 rpm and 1400 rpm, respectively. It should be noted that the EEE is only used to evaluate the ideal cycle performance with the heat transfer effect and combustion loss removed. In general, the variation trend of EEE is similar to that of EER in all the operating conditions, which increases with the CR and load rising. As a matter of fact, the EEE is not directly related to the load. In other words, it is directly related to the EER rather than the load. As shown in Fig. 11(a) and (b), there is a linear relation between EER and EEE regardless of the change of IC engine load. Thus it can be seen, the EEE is mainly determined by the EER, while the IC engine operating conditions are the indirect factors. That is, IC engine operating conditions influence the EEE through the effect on the EER. Therefore, in terms of EEE, the EER is an optimization object, and this conclusion is helpful for improving the engine indicated thermal efficiency. In the practical process, one can choose to optimize the heat release process and EVO timing to obtain a high EER. 4.2. Disintegration of high-pressure cycle efficiency Furthermore, the in-cylinder thermodynamic process parameters 9

Fuel xxx (xxxx) xxxx

J. Fu, et al.

Fig. 13. Disintegration of high-pressure cycle efficiency at CR = 12.6. (Case 1: cycle efficiency for ideal gas and geometric CR; Case 2: cycle efficiency for real gas and geometric CR; Case 3: cycle efficiency for real gas and the maximum ER; Case 4: effective expansion efficiency; Case 5: high-pressure cycle efficiency).

Case 3 is based on real gas at the maximum ER). It should be noted that, the difference between Case 2 and Case 4 represents the EER loss, while the difference between Case 4 and Case 5 equals to the heat transfer loss and combustion loss. Since the EVO in this LMG engine is fixed (without variable valve timing, VVT), the maximum ER is a constant (10.96), and this is the reason why the ideal cycle efficiency calculated based on the real gas at the maximum ER is almost a constant under different IC engine operating conditions. However, things are changed when it comes to the EEE and high-pressure cycle efficiency. As illustrated, the high-pressure cycle efficiency and EEE have the similar variation trend, both of which increase with the engine load rising. At medium–low speed and low load, the EEE is very low, so does the high-pressure cycle efficiency (see the circle marked area in Fig. 13(a), (b) and (c)). Thus it can be seen, the variation trend of high-pressure cycle efficiency mainly depends on the EEE. The heat transfer loss and combustion loss account for a large proportion, which means not only the waste of energy, but also the decrease of heat-work conversion efficiency (compared with the ideal cycle efficiency). In general, the percentages of heat transfer loss and combustion loss decrease with the increase of IC engine speed, and this is one of the reasons for the increase of high-pressure cycle efficiency. When the IC engine speed rises to 2000 rpm, the EEE does not seem to have relevance to the load. According to the above analysis, one can deduce that the HRR and combustion timing are reasonable under the circumstances, thus the EER is not influenced even at low load, and this issue will be further discussed in the next section. In a word, these figures not only quantitatively display the energy

ER of 10.96 in real gas. Therefore, the difference between the green line and the blue line (about 2 percent points) gives the effect of EVO (which is a part of EER loss) on the cycle efficiency. Furthermore, the difference between EEE (obtained based on the tested in-cylinder pressure) and the blue line gives the loss due to the HRR (which is also a part of EER loss). Thus, Fig. 12(b) clearly displays various losses in the thermodynamic cycle and also points out the highest goal of cycle efficiency. Fig. 12(c) shows the high-pressure cycle efficiency of this LMG engine under the universal characteristic conditions. In general, the variation trend of high-pressure cycle efficiency is similar to that of EEE. However, differences are observed in the change range and the sensibility to the operating conditions. Different from the EEE, the highpressure cycle efficiency of this LMG engine changes with the speed even at high load. As illustrated above, the difference between EEE and high-pressure cycle efficiency gives the loss of heat transfer during the expansion process. It should be noted that the loss of heat transfer discussed here is different from the absolute heat transfer from the viewpoint of heat balance in cylinder. This is because other kinds of energy loss (e.g., EER loss) may also dissipate in the way of heat transfer. Fig. 13(a)-(d) show the disintegration of heat-work conversion efficiency in this LMG engine at four speeds (1000 rpm, 1200 rpm, 1600 rpm and 2000 rpm). Similar to Fig. 12(b), the ideal cycle efficiencies calculated by Formula (1) are also given (Case 1 and Case 2 are based on ideal gas and real gas at the geometric CR, respectively, while 10

Fuel xxx (xxxx) xxxx

J. Fu, et al.

Fig. 14. The combustion process parameters of LMG engine (CR = 12.6).

distribution (as well as energy loss) during the cycle process, but also points out the potential and direction for improving the cycle efficiency of IC engine.

