Investigation on steam direct injection in a natural gas engine for fuel savings

Energy 183 (2019) 958e970

Contents lists available at ScienceDirect

Energy journal homepage: www.elsevier.com/locate/energy

Investigation on steam direct injection in a natural gas engine for fuel savings Lifu Li a, Zhongbo Zhang a, * a

School of Mechanical and Automotive Engineering, South China University of Technology, Guangzhou, 510641, China

a r t i c l e i n f o

a b s t r a c t

Article history: Received 15 June 2018 Received in revised form 15 June 2019 Accepted 29 June 2019 Available online 1 July 2019

In order to improve the natural gas engine (NGE) fuel economy, a steam direct injection method has been presented in the present study. In this method, exhaust was employed to heat water to produce superheated steam firstly. After that, at the power stroke, steam was injected into the cylinder directly. The potentials for fuel savings by this method are evaluated. First, ideal thermodynamic cycle of steam injected NGE is analyzed. Then, a baseline NGE is modeled and validated through experiments. After that, based on the simulation model, the effects of different steam injection parameters on the NGE performance are discussed, including steam mass, temperature and injected timing. The results show that the NGE fuel economy is significantly improved with steam direct injection. With optimal steam mass, 3.9 e5.2% reductions of the NGE brake specific fuel consumption (BSFC) are obtained over different speeds, when steam temperature and injected timing are 550 K and 50 deg, respectively. Steam mass and injected timing have great influences on the NGE BSFC. However, steam mass is limited by pinch point temperature difference of the evaporator and exhaust temperature at the evaporator exit. In addition, steam injected timing is restricted by pressure inside the cylinder. © 2019 Elsevier Ltd. All rights reserved.

Keywords: Natural gas engine Fuel savings Steam direct injection Waste heat utilization

1. Introduction Use of alternative fuels is an important method to solve energy crisis and environmental pollution caused by conventional fuel engines [1e4]. As an alternative fuel for the engine, it is generally accepted that natural gas has many promising advantages, such as abundant resources, attractive cost and lower greenhouse gas emission [5e8]. However, with carbon dioxide (CO2) emissions regulations becoming increasingly stringent, it is important to reduce natural gas engine (NGE) fuel consumption. In conventional engines, a significant amount of fuel energy is taken away by exhaust without converting into useful work. Compared with traditional fuel engines, because of the slow flame propagation velocity of natural gas, NGE has long combustion duration [9e12]. The exhaust temperature of NGE is relatively higher than that of traditional fuel engines. As a result, there is a great deal of available energy in the exhaust of NGE that can be used to improve the overall engine efficiency if an effective waste heat utilization method is employed. Various kinds of methods

* Corresponding author. E-mail addresses: pmfl[email protected] (Z. Zhang).

(L.

Li),

https://doi.org/10.1016/j.energy.2019.06.182 0360-5442/© 2019 Elsevier Ltd. All rights reserved.

[email protected]

have been presented to utilize waste heat of exhaust [13], including turbocompounding [14e17], Rankine cycle [18e20], exhaust gas turbine system [21], thermoelectric generation systems [22e25], etc. Steam injection, another useful technology to utilize waste heat of exhaust, has been widely researched in recent years. In such method, exhaust is employed to heat water to produce steam, which is injected into the engine to improve its performance. Therefore, part of thermal energy of exhaust is transferred into steam. With different targets, steam can be injected at different positions, including pre-turbine, intake manifold and cylinder. The main target of steam injected at the pre-turbine is to raise turbine power of turbocharged engines, so that the engine performance could be changed. For example, in studies of Zhu et al. [26,27], the influences of steam injected at the pre-turbine on a diesel engine were studied. It is shown that the engine fuel economy was improved by 3.2% [26]. Besides, with steam injected at the pre-turbine combined with Miller cycle, 2.8% improvement of the engine fuel economy was further obtained [27]. The concepts of steam-assisted turbocharging and steam turbocharging were presented by Fu et al. [28,29]. In the steam-assisted turbocharging approach [28], steam was injected at the pre-turbine and mixed with exhaust firstly. After that, the mixture flowed through the turbine to improve its power. However, for the steam

L. Li, Z. Zhang / Energy 183 (2019) 958e970

Nomenclature

Latin symbols Q heat quantity (J) Q_ heat flux (J/s) m mass (kg) m_ mass flow rate (kg/s) constant volume specific heat (J/(kg
K)) cV cp constant pressure specific heat (J/(kg
K)) P power (kW) u specific internal energy (J/kg) h specific enthalpy (J/kg) p pressure (bar) V cylinder volume (m3) v specific volume (m3/kg) Rg gas constant (J/(kg
K)) a; n Wiebe constant A area (m2) D cylinder diameter (m) w average cylinder gas velocity (m/s) S stroke (m) l ratio of connecting rod and crank length M mole mass (kg/mol) vf =zi coefficient of error transfer zi measured parameter y calculated parameter Dz i absolute error of measured parameter r radius of crank (m) l length of connecting rod (m) B fuel consumption mass (g) be brake specific fuel consumption (g/(kW
h))

turbocharging method [29], instead of flowing through the turbine, exhaust was just employed to heat water to steam, which was used to drive the turbine for turbocharging. It is found that at low engine speed conditions, the two methods both could improve engine performance. Zhao et al. [30,31] investigated the effects of steam injected at the pre-turbine combined with turbocompounding on a diesel engine. It is shown that 6.0e11.2% reductions of the engine fuel consumption were obtained at all speed conditions. Although steam injected at the pre-turbine can improve turbine power, exhaust back pressure is raised as well. As a result, it leads to an increase of engine pumping loss. Besides, instead of directly outputting power, the increase of turbine power is only used for turbocharging. Therefore, the improvement of engine power mainly depends on the matching strategy of the turbocharging system and the engine. Steam injected at the intake manifold aims to reduce nitrogen oxides (NOx) emissions and solve the corrosion and the destruction of the lubricant film caused by liquid water injection firstly [32], since it can replace liquid water to reduce in-cylinder temperature. Moreover, because in-cylinder pressure is also raised by this method, engine performance is improved. For example, Parlak et al. [32,33] presented an intake manifold steam injection method to control NOx emissions. In his method, exhaust served to heat water to saturated water firstly. Then, the saturated water was introduced into the inlet manifold, where it turned to steam and mixed with intake air. It is shown that, employing this method on a diesel engine, 5% reduction of brake specific fuel consumption (BSFC) was obtained [32]. Similarly, the impacts of the steam injected at the

Dy n Ttq

959

absolute error of calculated parameter engine speed (rpm) brake torque (N
m)

Greek symbols h efficiency g specific heat ratio ε compression ratio 2 volume ratio q crank angle (deg) qs start angle of combustion (deg) qb combustion duration (deg) a coefficient u combustion progress variable Acronyms NGE BSFC EVO PCPC

natural gas engine brake specific fuel consumption exhaust valve opening per cylinder per cycle

