Theoretical and experimental investigations on energy balance on DMDF engine

Fuel 164 (2016) 393–402

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Theoretical and experimental investigations on energy balance on DMDF engine Guopeng Han a, Anren Yao b, Chunde Yao a,⇑, Quangang Wang a, Bin Wang a, Hongyuan Wei a, Taoyang Wu a, Meijuan Liu a a b

State Key Laboratory of Engines, Tianjin University, Tianjin 300072, China School of Materials Science and Engineering, Tianjin University, Tianjin 300072, China

h i g h l i g h t s
Model to analyze energy balance of DMDF mode was built.
Reduction in cooling loss of DMDF mode is the most effective factor at high loads.
Incomplete combustion loss in DMDF mode is the dominant factor for high SR at low load.
Cooling loss of DMDF mode is lower than that of D mode and even lower as SP rises.
High methanol temperature makes SR and incomplete combustion loss reduced.

a r t i c l e

i n f o

Article history: Received 5 August 2015 Received in revised form 23 September 2015 Accepted 9 October 2015 Available online 22 October 2015 Keywords: Diesel methanol dual fuel Energy balance Diesel engine

a b s t r a c t This paper presented the analytical results for the energy balance of a diesel methanol dual-fuel (DMDF) engine. A methanol injection system was fixed on a diesel engine and let the engine to alternate to run in either the pure diesel (D) mode or DMDF mode. Primary application of the thermal balance analysis was to investigate the cause of methanol replacement ratio SR changing with engine loads. Calculations were conducted within the control volume by comparing each of the energy terms. The results show that the reduction in cooling loss in DMDF mode is the dominant factor for low SR at high loads. While, a substantial increase in incomplete combustion loss in DMDF mode is the most important reason for high SR at low loads. Another application of the thermal balance analysis was to research the cooling loss in DMDF mode which implied unique characteristics. Therefore, the changing methanol replacement rate SP experiments were carried out in various engine loads, giving the results that the cooling loss in DMDF mode is always lower than that in D mode, and it goes even lower as SP rises. Finally, the effect of methanol temperature on SR was also investigated from the thermal balance point of view. The results reveal that the incomplete combustion loss decreasing in higher methanol temperature conditions will lead to a certain extent of SR drop. All these results are related to the energy flow distribution within the engine. Ó 2015 Elsevier Ltd. All rights reserved.

1. Introduction As a new kind of alternative fuel used in engine, methanol has more advantages than traditional fuels, such as higher thermal efficiency and lower exhaust emissions. Decided by physical properties, methanol cannot be blended with diesel without the help of additives or some other physical methods. To apply methanol on diesel engines, Yao et al. [1,2] proposed a DMDF combustion method, which has been proved to be the most feasible way to apply methanol on heavy duty diesel engines. To be specific, DMDF ⇑ Corresponding author. Tel.: +86 22 2740 6649; fax: +86 22 2738 3362. E-mail address: [email protected] (C. Yao). http://dx.doi.org/10.1016/j.fuel.2015.10.024 0016-2361/Ó 2015 Elsevier Ltd. All rights reserved.

is realized by fixing methanol injectors on intake pipeline, and those injectors are controlled by an independent methanol Electronic Control Unit (ECU). A 4 bar methanol injection pressure is ensured by the methanol supplying system. While running in DMDF mode, the methanol/air mixture is ignited by diesel spray in the cylinder. Previous researches show that DMDF combustion can alleviate the trade-off relationship between soot and NOx emissions in traditional diesel engine. Meanwhile, it also has a higher thermal efficiency [3,4]. The methanol replacement rate SP and replacement ratio SR are defined to evaluate the using effect of methanol on diesel engine. Calculations of the two parameters are shown in Eqs. (1) and (2):

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G. Han et al. / Fuel 164 (2016) 393–402