4.3. Influences of combustion process on cycle efficiency of LMG engine According to the above analysis, the variation trend of high-pressure cycle efficiency is mainly determined by EEE, since the heat transfer loss and combustion loss vary little with the change of IC engine load.

11

Fuel xxx (xxxx) xxxx

J. Fu, et al.

(a) 1000 rpm

(b) 1600 rpm

Fig. 15. The effects of CR on the COV of IMEP in LMG engine.

indicated thermal efficiency and various influencing parameters were built, which are then used to quantitatively study the effects of CR and operating parameters on the thermodynamic performance of spark ignition LMG engine. Through this research, the following conclusions can be arrived.

As analyzed previously, the EEE mainly depends on the EER, thus the influence factor of EER is further discussed. According to Formula (7), the EER is related to the heat release process, so the influences of combustion process on cycle efficiency and EER of LMG engine are concerned. Fig. 14(a)-(e) give the combustion process parameters of LMG engine at the CR of 12.6. As it can be observed from Fig. 14(a), the ignition timing is advanced largely with the engine speed increasing, which is one of the reasons for the longer ignition delay period at high speed, as shown in Fig. 14(b). Moreover, the ignition delay period is very long at low load especially for high speed condition, which is even comparable to the 10–90% combustion duration (see Fig. 14(e)). This is due to the higher residual gas fraction and combustion instability [40-42]. Due to the longer ignition delay period, the SOC is also very late and even occurs after TDC at low load (see Fig. 14(c)). In general, the influence of 50% combustion position on the EER is much greater than that of 10–90% combustion duration. Taking the cases at 1000 rpm and 1200 rpm as an example, when the BMEP is > 8 bar, the 50% combustion position at the two speeds is close (see Fig. 14(d)), while the 10–90% combustion duration is significantly different (see the marked area in Fig. 14(e)). Under the circumstances, both the EER and EEE are similar. However, when the BMEP is relatively low (e.g., below 4 bar), although the 10–90% combustion duration at various speeds is similar, both the EER and EEE are significantly different due to the large difference in 50% combustion position (see the marked area in Fig. 14(d)). In fact, both the EER and EEE deviate significantly from the ideal value at low load due to the too late 50% combustion position. Therefore, to optimize the combustion phase (SOC, 50% combustion position, etc.) is a direct and effective way to improve the heat-work conversion efficiency of engine. Another interesting issue is the effects of CR on the cycle-by-cycle variations in SI NG engine [43]. Fig. 15(a) and (b) show the effects of CR on the coefficient of variation (COV) of IMEP in the LMG engine at the speed of 1000 rpm and 1600 rpm, respectively. As it can be seen, only at very low load (BMEP < 3 bar) the COV of IMEP is sensitive to the CR, which decreases with the CR increases except for the CR of 12.6. At other load (BMEP > 3 bar), it seems that the COV of IMEP has no relation with the CR. On the other hand, the COV of IMEP deceases as the load increases, especially at very low load the downtrend is more obvious. As stated by Zheng et al. [44], the COV of IMEP decreases as the CR increases from 8 to 12, while it is almost unchanged when the CR further increases (> 12). Since the CR in this study ranges from 12.6 to 15.6 (> 12), in general the results are consistent with the study of Zheng et al. [44].