Subscripts and superscripts e engine m mixture min minimum s stroke or steam u unburned b burned w wall or water 0-8,0w-3w,4s-5s locations in the system with ’ with steam injection without ’ without steam injection

intake manifold method on a diesel engine running with different alternative fuels were investigated by Gonca et al. [34e36]. The results showed that the performance of the engine fueling with different alternative fuels was improved with different degrees. Moreover, the impacts of the steam injected at the intake manifold method combined with other technologies on engine performance were further investigated, such as Miller cycle and turbocharging. In studies of Gonca et al. [37e41], the influences of steam injected at the intake manifold combined with turbocharging and Miller cycle on a diesel engine were studied. It is shown that the engine performance was significantly improved with the combination. The engine power was enhanced by 17% and the thermal efficiency was raised by 11% [41]. Even though engine performance is remarkably improved by the steam injected at the intake manifold method [32e41], the steam is produced by injecting saturated water. Fig. 1 is the pressure-enthalpy diagram of water [42]. According to the figure, because of large latent heat of evaporation, a significant amount of thermal energy is required in the steam generation process. However, the temperature of saturated water is much higher than that in the intake manifold. As a result, in the intake manifold the saturated water cannot turn to steam completely. The corrosion and destruction of the lubricant film are unavoidable in this method. Besides, as shown in Fig. 1, the water enthalpy only increases slightly when water is just heated to saturated water. Thus, the capacity of waste heat recovery is limited in this method. Moreover, for evaporation and expansion of saturated water inside the intake manifold, the air intake flow rate is reduced, which is not

960

L. Li, Z. Zhang / Energy 183 (2019) 958e970

1000

Pressure (bar)

300K

700K

800K 900K

600K

100

500K

10 400K

1 0

500 1000 1500 2000 2500 3000 3500 Fig. 2. Schematic of steam injected NGE.

Enthalpy (kJ/kg) Fig. 1. Pressure-enthalpy diagram.

2. Ideal thermodynamic analysis 2.1. System description Fig. 2 indicates the schematic of steam injected NGE. The working principle of the steam direct injection system can be illustrated as follows. First, room-temperature water from the water tank was pressured by the high-pressure pump and inlet into the condenser. In the condenser, the pressured water was preheated by exhaust. Then, the preheated water flowed into the evaporator and was heated to superheated steam. As shown in Fig. 1, the enthalpy of water is greatly raised in the process. Therefore, in the evaporator, a significant amount of thermal energy of exhaust is recovered by the

2.2. Ideal thermodynamic cycle analysis of steam injected NGE In this subsection, ideal thermodynamic cycle of steam injected NGE is analyzed. Fig. 4 presents the ideal thermodynamic cycle

In-cylinder pressure

good for the improvement of engine performance. To solve those problems of the steam injected at the intake manifold method, a steam direct injection method has been presented and employed on a diesel engine by Zhang et al. [43]. In his method, exhaust was employed to heat water to create superheated steam firstly. Then, at the compression stroke, the steam was injected into the cylinder directly. The results showed that, with optimal steam mass, 8.6e9.9% decreases of BSFC were obtained at various engine speeds. However, since a great deal of steam is introduced into the engine cylinder, it will have complex influences on the in-cylinder combustion [44,45]. Those influences are not taken into consideration in the simulation model in Zhang's study. Therefore, the improvement of engine performance is uncertain in this method unless being proved by experiments. However, for this method, if the superheated steam is introduced into the cylinder at the later stage of the power stroke, due to near the end of the combustion, the steam has little influences on the in-cylinder combustion. In the present paper, in order to enhance the NGE fuel economy, the effects of steam direct injection at the power stroke on NGE performance have been investigated. First, ideal thermodynamic cycle of NGE with steam direct injection at the power stroke is analyzed. Then, a baseline NGE is modeled by using commercial software GT-POWER and validated with experiments. After that, based on the simulation model, the effects of different steam direct injection parameters on the NGE performance are discussed, including steam mass, temperature and injected timing. Finally, the potentials of the steam direct injection method for fuel savings of the NGE are evaluated.

water. Next, at the power stroke, the steam was introduced into the engine cylinder by the steam injector. Fig. 3 is the timing range of steam direct injection. At the early stage of the power stroke, the pressure in the cylinder is ultra-high. Therefore, it is difficult to inject the steam into the engine cylinder under this condition. The steam is only injected at the later stage of the power stroke and before the exhaust valve opening (EVO). After injection, the steam mixed with the original mixture inside the cylinder. Therefore, the thermodynamics cycle of the engine is affected and its performance is also impacted. After that, at the exhaust stroke, the mixture of emissions and steam was exhausted from the engine cylinder and flowed through the evaporator and condenser in turns. In the condenser, the steam in the exhaust was condensed into droplets and flowed into the radiator. At last, in the radiator the condensation continued to be cooled and then flowed back into the water tank. According to the description of the steam direct injection system above, in this method, the liquid water does not exist in the engine cylinder or exhaust pipes. Therefore, the corrosion and destruction of the lubricant film can be completely prevented in this method.

EVO Steam injection timing range

Crank angle Fig. 3. Steam injected timing range.

L. Li, Z. Zhang / Energy 183 (2019) 958e970

with steam injection

without steam injection 5m m-original mixture in the cylinder 5'm s-steam 6's w-water 6m

6' Qin Qw 4

S

W'out

V

Temperature

In-cylinder pressure

with steam injection

without steam injection

5

961

5' 7'

3

7's

4s

2w 1w 0w

Q'out

Win

5's

3w

V

6

S

6'm 7'm

4m

3m

Cylinder volume

Entropy

(a) In-cylinder pressure-cylinder volume diagram

(b) Temperature-entropy diagram

Fig. 4. Ideal thermodynamic cycle of steam injected NGE.

expressed as the in-cylinder pressure-cylinder volume diagram and temperature-entropy diagram, respectively. In the figure, the dotted line is the original ideal thermodynamic cycle of NGE, which is simplified to the Otto cycle. The solid line represents the ideal thermodynamic cycle of NGE with steam direct injection at the power stroke, including 3e4 isentropic compression process of original working fluids, 4e5 constant volume heating process of original working fluids, 5-50 isentropic expansion process of original working fluids, 50 -60 constant volume adiabatic gas charging process, 60 -70 isentropic expansion process of mixture, 70 -3 constant volume heat release process of mixture. According to Fig. 4 (b), since the added energy of 50 -60 process Qw stems from the heat release Q 0 out , the thermal efficiency of the ideal thermodynamic cycle of steam injected NGE h0 e is expressed as Equation (1).

m0 m u0 60 ¼ mm u50 þ ms h5s

(8)

ðmm þ ms Þc0 V T 0 60 ¼ mm cV T50 þ ms cp;s T5s

(9)

Q 0 out  Qw h e ¼1  Qin

After steam injection, the specific heat capacities at constant volume and constant pressure of the mixture inside the cylinder are expressed as Equations (10) and (11), respectively.

0

(1)

εg1 p T5 5 T50 ¼ g1 ¼ T3 2 p4 2

(7)

where 2 ¼ V50 =V5 . Since the superheated steam is introduced into the engine cylinder at point 50 in a very short time, 50 -60 is simplified as a constant volume adiabatic gas charging process. On the basis of the first law of thermodynamics, Equation (8) is acquired. Besides, for the ideal gas with constant specific heat, Equation (8) is also written as Equation (9).