M d  M dm  100% Md Mm SR ¼ M d  M dm

SP ¼

ð1Þ ð2Þ

where Md and Mdm are mass flow rates of diesel in D and DMDF modes, kg/h. Mm is the mass flow rate of methanol, kg/h. Methanol has a much lower low heat value (LHV) than diesel. Theoretically, 2.16 kg methanol equals to 1 kg diesel in LHV, so the theoretic value of SR is also 2.16 [5,6]. However, SR is frequently found to be lower than 2.16 in vehicle road tests [7]. This phenomenon was ascribed to thermal efficiency improvement in DMDF engine by previous studies. However, the thermal efficiency improvement and the SR lower than theoretic value just describe the same thing in two different means of expressions. They cannot explain each other. So, it is necessary to conduct the energetic analysis on DMDF engine and find the exact factors affecting SR. Energetic analysis is widely used in engine performance researches. Huang et al. [8,9] investigated the combustion and emissions of a compression–ignition engine fuelled with diesel– oxygenate blends. Singh et al. [10] and Donn et al. [11] studied the effect of different running parameters, such as cooling water temperature, injection strategy, and EGR rate on the engine energy balance. Romero et al. [12] proposed an energy balance calculation method direct at NEDC driving cycles and analyzed the energy balance during the warming-up process. Rabeau and Magand [13] built a thermal balance model on AMEsim platform, and gave some suggestions on improving thermal efficiency basing on the simulation results. Da Costa et al. [14] performed both energetic and exergy balance analyses on a diesel/nature gas dual-fuel engine. Durgun and Sßahin [15] built a mathematical model to analyze gas fumigation diesel engine’s energy balance, and showed some predictive results. However, studies on energy balance characteristics of DMDF engine is still rare to see in literatures. This paper presented the analytical results for the energy balance on a DMDF engine. All the energy terms, such as output power, cooling loss, exhaust enthalpy, incomplete combustion loss and residual loss can be quantified according to the experimental data and basic calculation equations. Therefore, the reasons why SR various at different engine loads, as well as the characteristics of cooling loss in DMDF mode and the effect of methanol temperature on SR were investigated. 2. Experimental apparatus and setup 2.1. Experimental apparatus and fuels Prototypical engine used in this study is an YC4D120-30 turbocharged EUP diesel engine. Table 1 shows the detailed engine specifications. Schematic of the engine test system is shown in Fig. 1. Three methanol injectors are arranged at the intake port. When the engine runs in DMDF mode, methanol is injected into the intake pipeline along the air flow and then forms the

methanol/air mixture. Main experimental facilities are listed in Table 2 together with their applications. Commercial methanol (99.9% purity) and diesel fuel were used in the experiments. Table 3 shows the main properties of the two fuels. Methanol used in this study was produced from coal. An equivalent molecular formula CH1.95 can be obtained according to the mass ratio of C and H in the diesel fuel. This formula will be used in the following calculation parts. 2.2. Engine setup The DMDF experiments at a fixed operating point are describe here: The engine is set to run in constant speed and torque mode with the help of dynamometer. The diesel injection control is realized by CAN communication between the engine and computer. When the engine runs in DMDF mode, the diesel consumption will decrease with rising methanol injection map. Therefore, the SP and SR can be obtained from the measuring results of fuel consumption rates according to Eqs. (1) and (2). The boundary conditions are maintaining a steady combustion with a moderate pressure rise rate, as well as a maximum pressure lower than the safely limitations. For this engine, the 1.5 MPa/deg for maximum pressure rise rate and the 15 MPa for maximum explosion pressure are used for limitations. While at low loads, the flameout should be avoid when higher SP reaches. These can be realized by monitoring the in-cylinder pressures with the help of pressure transducers and combustion analyzer. The experiments conducted in this study are listed in Table 4. These experiments can be divided into three groups. The first group aims to investigate the effect of loads on DMDF engine thermal balance. Hence, the 1660 r/min–105 N m and 1660 r/min– 420 N m operating points, which correspond to A25 and A100 with 33% and 43% replacement rates respectively, were selected for analysis. The SR deviates from 2.16 for ±0.5 at the two points. Group 2 concerns the characteristics of cooling loss in DMDF engine. In this part, A25-100 working points were selected for the increasing SP experiments. The maximum of SP wasn’t controlled such high for only an integral changing law of cooling loss in DMDF mode was expected there. The last group of experiments concentrates on the thermal balance changing with methanol temperatures. This idea comes from the observing results that when the methanol supplying pipeline is close to the exhaust pipe, or the methanol tank is placed under blazing sun for a long time, the economic efficiency of DMDF engine will be improved. The methanol temperature was controlled by a combination of heat exchanger and heat tap installed on the methanol supplying pipeline. Still, the SP was not controlled strictly in this section. One instruction for the experiments is that the unspecified parameters, such as the intake air temperature, cooling water temperature and the diesel injection timing are kept constant at a specified speed–torque condition. 3. Mathematic model for energy balance calculation

Table 1 Engine specifications.