(1) In general, the CR has a less effect on the indicated thermal efficiency than engine speed and load. With the increase of CR, the indicated thermal efficiency of LMG engine increases slightly, and the increase rate depends on engine operating conditions. As the CR changes from 12.6 to 15.6, the variation range of indicated thermal efficiency is within 4 percent points, and the maximum increase is close to the theoretical value (4.2 percent points) and occurs at high load. In contrast, the variation range of indicated thermal efficiency due to the change of operating condition is > 20 percent points. (2) Different from ideal cycle efficiency, the actual cycle efficiency of LMG engine is mainly influenced by the specific heat ratio of medium gas, followed by the heat transfer loss and combustion loss, then the losses due to the HRR and EVO. Under low speed and low load conditions, the loss due to the HRR (EER loss) is relatively high and becomes the major influence factor, and it decreases obviously under other operating conditions. (3) EER is the direct influence factor for EEE, while IC engine operating conditions and combustion parameters are the indirect influence factors. That is, other parameters influence the EEE in the way of EER. The EER almost increases linearly with the CR rising. Nevertheless, the EER is more sensitive to the load compared with the CR. Moreover, the EER is also affected by the engine speed, but the effect of engine speed is only obvious at low load. Compared with combustion duration, the combustion phase plays a more important role in EER loss. (4) There are also many measures to improve the thermal efficiency of LMG engine. Under low load conditions, it is better to optimize the combustion heat release process (e.g., HRR and combustion phase), while in other conditions it should give priority to the reduction of heat transfer loss. Meanwhile, to improve the specific heat ratio of mixture gas in combustion process and optimize the EVO is also an effective measure to improve the thermal efficiency of engine. In the next study, we will continue the research on the improvement of engine thermal efficiency through the last two ways, e.g., low temperature combustion coupling with EVO optimizing, which can reduce the heat transfer loss and increase the EER. Author contributions

5. Conclusions Jianqin Fu analyzed the tested data and wrote this manuscript. Banglin Deng conceived and designed this study. Jingping Liu designed the engine test. Jun Shu, Feng Zhou and Lianhua Zhong performed the

In this study, a new method was proposed to predict the indicated thermal efficiency of IC engine and the quantitative relations between 12

Fuel xxx (xxxx) xxxx

J. Fu, et al.

engine test and collected experimental data.