In the ideal thermodynamic cycle, all gases are assumed to ideal gas. Thus, Q 0 out and Qin are represented as Equations (2) and (3), respectively,

c0 V ¼

mm cV þ ms cV;s ms þ mm

(10)


 Q 0 out ¼ m0 m c0 V T3  T 0 70

(2)

c0 p ¼

mm cp þ ms cp;s ms þ mm

(11)

Qin ¼ mm cV ðT5  T4 Þ

(3)

According to Equations (9) and (10), the temperature of the mixture in the cylinder at point 6’ T 0 60 is expressed as Equation (12).

where m0 m ¼ mm þ ms . By substituting Equation (2) and (3) into Equation (1), Equation (4) is obtained.


 m0 m c0 V T 0 70  T3  Qw h0 e ¼ 1  mm cV ðT5  T4 Þ

(4)

For 3e4 process, the temperature of original working fluids inside the cylinder at point 4 T4 is expressed as Equation (5).

T 0 60

 g1 p5 mm cV T3 ε2 p4 þ ms cp;s T5s ¼ mm cV þ ms cV;s

(12)

Besides, for 60 -70 process, the in-cylinder temperature of the mixture at point 7’ T 0 70 is obtained by Equation (13) as follow,

T 0 70 ¼ T 0 60



 0 2g0 1 V60 g 1 ¼ T 0 60 V70 ε

(13)

0

T4 ¼ T3 εg1

(5)

For 4e5 process, the in-cylinder temperature of original working fluids at point 5 T5 is obtained by Equation (6).

T5 ¼ T4

p5 p ¼ T3 εg1 5 p4 p4

(6)

And, for 5-50 process, the in-cylinder temperature of original working fluids at point 5’ T50 is expressed as Equation (7),

where 2 =ε ¼ V60 =V70 and g ¼ c0p =c0V . By substituting Equation (12) to Equation (13), the in-cylinder temperature of the mixture at point 7’ T 0 70 can be also expressed as Equation (14).

T 0 70

 g1 p5 0 mm cV T3 ε2 p4 þ ms cp;s T5s 2g 1 ¼ mm cV þ ms cV;s ε

(14)

By substituting Equations (5), (6), (10) and (14) to Equation (4), the thermal efficiency h0 e is further expressed as Equation (15).

962

L. Li, Z. Zhang / Energy 183 (2019) 958e970

  g0 1 εg1 p 2 5 mm cV T3 þ ms cp;s T5s 2 p4 ε 0   he ¼1 p mm cV T3 εg1 5  1 p4
 mm cV þ ms cV;s T3 þ Qw   þ p mm cV T3 εg1 5  1 p4


 m_ w ðh5s  h2w Þ ¼ m_ m c0 p T 0 80  T 0 100

(19)

The total energy absorbed by the water from exhaust Q_ w is expressed as Equation (20).

(15)

According to Equation (15), it is clear that the thermal efficiency of steam injected NGE is impacted by steam mass, temperature and injected timing.


 Q_ w ¼ m_ w ðh5s  h1w Þ ¼ m_ m c0 p T 0 80  T 0 110

(20)

According to Equation (20), the exhaust temperatures at the evaporator inlet and the condenser outlet are important parameters that affect the energy recovered by water. However, the exhaust temperature at the evaporator inlet T 080 is determined by the engine. Therefore, to recover more energy from exhaust, a low exhaust temperature at the condenser outlet T 0110 should be obtained by the steam generation system.

2.3. Analysis of ideal thermodynamic process of steam generation Fig. 5 is the ideal thermodynamic process of superheated steam generation expressed as the temperature-entropy diagram. In the figure, 0w-1w is the isentropic compression process of water, 1w2w is the water preheated process in the condenser, 2w-5s is the steam generation process in the evaporator, and 80 -110 is the heat exchange process of the mixture (or exhaust) in the exhaust system. As for 0w-1w process, the power consumed by the pump Ppump is calculated by Equation (16).

Ppump ¼ m_ w ðh1w  h0w Þ

Pg ¼ Pe  Ppump

(17)

In 1w-2w process, the pressured water was preheated in the condenser by exhaust. According to the energy conservation principle, the heat transfer process in the condenser is written as Equation (18).


 m_ w ðh2w  h1w Þ ¼ m_ m c0 p T 0 100  T 0 110

(18)

After the water is preheated in the condenser, in 2w-5s process, the preheated water is heated to superheated steam in the evaporator by exhaust. The heat transfer process in the evaporator is expressed as Equation (19).

Temperature

8'

T0

11'

To evaluate the influences of steam direct injection on NGE performance, a thermodynamic cycle model for NGE was developed by using the commercial software GT-POWER. The model is based on a one-dimensional description of the flow inside the intake and exhaust pipes. The intake and exhaust systems are carefully schematized by a network of pipes and junctions. Besides, the combustion in the cylinder is described by a zero-dimensional two-zone model.

(16)

Therefore, with steam injection, the engine gross power Pg can be calculated by Equation (17).

10'

3. Simulation model

9' 3w

3.1. Engine cycle modelling In two-zone combustion model, the in-cylinder combustion is divided into unburned zone and burned zone. Energy conservation equations of unburned and burned zone are expressed as Equations (21) and (22) [46e48], respectively.

dðmu uu Þ dmu dVu dQw;u ¼ hu p þ dq dq dq dq

(21)

dQw;b dðmb ub Þ dmb dV ¼ hb p bþ dq dq dq dq

(22)

During the working process of engine, the blow-by is ignored. Therefore, based on the mass conservation, the total mass inside the cylinder is expressed as Equation (23). By taking derivative, Equation (23) can be further expressed as Equation (24).

m ¼ mu þ mb

(23)

dmu dmb ¼  dq dq

(24)

Besides, based on the volume conservation, the cylinder volume is expressed as Equation (25).

5s 4s

2w 1w 0w Entropy

Fig. 5. Temperature-entropy diagram of superheated steam generating process.