3.1. Theoretical analysis of energy balance

Parameters

Specifications

Bore  stroke (mm) Compression ratio Number of cylinders Displacement (L) Nozzle (number  bore diameter/mm) Injection pressure (MPa) IVC (°CA ATDC) EVO (°CA ATDC) Max. power (kW@r/min)

108  115 17:1 4-stroke, 4-inline 4.214 7  0.16 28 130.3 112.2 103  1600

In this study, an opening system from the inlet of the compressor to the outlet of the turbine was selected as the control volume. The turbocharger was included within the control volume, so the energy transformation between turbine and compressor needn’t to be analyzed separately [11,15]. When the engine runs in D mode, the energy terms entering the control volume consist of intake air enthalpy Hin and chemical energy of diesel Qf,d, while the energy terms exiting the control volume consist of output power We, exhaust gas enthalpy Hexh (exhaust loss), energy taken

G. Han et al. / Fuel 164 (2016) 393–402

395

Fig. 1. Schematic diagram of experimental system.

Table 2 Main experimental facilities. Apparatus

Type

Applications

Hydraulic dynamometer Fuel consumptions meter Cylinder pressure sensor Combustion analyzer Exhaust gas analyzer

Yi KeCW5-5000/ 15000 FCMM-2 Kistler 6052CU20

Engine speed and torque controlling Diesel and methanol flow measurement Cylinder pressure measurement

AVL6120

Combustion analyzing Exhaust species analyzing

Liquid flow meter

HORIBA MEXA7100D LWY-50L

Air flow meter Hygrometer

ToCeil-20N100 272-A

3.7 times as gasoline, which leads to a substantial temperature drop of the intake air. If the energy released by intake air due to decline of its temperature cannot satisfy the methanol evaporation, a portion of ambient energy will be absorbed into the control volume. This amount of heat absorption Qab is the difference of methanol latent heat of vaporization and energy released by intake air. Schematic of energy flow distribution in DMDF mode is shown in Fig. 2(b), and the energy balance equation is shown in Eq. (4):

Hin þ Q f ;d þ Q f ;m þ Q ab ¼ W e þ Hexh þ Q c þ Q ic þ Q ub þ Q misc 3.2. Calculation of energy terms

Cooling water flow measurement Air flow measurement Air humidity measurement

Calculation of some energy terms in the energy balance equations are shown in Eqs. (5)–(8):

Q f ;d ¼ Hu;d  M r;d by cooling water Qc (cooling loss), heat taken from intake air by intercooler water (intercooler loss) Qic, incomplete combustion loss Qub, and residual loss Qmisc. Schematic of energy flow distribution in D mode is shown in Fig. 2(a), and the energy balance equation is shown in Eq. (3):

Hin þ Q f ;d ¼ W e þ Hexh þ Q c þ Q ic þ Q ub þ Q misc

ð4Þ

ð3Þ

When the engine runs in DMDF mode, the methanol is injected into intake air from the intake port and then forms the methanol/ air mixture. Therefore, chemical energy of methanol Qf,m and enthalpy of methanol steam Hm joined into the entering energy terms. Furthermore, latent heat of vaporization for methanol is

Table 3 Fuel properties of diesel and methanol [4]. Parameters

Diesel

Methanol

Density at 20 °C (kg/m3) C (wt.%) H (wt.%) O (wt.%) Cetane number Heat of evaporation (kJ/kg) Lower heating value (MJ/kg) Sulfur content (mg/kg) Flash point (°C) Self-ignition temperature (°C)

830 86 14 0 P49 250 42.5 6350 55 250

796 37.5 12.5 50 5 1110 19.7 0 12 460

ð5Þ

Q f ;m ¼ Hu;m  Mr;m

ð6Þ

W e ¼ P e  3600

ð7Þ

Q c ¼ cp;w  Mr;w  ðtout;w  t in;w Þ

ð8Þ

where Hu,d, Hu,m are LHVs of diesel and methanol respectively, kJ/kg. As water always exists in gaseous state among exhaust gases, the LHV is used here rather than high heat value (HHV) [14]. Mr,d, Mr,m, Mr,w are mass flow rates of diesel, methanol and cooling water respectively, kg/h. Pe is the output power, kW. cp,w is specific heat capacity of cooling water, kJ/(kg K). tin,w, tout,w are temperatures at the entrance and exit of cooling water, K. In the following paragraphs, the calculation of the other energy terms will be introduced. 3.2.1. Calculation of intake enthalpy Both the intake and exhaust gases can be dealt with as ideal gases and their enthalpies only related to temperatures [16,17]. The intake enthalpy Hin composed of intake air enthalpy Hair and methanol steam enthalpy Hm. Their relationship is shown in Eq. (9):

