Thermodynamic, performance and emission investigation of a diesel engine running on dimethyl ether and diethyl ether

International Journal of Thermal Sciences 50 (2011) 1594e1603

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International Journal of Thermal Sciences journal homepage: www.elsevier.com/locate/ijts

Thermodynamic, performance and emission investigation of a diesel engine running on dimethyl ether and diethyl ether _ Ismet Sezer* Mechanical Engineering Department, Gümüs¸hane University, 29100 Gümüs¸hane, Turkey

a r t i c l e i n f o

a b s t r a c t

Article history: Received 12 August 2010 Received in revised form 24 March 2011 Accepted 26 March 2011 Available online 6 May 2011

This study investigates the use of dimethyl ether and diethyl ether in diesel engines as alternative fuels. A direct injection diesel engine was simulated via a thermodynamic cycle model for investigation. Thermodynamic and performance parameters besides emissions determined and compared for diesel, dimethyl ether and diethyl ether fuels at two different states. The results showed that dimethyl ether and diethyl ether presented a lower cylinder temperature and pressure, and thus a lower engine performance than diesel fuel for the equal injection conditions. The brake power declines about 32.1% and 19.4% at 4200 rpm while brake specific fuel consumption increases about 47.1% and 24.7% at 2200 rpm for dimethyl ether and diethyl ether, respectively. Engine performance for dimethyl ether and diethyl ether extensively improves for the same equivalence ratio condition, but a more amount of fuel is needed about 64% for dimethyl ether and 32% for diethyl ether. The gains in the brake power by dimethyl ether and diethyl are about 13.6% and 6% at 4200 rpm compared to diesel fuel. The brake specific fuel consumption is also higher about 43.5% for dimethyl ether and 23.6% for diethyl ether than diesel fuel. The brake thermal efficiency for dimethyl ether and diethyl is generally better than diesel fuel. The lower carbon dioxides are obtained by dimethyl and diethyl ethers at all conditions, while carbon monoxide and nitrogen oxide are slightly higher for dimethyl and diethyl ethers at equal equivalence ratio condition. Ó 2011 Elsevier Masson SAS. All rights reserved.

Keywords: Alternative renewable fuels Dimethyl ether Diethyl ether Diesel engine Engine performance Fuel consumption Exhaust emissions

1. Introduction Diesel engines have been widely used in recent decades as an alternative power source for light-duty vehicles because of the economical and environmental reasons. Therefore, the global diesel fuel consumption has increased with the pollutions sourced from diesel engines [1,2]. The unburned or partially burned (total) hydrocarbon (THC) emissions, smoke (soot) or particulate matter (PM), nitrogen oxides (NOx), sulfur oxides (SOx) emitted from compression ignition (CI) engines and particularly carbon dioxide (CO2) create severe environmental problems [2e4], which have been tried to be reduced by the stringent emission legislations. The different alternatives such as the investigation of viable alternative fuels and the reformulation of conventional fuels have been evaluated for meeting the emission standards and future energy demand [5,6]. The reformulation of diesel fuel contains the reduction of the sulfur and aromatic contents or the oxygen addition to diesel fuel [7]. A lot of works have been performed to show the effects of using alternative diesel fuels and additives including synthetic diesel fuels, biodiesels, alcohols and ethers. * Tel.: þ90 456 233 74 25; fax: þ90 456 233 74 27. E-mail address: [email protected]. 1290-0729/$ e see front matter Ó 2011 Elsevier Masson SAS. All rights reserved. doi:10.1016/j.ijthermalsci.2011.03.021

The synthetic or gas-to-liquids (GTL) diesel fuels are very low sulfur and aromatic containing alternatives to conventional diesel fuel. The GTL diesel fuels are produced from wide variety of carbon-based feedstocks, including coal, natural gas, biomass and oil sands via a FischereTropsch process [8]. The synthetic diesel fuels have also very high cetane number and excellent autoignition characteristics. These fuels offer significant reductions in PM and NOx emissions [9,10]. The GTL fuels can be used neat [9] or blended in any proportions with conventional diesel fuel [10] without any engine modification. However, further research on synthetic fuel production technologies is essential for lower costs to make it more economically competitive with conventional fuels [11]. Biodiesels are one of the oxygen containing and sulfur free alternatives to petroleum-based diesel fuel, which are the fatty acids (or triglycerides) obtained from various straight vegetable oils [12e16] or the recycled waste oils [17e19] and animal fats [20e22]. Neat vegetable oils can be directly used in diesel engines because of a high cetane number and very close calorific value to diesel [5]. However, it was reported that they leaded to operational problems such as a lower engine power and efficiency, engine deposits under long-term use because of the high viscosity and low volatility [23,24]. Several methods to reduce the high viscosity of vegetable oils or

_ Sezer / International Journal of Thermal Sciences 50 (2011) 1594e1603 I.