V ¼ Vu þ Vb

(25)

The equations of state of the two zones are expressed as Equations (26) and (27), respectively.

pVu ¼ mu Rg;u Tu

(26)

pVb ¼ mb Rg;b Tb

(27)

The specific internal energy u is obtained by Equation (28). The constant volume specific heat cV and the constant pressure specific heat cp are expressed as Equation (29) and Equation (30), respectively,

L. Li, Z. Zhang / Energy 183 (2019) 958e970

du ¼ cV dT cV ¼

Rg

g1

¼ cp  Rg

  gRg ¼ R a1 þ a2 T þ a3 T 2 þ a4 T 3 þ a5 T 4 cp ¼ g1

(28)

dQw ¼ aAw ðT  Tw Þ dq

(37)

(29)

a ¼ 3:26D0:2 p0:8 T 0:55 w0:8

(38)

(30)

where R is the universal gas constant (J/(mol
K)) and a1 -a5 are the coefficients that can be obtained by least-squares matching with thermodynamic property data from the JANAF tables [49]. In addition, the units of coefficients a1 -a5 in the Equation (30) are different, they are mol/kg, mol/(kg
K), mol/(kg
K2) mol/(kg
K3) and mol/(kg
K4), respectively. According to Equations (21)e(30), the control equations for the two-zone thermodynamic model can be derived, as shown in Equations (31)e(33), including the temperature of the two zones and the pressure in the cylinder. The detailed derivation of the control equations can be found in Refs. [50,51].

  dTu gu  1 dp dQw;u Vu þ ¼ mu gu Rg;u dq dq dq

(31)

   dTb 1 g 1 dp dQw;u  u Vu þ  ¼ mb Rg;b dq dq gu dq   dm dp dV b Rg;b Tb  Rg;u Tu þV þp dq dq dq

(32)

" !# Rg;b Rg;u dp ð1  gb Þgu dmb T  ¼ f ðub  uu Þ  T  dq Vb gu þ gb Vu gb  1 b Rg;b u dq 

dQw;u dQw;b þ dq dq

 þ

gb  gu dQw;u gb dV þp g gu ðgb  1Þ dq gb  1 dq

(33)

From Equations (31)e(33), it is clear that the in-cylinder pressure and temperature are related to the mass of burned fuel, stroke volume, cylinder wall heat lost and specific heat ratio. For the mass of burned fuel mb , it can be expressed as Equation (34). The combustion progress variable u is estimated according to Wiebe function, which is given in Equation (35).

mb ¼ um

(34)

#    q  qs nþ1 u ¼ 1  exp  a

qb

(35)

It should be noted that at the later stage of the power stroke, the combustion inside the cylinder is near the end. Therefore, the incylinder combustion is weakly affected by the steam injection. There is no need to take the impacts of steam on the mass of burned fuel into consideration. The instantaneous cylinder volume V is obtained by Equation (36).

i p h r Vs V ¼ D2 r 1  cos q þ ð1  cos 2 qÞ þ 4l 4 ε1

963

(36)

The cylinder wall heat lost Qw is expressed as Equation (37). The instantaneous heat transfer coefficient a is calculated by Woschni formula [49], which is presented in Equation (38).

For the specific heat ratio, after steam injection, it can be calculated by Equation (39).

g0 m ¼

c0 p mm cp þ ms cp;s ¼ c0 V mm cV þ ms cV;s

(39)

3.2. Validation of NGE model Calibration of the accuracy of the NGE model is carried out by employing the experimental results from a baseline NGE. Table 1 shows the baseline NGE specifications. The engine is a 4-stroke, in-line 6-cylinder, turbocharged, after-cooled, premixed lean burn, spark ignition NGE, which is modified from a conventional diesel engine together with a Woodward electronic control system. The NGE is usually used as a power of city buses. The experiments were conducted based on the NGE test bench, which is illustrated in Fig. 6. The engine was loaded by a 250 kW eddy current dynamometer, which is installed with a SJ1091H Halleffect speed sensor and a PST750Kg in-line torque meter. The fuel consumption mass was measured by a HZF-100 electronic scale. The exhaust temperature at the exit of the turbine (point 5 in Fig. 2) was obtained by a K-type thermocouple temperature sensor. The signals collected from sensors need to be converted from an original analogue form to a digital form. This was achieved by using a FC2020 data acquisition system, which has 16 channels and 1 MHz bandwidth. The engine operating conditions were controlled by a computer. In the experiments, full load operating conditions from engine speed 1000 rpme2000 rpm were conducted. The steadystate average experimental data was acquired during the experiments. Error analysis of parameters is needed to ensure the accuracy of experimental results. Table 2 shows the measurement range and absolute error of different parameters. The error of the calculated parameter is related to the measured one. According to the propagation of error, the absolute error of the calculated parameter can be expressed as Equation (40).

Dy ¼

n

X

vf Dzi ; y ¼ f ðz1 ; z2 ; :::zn Þ

z

i¼1

(40)

i

Since the engine brake power Pe is calculated by Equation (41) and the fuel consumption mass B is converted to BFSC be by Equation (42), they are both calculated parameters.

Table 1 NGE specifications. Parameter

Value

Number of cylinders Compression Bore Stroke Maximum power Maximum torque Emission standard

6 10.5 112 mm 210 mm 200 kW 980 N m Euro III

964

L. Li, Z. Zhang / Energy 183 (2019) 958e970

Exhaust temperature (K)

1000

Fig. 6. NGE test bench.

Table 2 Errors of the measured parameters. Measurement range

Absolute error

Engine torque Engine speed Fuel consumption mass Static temperature

0e1000 N m 0e7500 rpm 0e100 kg 0-980  C

±2.2 N m ±1 rpm 5g ±2.2  C

be ¼

pnTtq 30

 103

1200

1400

1600

1800

2000

timing. By employing the NGE simulation model established in section 3, the influences of the steam direct injection parameters on the NGE performance are analyzed in this section. Besides, the potentials of steam direct injection for fuel savings are also presented. 4.1. The restrictions of steam direct injection parameters

According to subsection 2.2, the thermal efficiency of steam injected NGE is affected by steam mass, temperature and injected

240

Max:1.98% Avg:0.77%

Experiment Simulation

Max:3.47% Avg:1.78%

Experiment Simulation

220

BSFC (g/kW.h)

200

Power (kW)

1000

(42)

4. Results and discussion

150

100

200

180

160

50 1200

800

Fig. 8. Validation of exhaust temperature.

According to Equations (40)e(42), as a result, the maximum absolute errors of engine brake power and BSFC are 1.8 kW and 2.8 g/kW
h, respectively. In order to precisely predict engine performance, the NGE model should be validated in detail. Fig. 7 compares the simulation results of engine power and BSFC with the experimental data. It is shown that the maximum deviations of engine power and BSFC are both within 4%. The average deviations of engine power and BSFC are 0.77% and 1.78%, respectively. Therefore, the engine model can accurately predict the NGE performance. In the present study, exhaust temperature is an important parameter, which indicates the available thermal energy of exhaust. The comparison of model results and experimental data of exhaust temperature is shown in Fig. 8. The average and the maximum deviations of exhaust temperature are 0.82% and 1.55%, respectively. Consequently, an accurate exhaust temperature can be obtained by the NGE model.

1000

900

Engine speed (rpm)

(41)

B  1000 Pe

250

Max:1.55% Avg:0.82%

700

Parameter

Pe ¼

Experiment Simulation

1400

1600

Engine speed (rpm)

(a) Power

1800

2000

1000

1200

1400

1600

Engine speed (rpm)

(b) BSFC

Fig. 7. Validation of engine power and BSFC.