Hin ¼ Hair þ Hm

ð9Þ

Intake air enthalpy can be calculated by Eq. (10):

Hair ¼ ðhair;tair  hair;298:15 Þ  M r;air

ð10Þ

where hair is air enthalpy per kilogram, kJ/kg. Mr,air is air mass flow rate, kg/h. tair is air temperature at the inlet of the control volume, K.

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Table 4 Experimental content. Speed (r/min)

Torque (N m)

SP (%)

Methanol temperatures (°C)

Experimental objectives

Group 1

1660 1660

105 420

33 43

20 20

The effect of loads on DMDF engine thermal balance

Group 2

1660 1660 1660 1660

105 210 315 420

0–13 0–35 0–55 0–48

20 20 20 20

The characteristics of cooling loss in DMDF mode

Group 3

1660 1660 1660 1660

420 420 420 420

0–29 0–40 0–40 0–38

30 40 50 60

The effect of methanol temperatures on DMDF engine thermal balance

The LHVs of the two fuels are based on 298.15 K, so the zero gas enthalpies are also referred to 298.15 K [10]. According to the definition of ideal gas enthalpy difference [16,17], the methanol steam enthalpy can be calculated by Eq. (11):

Z Hm ¼ M r;m 

t air;a

cp;m  dt

ð11Þ

where tair,a is the air temperature downstream the methanol injectors, which is considered to be the methanol steam temperature, K. cp,m is the specific heat capacity of methanol steam at constant pressure, kJ/(kg K). The calculation of cp,m is shown in Eq. (12):

cp;m ¼ C 0 þ C 1  h þ C 2  h2 þ C 3  h3

298:15

(a) Energy flow distribution in D mode

(b) Energy flow distribution in DMDF mode Fig. 2. Energy flow distribution in two running modes.

ð12Þ

G. Han et al. / Fuel 164 (2016) 393–402

397

Fig. 3. Calculation process of the exhaust gas enthalpies.

where h = tair,a/1000, C0 = 0.66, C1 = 2.21, C2 = 0.81, C3 = 0.89 [16,17]. In summary, the intake enthalpy can be inferred from Eq. (9) combined with Eqs. (10)–(12).

of the two fuels Mr,d, Mr,m. Calculation of the equivalent fuel formula is shown in Eqs. (14)–(18):

g¼ 3.2.2. Calculation of exhaust enthalpy The main species in exhaust gases are CO2, H2O, N2, O2, which account for more than 98% of the total exhaust volume flow. Equipment for exhausts measuring used in this study is HORIBA MEXA7100D gas analyzer. Dry concentration of CO2, CO, O2, THC and NOx can be measured directly by the analyzer, while the concentration of N2 and H2O are still unknown. Basing on the calculation method recorded in literature [18,19], an improved mathematic model for calculating the unknown gas concentrations was put forward considering the combustion characteristics of DMDF mode. The method used here is described in the following paragraphs. Fuel reaction equation can be simplified as Eq. (13):

Ca Hb Oc Nd þ A ð3:7292N2 þ O2 þ 0:0444Ar þ 0:0019CO2 Þ þ BH2 O ! v 1 CO þ v 2 CO2 þ v 3 O2 þ v 4 H2 O þ v 5 Cz Hx þ v 6 H2 þ v 7 N2 þ v 8 NO þ v 9 NO2 þ v 10 Ar

ð13Þ

where CaHbOcNd is the equivalent fuel formula in DMDF combustion. This can be obtained from the methanol formula CH3OH and equivalent diesel formula CH1.95 [18], as well as the mass flow rates