Nomenclature b bmep bsfc C Cp f h hconv Fs Lcr LVHf m _ m N p P Q T u U v V Vd W Z

bore, m brake mean effective pressure, bar brake specific fuel consumption, g/kWh blowby coefficient, dimensionless specific heat at constant pressure, J/kg K residual gas fraction, dimensionless enthalpy, J/kg convective heat transfer coefficient, W/m2K stoichiometric fuel ratio, dimensionless connecting rod length, m lower heating value (or calorific value) of fuel, kJ/kg mass, kg mass flow rate, kg/s engine speed, rpm pressure, bar power, kW or BG heat or heat transfer, J temperature, K specific internal energy, J/kg internal energy, J specific volume, m3/kg volume, m3 cylinder displacement volume, m3 work, J cylinder number, dimensionless

animal fats to enable their use in diesel engines without any operational problem have been applied: blending with hydrocarbons (dilution), micro-emulsification (co-solvent blending), pyrolysis (craking), and transesterification [24e28]. The dilution of vegetable oils can be accomplished with such materials as petroleum diesel fuels, alcohols or solvent. It was reported that this approach created similar engine problems to those of neat vegetable oils and a longterm using of vegetable oil/petroleum-diesel blends were not suitable in direct injection diesel engines [25e28]. Micro-emulsification is another technique to produce biodiesel. The components of a biodiesel micro-emulsion include diesel fuel, vegetable oil or animal fat, alcohol, surfactant and cetane improver in suitable proportions. Alcohols such as methanol, ethanol and propanol are used as viscosity lowering additives, higher alcohols are used as surfactants and alkyl nitrates are used as cetane improvers. Viscosity reduction, increase in cetane number and good spray characters are achieved with the micro-emulsions but prolong usage causes problems like injector needle sticking, carbon deposit formation and incomplete combustion [25e29]. Pyrolysis or cracking is the process of a chemical change via heat or with the aid of catalyst. It involves heating in the absence of air or oxygen and cleavage of chemical bonds to yield small molecules. This process can be applied to vegetable oils, animal fats, natural fatty acids and methyl esters of fatty acids. This technique presents lower viscosity, flash point, and pour point than diesel fuel besides equivalent calorific values. Disadvantages of this process include high equipment cost and need for separate distillation equipment for separation of various fractions. Additionally, the product obtained was similar to gasoline with a lower cetane number containing sulfur, ash and carbon residue which makes it less ecofriendly [25e28]. The most popular method of producing biodiesel is the transesterification. In this process as shown in Eq. (1), the methyl esters and glycerine are formed through the reaction of fatty acids (or triglycerides) with methanol in the presence of a catalyst

1595

Greek letters equivalence ratio, dimensionless efficiency, % crank angle, CAD start of injection crank angle, CAD injection duration, CAD density, kg/m3 torque, Nm angular speed, 1/s or rad/s

f h q qs qd r s u

Subscripts a air b burned e effective (brake) f fuel IVC intake valve closed i injection L loss v volumetric w wall Abbreviations CAD crank angle degree CR compression ratio ER equivalence ratio MFI mass fuel injection MBT maximum brake torque AFRs stoichiometric airefuel ratio

such as sodium hydroxide (NaOH) or potassium hydroxide (KOH) [3,17,25e2830].

CH2 OOR I

CH2 OH I

CHOOR þ 3CH3 OH I

Catalyst





!