2001;26:55–70. [22] Nadaleti WC, Przybyla G, Filho PB, Souza SNM, Quadro M, Andreazza R. Methanehydrogen fuel blends for SI engines in Brazilian public transport: Potential supply and environmental issues. Int J Hydrogen Energ 2017;42(17):12615–28. [23] Sarli VD, Benedetto AD. Laminar burning velocity of hydrogen–methane/air premixed flames. Int J Hydrogen Energ 2007;32:637–46. [24] Sarli VD, Benedetto AD, Long EJ, Hargrave GK. Time-Resolved Particle Image Velocimetry of dynamic interactions between hydrogen-enriched methane/air premixed flames and toroidal vortex structures. Int J Hydrogen Energ 2012;37:16201–13. [25] Sarli VD, Benedetto AD. Effects of non-equidiffusion on unsteady propagation of hydrogen-enriched methane/air premixed flames. Int J Hydrogen Energ 2013;38:7510–8. [26] Sarli VD. Stability and emissions of a lean pre-mixed combustor with rich catalytic/ lean-burn pilot. Int J Chem React Eng 2014;12:77–89. [27] Chen Z, Xu BY, Zhang F, Liu JP. Quantitative research on thermodynamic process and efficiency of a LNG heavy-duty engine with high compression ratio and hydrogen enrichment. Appl Therm Eng 2017;125:1103–13. [28] Wang JH, Huang ZH, Fang Y, Liu B, Zeng K, Miao HY, et al. Combustion behaviors of a direct-injection engine operating on various fractions of natural gas-hydrogen blends. Int J Hydrogen Energ 2007;32:3555–64. [29] Huang ZH, Wang JH, Liu B, Zeng K, Yu JR, Jiang DM. Combustion characteristics of a direct-injection engine fueled with natural gas-hydrogen blends under different ignition timings. Fuel 2007;86:381–7. [30] Dimopoulos P, Rechsteiner C, Soltic P, Laemmle C, Boulouchos K. Increase of passenger car engine efficiency with low engine-out emissions using hydrogen–natural gas mixtures: A thermodynamic analysis. Int J Hydrogen Energ 2007;32(14):3073–83. [31] Amirante R, Distaso E, Di Iorio S, Sementa P, Tamburrano P, Vaglieco BM, et al. Effects of natural gas composition on performance and regulated, greenhouse gas and particulate emissions in spark-ignition engines. Energy Convers Manage 2017;143:338–47. [32] Ma FH, Wang Y, Liu HQ, Li Y, Wang JJ, Zhao SL. Experimental study on thermal efficiency and emission characteristics of a lean burn hydrogen enriched natural gas engine. Int J Hydrogen Energ 2007;32:5067–75. [33] Fu JQ, Yuan ZP, Liu JP, Zhou F, Wang SQ, Feng RH. A hybrid method of tests coupled with simulations used to detect the working processes of an automotive engine from cycle to cycle in transient conditions. Proc IMechE Part D: J Automobile Engineering 2017;231(13):1766–79. [34] Yuan ZP, Liu JP, Fu JQ, Liu Q, Wang SQ, Xia Y. Quantitative analysis on the thermodynamics processes of gasoline engine and correction of the control equations for heat-work conversion efficiency. Energy Convers Manage 2017;132:388–99. [35] Liu Q, Liu JP, Fu JQ, Li YY, Luo BJ, Zhan ZS, et al. Comparative study on combustion and thermodynamics performance of gasoline direct injection (GDI) engine under cold start and warm-up NEDC. Energy Convers Manage 2019;181:663–73. [36] Yüksek L, Özener O, Sandalcı T. Cycle-skipping strategies for pumping loss reduction in spark ignition engines: An experimental approach. Energy Convers Manage 2012;64:320–7. [37] Galindo J, Serrano JR, Climent H, Varnier O. Impact of two-stage turbocharging architectures on pumping losses of automotive engines based on an analytical model. Energy Convers Manage 2010;51:1958–69. [38] Zheng JJ, Wang JH, Wang B, Huang ZH. Effect of compression ratio on the performance and combustion of a direct-injection natural gas engine. Proc Inst Mech Eng: Journal of Automobile Engineering 2009;223(1):85–98. [39] Zheng JJ, Huang ZH, Wang JH, Wang B, Ning DZ, Zhang YJ. Effect of compression ratio on cycle-by-cycle variations in a natural gas direct injection engine. Energ Fuel 2009;23(11):5357–66. [40] Zhao D, Li JW. Feedback control of combustion instabilities using a Helmholtz resonator with an oscillating volume. Combust Sci Technol 2012;104:694–716. [41] An YZ, Raman V, Tang QL, Shi H, Sim J, Chang J, et al. Combustion stability study of partially premixed combustion with low-octane fuel at low engine load conditions. Appl Energy 2019;235:56–67. [42] Zhao H, Li GN, Zhao D, Zhang ZG, Sun DK, Yang WM, et al. Experimental study of equivalence ratio and fuel flow rate effects on nonlinear thermoacoustic instability in a swirl combustor. Appl Energy 2017;208:123–31. [43] Hu EJ, Li XT, Meng X, Chen YZ, Cheng Y, Xie YL, et al. Laminar flame speeds and ignition delay times of methane-air mixtures at elevated temperatures and pressures. Fuel 2015;158:1–10. [44] Hu EJ, Huang ZH, He JJ, Zheng JJ, Miao HY. Experimental and numerical study on lean premixed methane-hydrogen-air flames at elevated pressures and temperatures. Int J Hydrogen Energ 2009;34(16):6951–60.