M r;m 12:11  1 þ 1:008  4 þ 15:999  1 12:11  1 þ 1:0084  1:95  M r;d

a¼1þg

ð14Þ

b ¼ 1:95 þ 4  g

ð15Þ ð16Þ

c¼g

ð17Þ

d¼0

ð18Þ

where CzHx is the formula of THC, and it is equivalent to C3H8 by the exhaust gas analyzer. All the calculations in this study were carried out on MATLAB platform. Fig. 3 shows the calculation process of the exhaust gas enthalpies. The configuration file contains the values of Mr,d, Mr,m, Mr,air, relative humidity of air U, interpolation table of saturated vapor pressure Pg, atmospheric pressure Pa, exhaust temperature Texh, dry concentrations of exhaust gases (CO2, CO, O2, THC, NO, NO2) and their enthalpy interpolation tables. Eqs. (19)–(21) and Eq. (24) are mass conservation equations of C, H, O, N. Eq. (22) gives the relationship between B and A⁄. K is the equilibrium constant of water gas shift reaction, see in Eq. (25), which is introduced to keep the equation set closed. The recommended K = 3.5 is used in Eq. (23) [18].

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Heat absorption Chemical energy of fuel Intake enthalpy

800000 D

300000

Residual loss Incomplete combustion loss Cooling loss Exhaust loss Intercooler loss Output power

D DMDF

Chemical energy of fuel

DMDF D

D

Energy (kJ)

600000

Energy (kJ)

DMDF

250000

DMDF

400000

Residual loss Incomplete combustion loss Cooling loss Exhaust loss Intercooler loss Output Power

200000 150000 100000

200000 50000 0

0

Energy entering the control volume

Energy entering the control volume

Energy exiting the control volume

Energy exiting the control volume

(a) Energy flow distribution

(a) Energy flow distribution 100000

340000 Intercooler loss

320000

Heat absorption

Energy changes Engine power of the DMDF mode Engine power of the D mode

80000

Exhaust loss

300000

Energy (kJ)

Energy (kJ)

Energy changes Engine power of the DMDF mode Engine power of the D mode

Cooling loss

280000

Incomplete combustion loss Residual loss

60000

Residual loss Intercooler loss Heat absorption

40000

Cooling loss Incomplete combustion loss

Exhaust loss

260000 20000

240000 220000

0

(b) Energetic analysis

(b) Energetic analysis

Fig. 5. Energy flow analysis at 1660 r/min–105 N m.

Fig. 4. Energy flow analysis at 1660 r/min–420 N m.

ð25Þ

The conversion relationship equations between dry and wet concentration of specie i (y0i and yi) are also used in the calculation, see in Eqs. (26) and (27):

y0i ¼ yi =ð1  y4 Þ .X10 v yi ¼ v i i¼1 i

ð26Þ ð27Þ

3.2.3. Calculation of intercooler loss and heat absorption from ambient The intercooler loss can be obtained from the air internal energy changes before and after intercooler, see in Eq. (28):

Z Q ic ¼ M r;air 

t in;air

t out;air

cv ;air  dt

ð28Þ

where tin,air, tout,air are air temperatures before and after intercooler, K. cv,air is specific heat of air at constant volume, kJ/(kg K). When evaporating, methanol may absorb a large amount of heat from surroundings. If the energy released by intake air due to decline of its temperature cannot satisfy the methanol evaporation, a portion of ambient energy will be absorbed into the control volume. The amount of heat absorption can be calculated by Eq. (29):

Z Q ab ¼ cm  M r;m  Mr;air 

t b;air

t a;air

cv ;air  dt

ð29Þ

where cm is the methanol latent heat of vaporization, kJ/kg. tb,air, ta,air are air temperatures upstream and downstream the methanol injectors.

3.2.4. Calculation of incomplete combustion loss and residual loss In this study, the incomplete combustion loss is inferred from the difference of fuels’ chemical energy and the sum of cumulative heat release of all the four cylinders. Cumulative heat release is calculated from the in-cylinder pressure data using the ‘‘Polytropic Index First Law” (PIFL) heat release model recording in literature [20–22], which has been proved to be best suited for diesel engine thermodynamic analysis. The incomplete combustion loss can be obtained by Eq. (30):

Cooling water temperature difference (°C)

CO2 þ H2  kCO þ H2 O

1660r/min,100% load 1660r/min, 75% load 1660r/min, 50% load 1660r/min, 25% load

9

8

7

6

5

0

10

20

30

40

50

60

Replacement rate (%) Fig. 6. Cooling water temperature difference changing with SP at various engine loads.