3CH3 OOCR þ CHOH I

CH2 OOR |fflfflfflfflfflffl{zfflfflfflfflfflffl}

CH3 OH |fflfflfflffl{zfflfflfflffl}

CH3 OOCR |fflfflfflfflfflfflffl{zfflfflfflfflfflfflffl}

CH2 OH |fflfflfflffl{zfflfflfflffl}

Triglyceride

Methanol

Methyl esterðBiodieselÞ

Gylcerol

ð1Þ

The alcohols other than methanol have also been used to generate various esters, i.e. the ethyl, propyl, and butyl esters [31]. Consequently, biodiesel is technically competitive with or offer several technical advantages compared to conventional diesel. Besides being a renewable and domestic resource, biodiesel reduces most emissions while engine performance and fuel economy are nearly identical compared to conventional diesel fuel [32]. Some problems, however, have impaired the widespread use of biodiesel. They are related to the economics [12] and properties of biodiesel such as lower volatility and higher viscosity. Furthermore, problems related to combustion and emissions especially NOx remain to be solved [23]. Alcohols and ethers are other alternatives as oxygenated fuels or fuel additives for diesel engines. Alcohols produced fossil or renewable resources such as methanol [6,33], ethanol [34e36] etc. [37] are generally added to diesel fuel and the alcohol content of up to 20% can be used in compression ignition (CI) engines by little or no modification [5,33]. It is declared that alcoholediesel blends create to reduction especially smoke or particulate emissions [6,33,34e36], but the effects on carbon monoxide (CO), THC and NOx, which are influenced by a number of factors such as engine

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fuel metering technology, exhaust control technology, test procedure, and test conditions [34,38] are less clear; some studies also reported decrease in CO and NOx [6,34,35] different from THC emissions. It is also stated that alcohols have negative effects on cetane number and also energy content of diesel fuel, which cause to reduction in power output and increment of fuel consumption. Therefore, both cetane improver and lubricity additives might be needed when adding alcohols to diesel [35,36,39]. Ethers originated from treatment of alcohols with strong dehydrating agents, chiefly methyl tertiary-butyl ether (MTBE) and ethyl tertiary-butyl ether (ETBE) can be used as oxygen additives in diesel fuel. The use of such etherediesel blends in CI engines creates similar performance and emissions results to alcohols [40e42]. Among the oxygenated alternatives dimethyl ether (DME) and diethyl ether (DEE) are regarded as one of the promising alternative fuels or an oxygen additive for diesel engines with its advantages of a high cetane number and oxygen content [43,44]. The other attractive characteristics of DME include rapid evaporation, smokefree and low noise level combustion, virtual non-toxicity and being in general environmental friendly [45]. DME is a pure simple oxygenate having similar physical properties to liquefied petroleum gas (LPG), hereby also referred as synthetic LPG. It condenses at 25  C under atmospheric pressure, or under 5e6 bar at ambient temperature. Therefore, DME can be handled similarly as LPG so existing tankers and receiving terminals of LPG can be easily converted for DME distribution [43e45]. The production of DME is very similar to that of methanol; DME is produced through gasification of various renewable substances or fossil fuels, i.e. natural gas, coal, bio mass or other carbon containing materials [44,46]. However, the current production technology is costly and relatively energy inefficient. As well, some engine optimization is required to take advantage of the fuel’s properties. Conventional fuel injection system has to be modified due to the intrinsic properties of the DME [47]. Diethyl ether (DEE) is also the most similar to DME except for DEE is liquid at the ambient conditions, which makes it attractive for fuel storage and handling. DEE is produced from ethanol by dehydration process so it is a renewable fuel [48]. DEE has several favorable properties, including exceptional cetane number, reasonable energy density, high oxygen content, low autoignition temperature and high volatility. Therefore, it can be assist to improving of engine performance and reducing the cold starting problem and emissions when using as a pure or an additive in diesel fuel [49,50]. Consequently, both of DME and DEE are promising alternatives for diesel engines due to suitable fuel properties cited above and also given in Table 1. However, only limited number of studies performed on using of DME and DEE in CI engines [43e50]. Very few of works for DME [45,51] and DEE [52] were performed on investigation of engine performance. Additionally, DME and DEE have not been compared each other in any study.

2. Thermodynamic cycle model A thermodynamic diesel engine cycle model originally developed by Ferguson [53] adapted and used for the calculation of thermodynamic and performance parameters. The thermodynamic cycle model includes the only closed part of the engine cycle. 2.1. Fuel injection model The fuel injection rate is determined empirically from the fictitious fuel injection rate

    _ fi u q  qs n1 m ðq  qs Þ ¼ exp qd GðnÞ qd qd mfi

Property

Diesel

DME

DEE

Chemical formula Molecular weight Density of liquid at NTPa (kg/L) Viscosity at NTPa (cP) Oxygen content (wt %) Sulfur content (ppm) Boiling temperature ( C) Autoignition temperature in air ( C) Flammability limit in air (vol %) Stoichiometric airefuel ratio (AFRs) Heat of vaporization at NTPa (kJ/kg) Lower heating value (MJ/kg) Cetane number (CN)

CxHy 190e220 w0.84 2.6 e w250 180e360 315 0.6e6.5 14.6 250 42.5 40e55

C2H6O 46 0.668 e 34.8 e 24.9 235 3.4e17 9 460 (20  C) 28.4 55e66

C4H10O 74 0. 710 0.23 21 e 34.6 160 1.9e9.5 11.1 356 33.9 >125

NTP: Normal temperature and pressure.