Acknowledgements This research work is jointly sponsored by the National Natural Science Foundation of China (No. 51876056) and the Fundamental Research Funds for the Central Universities. The authors appreciate the reviewers and the editor for their careful reading and many constructive comments and suggestions on improving the manuscript. References [1] Wu H, Nithyanandan K, Zhang JX, Lin YL, Lee TH, Lee CF, et al. Impacts of AcetoneButanol-Ethanol (ABE) ratio on spray and combustion characteristics of ABE-Diesel blends. Appl Energy 2015;149:367–78. [2] An YZ, Jaasim M, Vallinayagam R, Vedharaj S, Im HG, Johansson B. Numerical simulation of combustion and soot under partially premixed combustion of lowoctane gasoline. Fuel 2018;211:420–31. [3] Zhao D. Waste thermal energy harvesting from a convection-driven RijkeeZhao thermo-acoustic-piezo system. Energy Convers Manag 2013;66:87–97. [4] Heywood JB. Internal Combustion Engine Fundamentals. 2nd Edition: New York: McGraw-Hill; 2018. [5] Fu J, Shu J, Zhou F, Liu J, Xu Z, Zeng D. Experimental investigation on the effects of compression ratio on in-cylinder combustion process and performance improvement of liquefied methane engine. Appl Therm Eng 2017;113:1208–18. [6] Selim MYE. Sensitivity of dual fuel engine combustion and knocking limits to gaseous fuel composition. Energy Convers Manage 2004;45:411–25. [7] Abdelaal MM, Rabee BA, Hegab AH. Effect of adding oxygen to the intake air on a dual-fuel engine performance, emissions, and knock tendency. Energy 2013;61:612–20. [8] Mahmood FG, Amir N, Mahdi DD, Hamid RR. Effects of natural gas compositions on CNG (compressed natural gas) reciprocating compressors performance. Energy 2015;90:1152–62. [9] Zhang ZG, Zhao D, Ni SL, Sun YZ, Wang B, Chen Y, et al. Experimental characterizing combustion emissions and thermodynamic properties of a thermoacoustic swirl combustor. Appl Energy 2019;235:463–72. [10] Kakaee AH, Rahnama P, Paykani A. Influence of fuel composition on combustion and emissions characteristics of natural gas/diesel RCCI engine. J Nat Gas Sci Eng 2015;25:58–65. [11] Cho HM, He BQ. Spark ignition natural gas engines–a review. Energy Convers Manage 2007;48(2):608–18. [12] Weaver CS. Natural gas vehicles - a review of the state of the art. SAE Technical Paper 892133 1989. [13] Shu J, Fu J, Liu J, Zhang L, Zhao Z. Experimental and computational study on the effects of injection timing on thermodynamics, combustion and emission characteristics of a natural gas (NG)-diesel dual fuel engine at low speed and low load. Energ Convers Manage 2018;160:426–38. [14] Shu J, Fu J, Liu J, Wang S, Yin Y, Deng B, et al. Influences of excess air coefficient on combustion and emission performance of diesel pilot ignition natural gas engine by coupling computational fluid dynamics with reduced chemical kinetic model. Energ Convers Manage 2019;187:283–96. [15] Liu Y, Dong Y, Yeom JK, Chung SS. An experimental investigation of the engine operating limit and combustion characteristics of the RI-CNG engine. J Mech Sci Technol 2012;26(11):3673–9. [16] Zhou F, Fu JQ, Liu DH, Liu JP, Lee CF, Yin YS. Experimental study on combustion, emissions and thermal balance of high compression ratio engine fueled with liquefied methane gas. Appl Therm Eng 2019;161:114–25. [17] Poorghasemi K, Saray RK, Ansari E, Irdmousa BK, Shahbakhti M, Naber JD. Effect of diesel injection strategies on natural gas/diesel RCCI combustion characteristics in a light duty diesel engine. Appl Energy 2017;199:430–46. [18] Papagiannakis RG, Hountalas DT. Combustion and exhaust emission characteristics of a dual fuel compression ignition engine operated with pilot diesel fuel and natural gas. Energy Convers Manage 2004;45:2971–87. [19] Yousefi A, Guo HS, Birouk M. Effect of swirl ratio on NG/diesel dual-fuel combustion at low to high engine load conditions. Appl Energy 2018;229:375–88. [20] Korb B, Kawauchi S, Wachtmeister G. Influence of hydrogen addition on the operating range, emissions and efficiency in lean burn natural gas engines at high specific loads. Fuel 2016;164:410–8. [21] Bauer CG, Forest TW. Effect of hydrogen addition on the performance of methanefueled vehicles. Part I: effect on S.I. engine performance. Int J Hydrogen. Energ

13