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G. Han et al. / Fuel 164 (2016) 393–402

Q ub ¼ ðHu;d  M r;d þ Hu;m  M r;m Þ 

4 X Q ac;i

calculated by the air flow rate multiplied by the temperature difference. In literature [10], the engine surface was divided into seven parts. Temperature of each part was measured by thermal couples. Then the convection and radiation heat in each area can be obtained from the measuring results of air flow through engine by hot-wire anemometer. However, veracity of these methods is limited by the measuring equipment and experiment conditions. To keep the closure of the energy balance equation, the residual loss in this study is calculated by the difference between total input energy and the sum of other energy terms.

ð30Þ

i¼1

where Qac,i is the cycle cumulative heat release of cylinder i, kJ/h. The residual loss is mainly caused by convection and radiation at engine surface and exhaust pipes [23]. There are many different methods to account for the residual loss. Smith et al. [24] built a wooden box around the engine and measured the air flow rate through the box as well as the temperature difference between the entrance and exit of the box. The residual loss therefore is

1400

1660r/min-25% load

1200

1100

1000

1300 1200 1100 1000 900

900 -20

-10

0

10

20

30

40

50

60

800 -20

70

-10

0

10

Crank Angle (deg)

30

40

50

60

70

2000

D mode DMDF mode

1660r/min-75% load

D mode DMDF mode

1660r/min-100% load 1800

Temperature (K)

1600

Temperature (K)

20

Crank Angle (deg)

1800

1400

1200

1000

800 -20

D mode DMDF mode

1400

Temperature (K)

Temperature (K)

1300

1660r/min-50% load

1500

D mode DMDF mode

1600 1400 1200 1000

-10

0

10

20

30

40

50

60

70

800 -20

-10

Crank Angle (deg)

0

10

20

30

40

Crank Angle (deg)

(a) In-cylinder temperatures

28

1660r/min D mode 1660r/min DMDF mode

24

20

16

12

25

50

75

100

Loads (%)

(b) Combustion durations Fig. 7. In-cylinder temperatures and combustion durations of the two running modes at 1660 r/min.

50

60

70

400

G. Han et al. / Fuel 164 (2016) 393–402

2.2

4.1. Energy balance analysis at different loads

2.0

4.2. Characteristics of cooling loss in DMDF mode The results in Section 4.1 showed that the cooling loss in DMDF mode was always lower than that in D mode. To research the characteristics of cooling loss in DMDF mode, a turbine flow meter was installed upstream the entrance of cooling water. Temperature of the cooling water was maintained 80 °C by a solenoid valve installed in the compensatory pipe of the water tank. According to the measuring data of water temperatures in and out of the engine, cooling loss can be obtained by Eq. (8). For this engine, the water pump shaft is connected with crankshaft by straps.

1.8

1.6

1.4

1.2

1.0

5

10

15

20

25

30

35

40

45

Replacement rate (%) Fig. 8. SR changing with SP at different methanol temperatures.

Hence, the cooling water mass flow rate is mainly decided by the engine speed. For a constant engine speed, the cooling water temperature differences can reflect the amount of cooling loss. Instructions of the experiments here are shown in Group 2 of Table 4. Fig. 6 shows the cooling water temperature differences changing with SP at various engine loads. It’s seen from Fig. 6 that, the SP has the maximum value at middle or even higher loads. At low loads, a higher SP makes the combustion unstable and uneconomical. While at high loads, the SP is

900000 800000

Residual loss Incomplete combustion loss Cooling loss Exhaust loss Intercooler loss Output power

Chemical energy of fuel

700000

Energy (kJ)

600000 500000 400000 300000 200000 100000 0

Energy entering the control volume

Energy exiting the control volume

(a) Energy flow distribution Energy changes

300000 Intercooler loss Heat absorption

285000

Energy (kJ)