(2)

where, qs, qd and mfi are start of fuel injection, measure of injection duration and total mass of injected fuel. The gamma function G(n) in Eq. (2) is determined from the asymptotic formula for n > 1.

  1 1 1 1 lnðnÞ  n þ lnð2pÞ þ lnGðnÞ ¼ n   2 2 12n 360n3 1 1 þ  1260n5 1680n7

ð3Þ

The Eq. (2) can be also used for the determination of the arbitrary heat release (AHR) as follow.

AHR ¼

_ fi 1m

u mf i

Qtotal;fi

(4)

2.2. Mass change in the cylinder The rate of change of burned fuel and air in the cylinder are determined as follows

  _ fF m dmf 1 _ fi  L s m ¼ u 1 þ fFs dq dma ¼ dq

Table 1 Properties of the fuels.

a

Therefore, this study was devoted to the use of DME and DEE in CI engines and effects of these fuels on thermodynamic and performance parameters were investigated comparatively. For this purpose, a commercial suction (naturally-aspirated) direct injection (SDI) automobile diesel engine modeled by using a thermodynamic cycle model. Thus, thermodynamic and performance parameters besides emissions such as CO2, CO and NO were determined and compared for diesel, DME and DEE fuels.

 

_ L =u m 1 þ fFs

(5)

 (6)

_ L ¼ Cm and equivalence ratio at any time is here, the blowby is m f ¼ mf =Fs ma . Obviously, if the blowby is neglected, i.e. C ¼ 0.0, the burning rate will be equal to injection rate in Eq. (5). The total mass in the cylinder at any time is

m ¼ ma þ mf

(7)

2.3. Energy equation For a closed system, the energy equation that is basically first law of thermodynamics applied to the cylinder contents gives

_ Sezer / International Journal of Thermal Sciences 50 (2011) 1594e1603 I.

_ h _ h m dU dV m Q_ ¼  L p  L Lþ fi f u u u dq dq

2.4. Engine performance parameters

(8)

Here,

the enthalpy of fuel is determined 1 0 0 ashf ¼ hf þ ðpf  patm Þ. hf , rf and pf are the enthalpy of forma-

rf

tion at atmospheric pressure, density and saturation vapor pressure of the fuel, respectively. The cylinder gas temperature and pressure are determined by NewtoneRaphson iteration as follow.

 T ¼T  p ¼ p

1597

    

   F vlnv vlnv vlnv D v 1 þF2 þ vlnT p vlnp T p vlnp T      

  Cp T pv vlnv F vlnv D F2   1  pv T vlnT p p vlnT p

(9)

The engine performance parameters are determined from widely known relations given below. Brake mean effective pressure is defined as follow

imep ¼

Wi Vd

bmep ¼ imep  fmep

      Cp T v2 vlnv vlnv D¼ þ  pv T vlnp T vlnT p        vlnv vlnv vlnv ; þ þ vlnT p vlnT p vlnp T

(18)

The brake power is defined as

Pe ¼

Here;

(17)

here, fmep is friction mean effective pressure which is determined following correlation

fmep ¼ 0:97 þ 0:15ðN=1000Þ þ 0:05ðN=1000Þ2 (10)

(16)

bmepVd NZ k60

(19)

where, k is a constant which is 1 for two-stroke engines and 2 for four-stroke engines. Brake torque is defined as

se ¼

Pe

u

(20)

where, u is the angular speed that can be expressed as

F1 ¼ uð1Þ  u

u¼

and

F2 ¼ vð1Þ  v

2pN 60

(21)

Brake thermal efficiency is defined as the ratio of the brake power to the energy of the supplied fuel to the cylinder which is expressed as

Initial conditions were taken as p1 ¼ 1 bar and T1 ¼ 325 K as suggested by Ferguson [53]. In Eq. (8), the second term at right hand side expresses work output

he ¼

dW dV ¼ p dq dq

where, mf is the mass fuel injection and LHVf is the lower heating value of the fuel. The brake specific fuel consumption represents the ratio of mass fuel injection to per unit brake power output and it is expressed as

(11)

Pe mf LHVf

(22)

The cylinder volume is determined as follow



VðqÞ ¼ V0 1þ



 1

CR 1 1cosq þ 3 2

1  2 1 1 32 sinq 2 

bsfc ¼ (12)