1800

2000

To ensure right use of steam direct injection, the restrictions of different steam direct injection parameters are analyzed in this subsection, including steam injected pressure, temperature and mass. Since steam is injected into the engine cylinder at the power stroke, the minimum steam injected pressure must be higher than the pressure in the cylinder at the injected timing. Moreover, the minimum steam temperature must be higher than the evaporating temperature under the injected pressure, so that the water can turn into steam completely. Therefore, the minimum steam temperature is restricted by steam injected pressure. Fig. 9 shows the minimum steam injected pressure and temperature. It can be found from Fig. 9 (a) that steam injected pressure is raised as steam injected timing advances. This is due to that in-cylinder pressure is higher near the top dead center. Besides, because in-cylinder pressure is lower at low engine speed condition than that at high engine speed condition, steam injected pressure is also raised as engine speed increases. It can be told from Fig. 9 (b) that, the minimum steam temperature has similar changes with the minimum steam injected pressure. Due to the fact that the heat transfer process between exhaust and water mainly occurs in the evaporator, the requirement for heat transferability of the evaporator is large. As a result, in order to ensure that the evaporator has high-efficiency heat exchange capability, pinch point analysis of the evaporator is needed [52]. Pinch point is the position where the minimum temperature difference between exhaust and water sides in the evaporator exists [30,52]. For a given heat flux, with a large pinch point temperature difference, the heat transfer area of the evaporator can be reduced. Therefore, the volume of the evaporator can be smaller. In the research, the minimum pinch point temperature difference of the evaporator is set at 20 K to keep high heat exchange efficiency [43]. As it is shown in Fig. 5, the pinch point in the evaporator may appear at point 2w, 3w or 5s. Therefore, the minimum pinch point temperature difference of the evaporator can be written as

L. Li, Z. Zhang / Energy 183 (2019) 958e970

(a) Minimum steam injected pressure

965

(b) Minimum steam temperature

Fig. 9. The minimum steam injected pressure and temperature.

Equation (43).


0 
0   T 100  T2w ; T 90  T3w ;
DT0min ¼ min T 80  T5s > 20K

(43)

In Equation (43), exhaust temperature at the evaporator exit T 0100 is calculated by Equation (18). To cool steam in the exhaust efficiently, the water temperature in the condenser should not be too high. The water temperature at the exit of the condenser T2w is set at 323.15 K in the research. Thus, the temperature difference at point 2w is acquired. The exhaust temperature at point 9’ T 090 is acquired by Equation (44). Besides, the temperature of saturated water T3w can be obtained by using REFPROP software [42], when steam injected pressure is given. Therefore, the temperature difference at point 3w is obtained.


 m_ w ðh5s  h3w Þ ¼ m_ m c0 p T 0 80  T 0 90

(44)

The exhaust temperature at point 8’ T 080 is obtained by the NGE model. In addition, the steam temperature T5s is a parameter of steam direct injection which is given in the study. Therefore, the temperature difference at point 5s is acquired. In summary, according to Equation (43) and temperatures at each point in the equation, steam mass and temperature are limited by the pinch point temperature difference of the evaporator DTmin .

(a) Minimum pinch point temperature difference of the

Moreover, exhaust temperature at the evaporator exit T 0100 must be higher than the temperature of the condensation point of water. Otherwise, condensation of steam in the exhaust will happen before it flows into the condenser. In the study, the lowest T 0100 is set at 393.15 K to keep steam from condensation in the evaporator. As a result, steam mass and temperature are further restricted. Fig. 10 displays the impacts of steam mass and temperature on the minimum pinch point temperature difference of the evaporator DTmin and the exhaust temperature at the evaporator exit T 0100 at engine 1400 rpm. The data are obtained under the condition that steam injected timing and pressure are 80 deg and 20 bar, respectively. From Fig. 10 (a), it is found that with a small steam mass and high steam temperature, the locations of DTmin mainly appear at point 5s (Fig. 5). But, when steam temperature is low and steam mass is large, the locations of DTmin occur at point 2w. When steam mass is larger than 850 mg per cylinder per cycle (PCPC) and steam temperature is higher than 560 K, the value of DTmin is lower than the limiting temperature difference value of 20 K. As for the exhaust temperature at the evaporator exit T 0100 , as shown in Fig. 10 (b), T 0100 is remarkably decreased as steam mass increases. However, steam temperature has little effects on T 0100 . When steam mass reaches 550 mg PCPC, the value of T 0100 is lower than the limiting temperature value of 393.15 K.

(b) Exhaust temperature at the evaporator exit

evaporator Fig. 10. The impacts of steam mass and temperature on DTmin and T 0100 at 1400 rpm.

L. Li, Z. Zhang / Energy 183 (2019) 958e970

4.2. The effects of steam mass

30

20

50

60

(45)

2500

250

1000 0

40

80

120

Crank angle (deg) Fig. 11. The influences of steam mass on cylinder temperature.

160

100

Baseline 100 mg PCPC 200 mg PCPC 300 mg PCPC

8

100 mg PCPC 200 mg PCPC 300 mg PCPC

6

150

4

100

2

0 1000

1200

1400

1600

1800

2000

Engine speed (rpm) Fig. 13. The effects of steam mass on NGE gross power.

direct injection. In addition, the increase of the NGE gross power is further improved as steam mass increases. With steam mass 300 mg PCPC, the NGE gross power is increased by 3.4e3.9 kW at different engine speed conditions. The largest increase of NGE gross

46 45

1500

90

50

baseline 100 mg PCPC 300 mg PCPC

2000

80

Fig. 12. The influences of steam mass on cylinder pressure.

Indicated efficiency (%)

steam injected timing=50deg

70

Crank angle (deg)

200

pV mRg

From the equation, the in-cylinder pressure is mainly affected by the in-cylinder temperature and the mass of the mixture inside the cylinder. However, from Fig. 11, the in-cylinder temperature is reduced remarkably with steam injection. Therefore, the increase of the in-cylinder pressure is mainly caused by the fact that the mass of working fluids inside the cylinder is remarkably increased with steam injection. At 80 deg crank angle, the in-cylinder pressures are raised by 2.5% and 8.2% in comparison with the baseline engine, when the steam masses are 100 mg PCPC and 300 mg PCPC, respectively. At the power stroke, the working fluids inside the cylinder do positive work on the piston. As a result, the positive work of the piston is increased as in-cylinder pressure is raised, which will result in an improvement of engine power. The influences of steam mass on the NGE gross power are indicated in Fig. 13. It can be seen that the NGE gross power is significantly improved over different engine speeds with steam

In-cylinder temperature (K)

steam injected timing=50deg

10

Gross power (kW)

T¼

baseline 100 mg PCPC 300 mg PCPC

Gross power increases (kW)

The limitations of steam direct injection parameters are analyzed above. In this subsection, the effects of steam mass on the NGE performance are presented. The results are acquired under the condition that steam injected timing is 50 deg, steam injected pressure is 40 bar and steam temperature is 550 K. The impacts of steam injection mass on the in-cylinder temperature are shown in Fig. 11. According to Equations (10) and (11), since the specific heat of steam is much larger than that of the original working fluids inside the cylinder, the specific heat of the mixture of original working fluids and injected steam is increased after steam injection. Therefore, as can be seen from the figure, the in-cylinder temperature is significantly reduced after steam injection. Besides, the in-cylinder temperature is lower when the steam injection mass increases. At 80 deg crank angle, compared with the baseline engine, the in-cylinder temperatures are reduced by 49.1 K and 139.9 K, when steam masses are 100 mg PCPC and 300 mg PCPC, respectively. Fig. 12 indicates the influences of steam mass on the pressure in the cylinder at engine 1400 rpm. As can be seen from the figure, incylinder pressure is remarkably raised after steam injection. According to the ideal gas equation, the in-cylinder temperature can be expressed as Equation (45).