Previous researches show that SR was always higher than its theoretical value 2.16 at low loads, while it was the opposite result at high loads [6]. To investigate the effect of loads on SR, the 1660 r/ min–420 N m and 1660 r/min–105 N m operating points, which correspond to A100 and A25 of this engine, were selected to conduct the energy balance analysis. Instructions of these experiments are shown in Group 1 of Table 4. The operating parameters in DMDF mode, such as the intake air temperature, cooling water temperature and diesel injection timing, were kept the same with those in the correspondent D mode. Fig. 4 shows the analytical result at 1660 r/min–420 N m working point. The SR is 1.69 in DMDF mode, which is 0.5 lower than the theoretical value. As mentioned before, zero enthalpy of the gases are based on 298.15 K in this study, while the intake air temperature at the entrance of the control volume was also kept around 298 K, so the absolute enthalpy value of intake air was quite small. When injected into the intake port, methanol will evaporate to form the methanol/air mixture. Therefore, a large amount of heat will be absorbed due to the high latent heat of vaporization for methanol, which leads to a certain amount of temperature drop of the intake air. If the mixture temperature is lower than 298.15 K, then a negative methanol steam enthalpy is obtained. The total intake enthalpy is the sum of intake air and methanol steam enthalpies. In fact, the absolute value of total intake enthalpy is always small (even cannot be displayed in the energy flow distribution figure) and can be ignored in the energy balance analysis. Fig. 4(a) shows the energy flow distribution of the two running modes at 1660 r/min–420 N m operating point. The starting point of dotted line in Fig. 4(b) is the sum of output power and fuel chemical energy difference between D and DMDF mode. Then this term subtracts the other terms’ difference successively. The solid line gained finally is the output power of D mode. We can see from Fig. 4(b) that, the exhaust loss, cooling loss and incomplete combustion loss in DMDF mode are lower than those in D mode. Furthermore, a certain amount of ambient energy is absorbed into the control volume due to methanol evaporation. All these contribute to thermal efficiency improvement of DMDF mode, and the reduction of cooling loss is the dominate factor. Fig. 5(a) shows the energy flow distribution of the two running modes at 1660 r/min–105 N m operating point. The SR is 2.67 in DMDF mode, which is 0.5 higher than the theoretical value. We can see from Fig. 5(b) that the cooling loss in DMDF mode is still lower than that in D mode. However, the exhaust loss, residual loss and incomplete combustion loss in D mode are lower than those in DMDF mode. In addition, for the reason of low methanol mass flow rate, there is no ambient heat absorption at low loads. All these contribute to higher SR than theoretical value, and the increase of incomplete combustion loss in DMDF mode is the most important factor.

Replacement ratio

4. Results and discussion

270000

Exhaust loss Cooling loss

Incomplete combustion loss Residual loss

255000

240000

225000

(b) Energetic analysis Fig. 9. Energy flow analysis at 30 °C and 60 °C methanol temperatures while keeping a 10% SP.

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limited by maximum explosion pressure [6]. Cooling water temperature differences decrease with rising SP at various engine loads. In other words, the cooling loss of DMDF mode is always lower than that in D mode, and it goes even lower as SP rises. Fig. 7 gives the calculated in-cylinder temperatures and combustion durations of the two running modes at 1660 r/min with various engine loads. The operating points in DMDF mode showed here are the maximum SP points at each load in Fig. 6. We can see that, at low loads, lower in-cylinder temperature in DMDF mode is the cause of decreasing cooling loss. While, at high loads, the reduction in combustion duration is the main reason for decreasing cooling loss, in spite the fact that the in-cylinder temperature is higher in DMDF mode. 4.3. Energy balance analysis in different methanol temperatures The quality of methanol/air mixture affects dual-fuel combustion process. Meanwhile, methanol temperature has substantial impact on the formation of methanol/air mixture. Previous observing results show that when the methanol supplying pipeline is close to the exhaust pipe, or the methanol tank is placed under blazing sun for a long time, the economic efficiency of DMDF engine will be improved. Therefore, it is necessary to investigate the effect of methanol temperature on the energy flow distribution in DMDF mode. The methanol temperature was controlled by a combination of heat exchanger and heat tap installed on the methanol supplying pipeline. Then the rising replacement rate experiments were

conducted at 1660 r/min–420 N m operating point with various methanol temperatures. Instructions of the experiments here are shown in Group 3 of Table 4. Fig. 8 gives the experimental results of SR changing with rising SP at different methanol temperatures. It’s seen from Fig. 8 that SR decreases as methanol temperature rises at the same SP. When the methanol temperature is lower than 50 °C, the SR decreases with rising SP. However, with a 60 °C methanol temperature, the SR approximates to 1.1 at low SP, and then it goes up as SP rises. Fig. 9 shows the energy flow distributions at 30 °C and 60 °C methanol temperatures with 10% SP. We can see that, although the methanol temperature has a great effect on SR, as shown in Fig. 8, the total input energy has little difference because the SP here is too small. A significant reduction in incomplete combustion loss at 60 °C methanol temperature is the main cause of SR decline. Fig. 10 shows the energy flow distribution at 30 °C and 60 °C methanol temperatures with 30% SP. Compared with results shown in Fig. 8, the effect of methanol temperature on SR is weaker, so does the difference in total input energy. The decline of SR at higher methanol temperature is also the result of reduction in incomplete combustion loss.

4.4. Energy balance analysis in different replacement rates We can see from Fig. 8 that the SR changing with SP showed different characteristics at various methanol temperatures. Fig. 11 gives the energy flow distribution of the DMDF engine at 50 °C and 60 °C methanol temperatures with rising SP.

900000 Heat absorption Chemical energy of fuel

700000

Residual loss Incomplete combustion loss Cooling loss Exhaust loss Intercooler loss Output power

800000

600000

700000

500000

600000

Energy (kJ)

Energy (kJ)

Residual loss Incomplete combustion loss Cooling loss Exhasut loss Intercooler loss Output power

500000 400000 300000

400000 300000 200000

200000

100000

100000 0 Energy entering the control volume

0

Energy exiting the control volume

(a) Energy flow distribution 300000

D

(a) 50

Energy changes

11% 19% 31% 40%

methanol temperature Residual loss Incomplete combustion loss Cooling loss Exhasut loss Intercooler loss Output power

700000

Intercooler loss

600000

Heat absorption Exhaust loss

Cooling loss

500000 270000

Incomplete combustion loss Residual loss

255000

Energy (kJ)

Energy (kJ)

285000

400000 300000 200000 100000

240000

0

(b) Energetic analysis Fig. 10. Energy flow analysis at 30 °C and 60 °C methanol temperatures while keeping a 30% SP.

D

(b) 60

10% 19% 31% 38%

methanol temperature

Fig. 11. Energy flow distribution at 50 °C and 60 °C methanol temperatures with rising SP.

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Referring to Fig. 8, we can infer from Fig. 11(a) that the thermal efficiency of the engine improves with rising SP. This is the result of the exhaust loss, cooling lose and incomplete combustion loss all decrease linearly with rising SP. At 60 °C methanol temperature, the SR increases with rising SP. But the results in Fig. 11(b) show that, the total input energy also decreases with rising SP, which means the thermal efficiency of the engine is improved. This can be explained that the input energy increase caused by SR rises is weaker than the input energy decrease caused by SP rises, for SR is always lower than theoretical value at high loads. Meanwhile, the cooling loss, exhaust loss and residual loss all decrease with rising SP at 60 °C methanol temperature, but the incomplete combustion loss shows the first and then decreases characteristics.

5. Conclusions The theoretical and experimental investigations on energy balance on DMDF engine were conducted. The following conclusions can be drawn from this study: (1) A mathematic algorithm basing on the experimental data was propose, which can be used to conduct the energy balance calculation of both diesel fuel and diesel/methanol dual fuel engines. (2) The SR is lower than its theoretical value 2.16 at high loads. This can be explained that the exhaust loss, cooling loss and incomplete combustion loss in DMDF mode are all lower than that in D mode at high load conditions. Furthermore, a certain amount of ambient energy is absorbed into the control volume in DMDF mode. Among all those reasons above, the reduction of cooling loss is the dominate factor. (3) The SR is higher than its theoretical value 2.16 at low loads. The cooling loss in DMDF mode is still lower than that in D mode. However, the D mode has lower exhaust loss, residual loss and incomplete combustion loss than the corresponding DMDF mode. In addition, for the reason of low methanol mass flow rate, there is no ambient heat absorption at low load conditions. Among all these reasons, the increase of incomplete combustion loss in DMDF mode is the most important factor. (4) In the comprehensive engine operating range, the cooling loss of DMDF mode is always lower than that of D mode, and it goes even lower as SP rises. (5) The SR decreases with rising methanol temperatures at the same SP, which is the result of decreasing incomplete combustion loss.

Acknowledgments The authors acknowledge the financial support from the Natural Science Foundation Committee of China (Contract No. 51336005).

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