In-cylinder pressure (bar)

40

Baseline 100 mg PCPC 200 mg PCPC 300 mg PCPC

2.5

100 mg PCPC 200 mg PCPC 300 mg PCPC

2.0

44 1.5 43 1.0 42 0.5

41 40

0.0 1000

1200

1400

1600

1800

2000

Engine speed (rpm) Fig. 14. The effects of steam mass on NGE indicated efficiency.

Indicated efficiency increases (%)

966

L. Li, Z. Zhang / Energy 183 (2019) 958e970

The impacts of steam temperature on the NGE performance are discussed in the subsection. The results are acquired under the condition that steam mass is 300 mg PCPC, steam injected pressure is 20 bar, and steam injected timing is 80 deg. Fig. 16 presents the influences of steam temperature on pressure in the cylinder at engine 1400 rpm. As shown in the figure, steam temperature has little impacts on the pressure in the cylinder. At 100 deg crank angle, the in-cylinder pressure is only raised by 0.8%, when steam temperature is increased from 500 K to 600 K. This is

steam injection timing=80 deg

16 14 12 10 8 80

190 4 180 2

110

120

due to that as steam temperature increases, only a slight increase of the steam enthalpy is obtained. For instance, at 20 bar pressure, the specific enthalpies of steam are 2841.4 kJ/kg and 3085.6 kJ/kg, when steam temperatures are 500 K and 600 K, respectively. In other words, the specific enthalpy of steam is only increased by 8.6%, when steam temperature increases from 500 K to 600 K. Therefore, as steam temperature increases, the positive work of the piston will be only slightly changed. As a result, steam temperature has little influences on the NGE performance. Fig. 17 shows the influences of steam temperature on the NGE gross power and BSFC at 1400 rpm. As it is observed from the figure, steam temperature has little impacts on the NGE gross power and BSFC. When steam temperature increases from 500 K to 600 K, the improvements of NGE gross power and BSFC are both less than 0.3%. In conclusion, even though the NGE gross power and BSFC have a little improvement by raising the steam temperature, the stronger heat transferability of the evaporator is needed to obtain higher steam temperature. Thus, in consideration of the benefits of the system, the superheated degree of the steam should not be too large.

Gross power (kW)

6

100

Fig. 16. The influences of steam temperature on cylinder pressure at 1400 rpm.

182

Gross power BSFC

8

100 mg PCPC 200 mg PCPC 300 mg PCPC

BSFC reductions (%)

BSFC (g/kW.h)

Baseline 100 mg PCPC 200 mg PCPC 300 mg PCPC

90

Crank angle (deg)

147 200

baseline 500K 600K

146

181

145

180

144 170

0 1000

1200

1400

1600

1800

2000

BSFC (g/kW.h)

4.3. The effects of steam temperature

18

In-cylinder pressure (bar)

power is 3.9 kW acquired at engine 1600 rpm, and the lowest improvement is 3.4 kW obtained at 1000 rpm. Since the increase of positive piston work with steam injection, more thermal energy is transferred to engine power. Therefore, steam direct injection benefits for the rise of NGE efficiency. The impacts of steam mass on the NGE indicated efficiency are shown in Fig. 14. From the figure, the NGE indicated efficiency is remarkably increased with steam injection. Besides, the improvement of NGE indicated efficiency is much larger at lower engine speed condition. Under the condition that steam mass is 300 mg PCPC, the NGE indicated efficiency is increased by 0.9e1.6% over different speeds. The largest increase of the NGE indicated efficiency is 1.6% achieved at engine 1000 rpm, and the smallest increase is 0.9% obtained at 2000 rpm. This is because that, for the basic NGE, the engine incylinder pressure at low speed condition is smaller than that at high speed condition. With the same steam injection parameter, the percentage increase of in-cylinder pressure at low speed condition is larger than that at high speed condition. Therefore, the increase percentage of indicated work at low speed condition is larger in comparison with that at high speed condition, and the increase of indicated efficiency at low speed condition is also larger than that at high speed condition. Because of the improvement of NGE efficiency, the NGE BSFC is decreased with steam direct injection. The impacts of steam mass on the NGE BSFC are presented in Fig. 15. It is found that the NGE BSFC is decreased with steam direct injection significantly. A larger reduction of NGE BSFC can be acquired with higher steam injection mass. With 300 mg PCPC steam mass, 2.3e4.0% reductions of the NGE BSFC are obtained over various engine speeds. The largest BSFC reduction is 4.0% acquired at engine 1000 rpm. And the lowest BSFC improvement is 2.3% at 2000 rpm. However, since the decrease of NGE indicated efficiency, as engine speed increases, the reduction of BSFC is decreased.

967

179 500

520

540

560

580

600

Steam temperature (K)

Engine speed (rpm) Fig. 15. The influences of steam mass on BSFC.

Fig. 17. The influences of steam temperature on NGE gross power and BSFC at 1400 rpm.

L. Li, Z. Zhang / Energy 183 (2019) 958e970

195

35

baseline 50deg 80deg

steam injected timing=50deg

30 25

steam injected timing=80deg

20

50deg 60deg 70deg 80deg

190 BSFC (g/kW.h)

In-cylinder pressure (bar)

40

8

50deg 60deg 70deg 80deg

6

185 4 180 2

175

BSFC reductions (%)

968

15 0

170 1000

10 50

60

70

80

90

100

1200

110

1400

1600

1800

2000

Engine speed (rpm)

Crank angle (deg)

Fig. 20. The effects of steam injected timing on BSFC.

Fig. 18. The influences of steam injected timing on cylinder pressure at 1400 rpm.

0.5-1.9% to 2.3e4.1%, when steam injected timing is advanced from 80 deg to 50 deg.

In this subsection, the impacts of steam injected timing on the NGE performance are analyzed. The data are obtained under the condition of 300 mg PCPC steam mass, 550 K steam temperature and 40 bar steam injected pressure. Fig. 18 illustrates the influences of steam injected timing on the pressure in the cylinder at 1400 rpm. As it can be observed from the figure, the initial time of the increase of the in-cylinder pressure occurs early as steam injecting timing advances. Thus, the positive piston work is further increased. This is good for the improvement of the NGE power. For the increase of the positive piston work, the NGE gross power is further improved by advancing steam injected timing. Fig. 19 indicates the impacts of steam injected timing on the NGE gross power with respect to engine speed. As steam injected timing is advanced from 80 deg to 50 deg, the improvement range of the NGE gross power at all engine speed is increased from 0.2-1.8% to 1.8e4.1%. Because of the improvement of engine power with advancing steam injected timing, the NGE BSFC is reduced. Fig. 20 shows the effects of steam injected timing on NGE BSFC over various speeds. The reduction range of BSFC at all engine speeds is increased from

4.5. The optimal steam mass over various speeds According to the discussion above, steam direct injection has great influences on the NGE performance, especially steam mass and injected timing. However, steam mass is limited by pinch point temperature difference of the evaporator DTmin and exhaust temperature at the evaporator exit T 0100 , and steam injected timing is limited by pressure in the cylinder. In this subsection, the optimal steam mass and the potentials of the steam direct injection method for fuel savings of the NGE are presented. Fig. 21 shows the optimal steam mass and the largest BSFC reductions of the NGE at all speeds. The results are obtained under the condition of 550 K steam temperature, 40 bar steam injected pressure and 50 deg steam injected timing. From the figure, the optimal steam mass value ranges from 390 mg PCPC to 520 mg PCPC. With the rise of the engine speed, the optimal steam mass is increased. With optimal steam mass, 3.9e5.2% reductions of the NG BSFC are obtained over various engine speeds. Therefore, the steam direct injection at the power stroke method obtains a good fuel reduction.

8 50deg 60deg 70deg 80deg

6

150

4

100

2

50

0 1000

1200

1400

1600

550

Optimal steam mass BSFC reductions

1800

2000

Change (%)

Gross power (kW)

200

50deg 60deg 70deg 80deg

8 Gross power increase (%)

250

6

500

4

450

2

400

350

0 1000

1200

1400

1600

1800

Engine speed (rpm)

2000

Optimal steam injection mass (mg PCPC)

4.4. The effects of steam injected timing

Engine speed (rpm) Fig. 19. The influences of steam injected timing on NGE gross power.

Fig. 21. The optimal steam mass and the largest BSFC reductions at various engine speeds.

L. Li, Z. Zhang / Energy 183 (2019) 958e970

5. Conclusions In this paper, steam direct injection at the power stroke for fuel savings in a NGE has been studied by employing a numerical model. Through the analysis above, some conclusions are drawn. (1) Steam direct injection can significantly reduce the NGE BSFC over various speeds. This is because that pressure in the cylinder is remarkably increased. More thermal energy is transferred to engine power. Therefore, the NGE efficiency is improved. With optimal steam mass, 3.9e5.2% decreases of BSFC are acquired at all speeds, when steam temperature, injected pressure and timing are 550 K, 40 bar and 50 deg, respectively. (2) Steam mass and injected timing have great impacts on the NGE BSFC. By increasing steam mass or advancing steam injected timing, the reductions of BSFC are further increased. However, steam mass is limited by pinch point temperature difference of the evaporator and exhaust temperature at the evaporator exit. Besides, steam injected timing is limited by pressure in the cylinder. (3) NGE fuel economy is little influenced by steam temperature. This is due to that as steam temperature increases, only a slight increase of the steam enthalpy is obtained. In consideration of the benefits of the system, the superheated degree of the steam should not be too large. In general, applying steam direct injection at the power stroke on the NGE may be a good method for fuel savings, but it must be further validated with experiments. In the further, the experimental activities will be carried out.

[10]

[11]

[12]

[13] [14]

[15]

[16]

[17]

[18]

[19]

[20] [21]

[22]

[23]

Conflict of interest statement [24]

The authors declare that there is no conflict of interest with any individual or organization for the present work. [25]

Acknowledgments This work was supported by the Guangdong Province Public Welfare Research and Capacity Building Grant (Contract No. 2014B010106004). The authors would like to thank Professor Rongchao Zhao of South China University of Technology for his great support during the research process. References [1] Bicer Y, Dincer I. Life cycle environmental impact assessments and comparisons of alternative fuels for clean vehicles. Resour Conserv Recycl 2018;132: 141e57. [2] Geng P, Cao EM, Tan QM, Wei LJ. Effects of alternative fuels on the combustion characteristics and emission products from diesel engines: a review. Renew Sustain Energy Rev 2017;71:523e34. [3] Othman MF, Adam A, Najafi G, Mamat R. Green fuel as alternative fuel for diesel engine: a review. Renew Sustain Energy Rev 2017;80:694e709. [4] Aslam MU, Masjuki HH, Kalam MA, Abdesselam H, Mahlia TMI, Amalina MA. An experimental investigation of CNG as an alternative fuel for a retrofitted gasoline vehicle. Fuel 2006;85(5e6):717e24. [5] Tang QJ, Fu JQ, Liu JP, Zhou F, Yuan ZP, Xu ZX. Performance improvement of liquefied natural gas (LNG) engine through intake air supply. Appl Therm Eng 2016;103:1351e61. [6] Hassan MH, Kalam MA, Mahlia TMI, Aris I, Nizam MK, Abdullah S, et al. Experimental test of a new compressed natural gas direct injection engine. Energy Fuel 2009;23(10):4981e7. [7] Korakianitis T, Namasivayam AM, Crookes RJ. Natural-gas fueled sparkignition (SI) and compression-ignition (CI) engine performance and emissions. Prog Energ Combust 2011;37(1):89e112. [8] Cho HM, He BQ. Spark ignition natural gas engines - a review. Energy Convers Manag 2007;48(2):608e18. [9] Wu ZK, Han ZY. Numerical investigation on mixture formation in a

[26] [27]

[28]

[29]

[30]

[31]

[32]

[33]

[34]

[35]

[36]

[37]

969

turbocharged port-injection natural gas engine using multiple cycle simulation. J Eng Gas Turbines Power 2018;140(5). Navarro E, Leo TJ, Corral R. CO2 emissions from a spark ignition engine operating on natural gas-hydrogen blends (HCNG). Appl Energy 2013;101: 112e20. Korakianitis T, Namasivayam AM, Crookes RJ. Natural-gas fueled sparkignition (SI) and compression-ignition (CI) engine performance and emissions. Prog Energ Combust 2011;37(1):89e112. Ran ZN, Hariharan D, Lawler B, Mamalis S. Experimental study of lean spark ignition combustion using gasoline, ethanol, natural gas, and syngas. Fuel 2019;235:530e7. Singh DV, Pedersen E. A review of waste heat recovery technologies for maritime applications. Energy Convers Manag 2016;111:315e28. Zhao RC, Zhuge WL, Zhang YJ, Yin Y, Chen Z, Li ZG. Parametric study of power turbine for diesel engine waste heat recovery. Appl Therm Eng 2014;67(1e2): 308e19. Zhao RC, Zhuge WL, Zhang YJ, Yang MY, Martinez-Botas R, Yin Y. Study of twostage turbine characteristic and its influence on turbo-compound engine performance. Energy Convers Manag 2015;95:414e23. Zhao RC, Zhuge WL, Zhang YJ, Yin Y, Zhao YT, Chen Z. Parametric study of a turbocompound diesel engine based on an analytical model. Energy 2016;115:435e45. Katsanos CO, Hountalas DT, Zannis TC. Simulation of a heavy-duty diesel engine with electrical turbocompounding system using operating charts for turbocharger components and power turbine. Energy Convers Manag 2013;76:712e24. Liona S, Michos CN, Vlaskos I, Rouaud C, Taccani R. A review of waste heat recovery and Organic Rankine Cycles (ORC) in on-off highway vehicle Heavy Duty Diesel Engine applications. Renew Sustain Energy Rev 2017;79: 691e708. Tian H, Shu GQ, Wei HQ, Liang XY, Liu LN. Fluids and parameters optimization for the organic Rankine cycles (ORCs) used in exhaust heat recovery of Internal Combustion Engine (ICE). Energy 2012;47(1):125e36. Chen H, Guo Q, Yang L, Liu SH, Xie XL, Chen ZY, et al. A new six stroke single cylinder diesel engine referring Rankine cycle. Energy 2015;87:336e42. Liu ZC, Yuan X, Tian J, Han YQ, Yu KB, Teng PK. Effects of injection timing on transient performance of a regulated two-stage turbocharged diesel engine with turbine bypass valve. Int J Auto Tech-Kor. 2018;19(5):783e94. Shu GQ, Liang YC, Wei HQ, Tian H, Zhao J, Liu LN. A review of waste heat recovery on two-stroke IC engine aboard ships. Renew Sustain Energy Rev 2013;19:385e401. Al-Nimr MA, Alajlouni AA. Internal combustion engine waste heat recovery by a thermoelectric generator inserted at combustion chamber walls. Int J Energy Res 2018;42(15):4853e65. Ji DX, Wei ZB, Mazzoni S, Mengarelli M, Rajoo S, Zhao JY, et al. Thermoelectric generation for waste heat recovery: application of a system level design optimization approach via Taguchi method. Energy Convers Manag 2018;172: 507e16. Wang RC, Yu W, Meng XP. Performance investigation and energy optimization of a thermoelectric generator for a mild hybrid vehicle. Energy 2018;162: 1016e28. Zhu SP, Deng KY, Qu S. Thermodynamic analysis of an in-cylinder waste heat recovery system for internal combustion engines. Energy 2014;67:548e56. Zhu SP, Liu S, Qu S, Deng KY. Thermodynamic and experimental researches on matching strategies of the pre-turbine steam injection and the Miller cycle applied on a turbocharged diesel engine. Energy 2017;140:488e505. Fu JQ, Liu JP, Deng BL, Feng RH, Yang J, Zhou F, et al. An approach for exhaust gas energy recovery of internal combustion engine: steam-assisted turbocharging. Energy Convers Manag 2014;85:234e44. Fu JQ, Liu JP, Yang YP, Ren CQ, Zhu GH. A new approach for exhaust energy recovery of internal combustion engine: steam turbocharging. Appl Therm Eng 2013;52(1):150e9. Zhao RC, Li WH, Zhuge WL, Zhang YJ, Yin Y. Numerical study on steam injection in a turbocompound diesel engine for waste heat recovery. Appl Energy 2017;185:506e18. Zhao RC, Zhang ZB, Zhuge WL, Zhang YJ, Yin Y. Comparative study on different water/steam injection layouts for fuel reduction in a turbocompound diesel engine. Energy Convers Manag 2018;171:1487e501. Parlak A, Ayhan V, Ust Y, Sahin B, Cesur I, Boru B, et al. New method to reduce NOx emissions of diesel engines: electronically controlled steam injection system. J Energy Inst 2012;85(3):135e9. Cesur I, Parlak A, Ayhan V, Boru B, Gonca G. The effects of electronic controlled steam injection on spark ignition engine. Appl Therm Eng 2013;55(1e2): 61e8. Gonca G. Investigation of the effects of steam injection on performance and NO emissions of a diesel engine running with ethanol-diesel blend. Energy Convers Manag 2014;77:450e7. Gonca G, Sahin B. Simulation of performance and nitrogen oxide formation of a hydrogen-enriched diesel engine with the steam injection method. Therm Sci 2015;19(6):1985e94. Gonca G. Investigation of the influences of steam injection on the equilibrium combustion products and thermodynamic properties of bio fuels (biodiesels and alcohols). Fuel 2015;144:244e58. Gonca G, Sahin B, Parlak A, Ust Y, Ayhan V, Cesur I, et al. The effects of steam injection on the performance and emission parameters of a Miller cycle diesel

970

L. Li, Z. Zhang / Energy 183 (2019) 958e970

engine. Energy 2014;78:266e75. [38] Gonca G, Sahin B, Ust Y, Parlak A, Safa A. Comparison of steam injected diesel engine and Miller cycled diesel engine by using two zone combustion model. J Energy Inst 2015;88(1):43e52. [39] Gonca G, Sahin B. The influences of the engine design and operating parameters on the performance of a turbocharged and steam injected diesel engine running with the Miller cycle. Appl Math Model 2016;40(5e6):3764e82. [40] Gonca G, Sahin B. Effect of turbo charging and steam injection methods on the performance of a Miller cycle diesel engine (MCDE). Appl Therm Eng 2017;118:138e46. [41] Gonca G, Sahin B, Parlak A, Ayhan V, Cesur I, Koksal S. Investigation of the effects of the steam injection method (SIM) on the performance and emission formation of a turbocharged and Miller cycle diesel engine (MCDE). Energy 2017;119:926e37. [42] REFPROP version 7.1. NIST standard reference database 23. America: The U.S.Secretary of Commerce; 2003. [43] Zhang ZB, Li LF. Investigation of in-cylinder steam injection in a turbocharged diesel engine for waste heat recovery and NOx emission control. Energies 2018;11(4). [44] Mazas AN, Fiorina B, Lacoste DA, Schuller T. Effects of water vapor addition on the laminar burning velocity of oxygen-enriched methane flames. Combust Flame 2011;158(12):2428e40.

[45] Galmiche B, Halter F, Foucher F, Dagaut P. Effects of dilution on laminar burning velocity of premixed methane/air flames. Energy Fuel 2011;25(3): 948e54. [46] Soylu S, Van Gerpen J. Development of an autoignition submodel for natural gas engines. Fuel 2003;82(14):1699e707. [47] Soylu S, Van Gerpen J. Development of empirically based burning rate submodels for a natural gas engine. Energy Convers Manag 2004;45(4):467e81. [48] Lounici MS, Loubar K, Balistrou M, Tazerout M. Investigation on heat transfer evaluation for a more efficient two-zone combustion model in the case of natural gas SI engines. Appl Therm Eng 2011;31(2e3):319e28. [49] Heywood JB. Internal combustion engine fundamentals. McGraw-Hill; 1988. [50] Verhelst S, Sheppard CGW. Multi-zone thermodynamic modelling of sparkignition engine combustion - an overview. Energy Convers Manag 2009;50(5):1326e35. [51] Ma FH, Wang Y, Wang MY, Liu HQ, Wang JJ, Ding SF, et al. Development and validation of a quasi-dimensional combustion model for SI engines fuelled by HCNG with variable hydrogen fractions. Int J Hydrogen Energy 2008;33(18): 4863e75. [52] Guo C, Du XZ, Yang LJ, Yang YP. Performance analysis of organic Rankine cycle based on location of heat transfer pinch point in evaporator. Appl Therm Eng 2014;62(1):176e86.