Here, V0 is clearance volume and 3 ¼ S/2Lcr is the half of the stroke to connecting rod length. One can easily derive from Eq. (12) (dV/dq) which is used as in Eq. (11). Assuming a cylindrical combustion chamber, the heat loss rate in Eq. (8) is determined

! pb2 4V _ Q L ¼ hconv þ ðT  Tw Þ 2 b

(13)

In Eq. (10), hconv, convective heat transfer coefficient is determined from Woschni correlation as follow

hconv ¼ 3:26b0:2 p0:8 T 0:55 w0:8

(14)

In Eq. (11), w is average gas velocity which is determined from

 w ¼

c1 Sp þ c2

Vd TIVC ðp  pm Þ pIVC VIVC

 (15)

here, Sp and pm are mean piston speed and motored pressure, c1 and c2 are constants. During compression c1 ¼ 2.28 and c2 ¼ 0; during combustion and expansion c1 ¼ 2.28 and c2 ¼ 3.24  103 were used. Further details of cycle model can be found in Ferguson [53].

mf 3600 Pe

(23)

2.5. Equilibrium combustion products The reacting species in any engine combustion includes the radicals of C, H, O, N compounds, however it has been shown that for equivalence ratio below 2 which is the case of most diesel engines, the 10 species are considered. The combustion reaction for any fuel with oxygen is given by;

3fCc Hh Oof Nn þ ð0:21O2 þ 0:79N2 Þ/n1 CO2 þ n2 H2 O þ n3 N2 þ n4 O2 þn5 CO þ n6 H2 þ n7 H þ n8 O þ n9 OH þ n10 N (24) To determine the mol fraction of combustion products 10 equations are required. Four of the equations are obtained from the mass balance of C, H, O and N. The other six equations are obtained from equilibrium constants for dissociations reactions. After linearization of the nonlinear equations via NewtoneRaphson method, Gaussian-elimination method is used for the solution of linear algebraic equations [53]. 3. Validation of the model To demonstrate the reliability of the thermodynamic cycle model, the predicted values were compared with experimental

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data for diesel [54], DME [55] and DEE [52] in Fig. 1(a)e(e) at the conditions specified on the figures and engine specifications in Table 2. The predictions are in good agreement with the experimental data as seen in the figures. The deviations are about 0.7e14% for pressure in Fig. 1(a), 0.2e3.8% for brake power in Fig. 1(b), 0.4e11% for pressure in Fig. 1(c), 0e1.5% for brake specific fuel consumption in Fig. 1(d) and 0.3e13% for normalized heat release rate in Fig. 1(e). Therefore, it can be said that the thermodynamic cycle model has a sufficient level of confidence for parametric investigations and engine performance analysis.

Table 2 The main specifications of the engines.

4. Results and discussion

The main specifications of the engine (Engine IV) can be seen in Table 2. The variation of thermodynamic parameters such as pressure, temperature, work output and heat loss for two different cases are shown in Fig. 2. At the first case in Fig. 2(a), the mass fuel

S

Lcr

CR

Type of engine

100

115

230

18

Engine II [55]

114

135

270

18

Engine III [52]

102

120

192

17.9

191

19.5

One-cylinder 903cc Experimental Engine Six-cylinder 8270 cc Turbo-Intercooler DI Engine One-cylinder 980 cc DI Experimental Engine Four-cylinder 1896cc Volkswagen 1.9 SDI Engine

Engine IV

A commercial suction (naturally-aspirated) direct injection (SDI) automobile diesel engine was used for the parametric investigation. 80

b Engine I [54]

79.5

11

N= 2000 rpm θs=-25.5 CAD

95.5

DIESEL Thermodynamic model Experimental [54]

10

Brake power, kW

Pressure, bar

60

40

9

8

20 7

Diesel Thermodynamic model Experimental [54]

0

6

-20

0

20

40

60

1400

Crank angle, CAD

120

a N= 1400 rpm θs=-9 CAD

80 60 40 DME Thermodynamic model Experimental [55]

0 -20

0

20

40

2000

2200

2400

60

b 240

DME Thermodynamic model Experimental [55]

220

200

180 800 1000 1200 1400 1600 1800 2000 2200

80

Engine speed, rpm

Crank angle, CAD

c

d

0.05

Normalized heat release rate, -

Pressure, bar

100

20

1800

Engine speed, rpm Brake specific fuel consumption, g/kWh

140

1600

N= 1500 rpm θs= MBT

0.04 0.03 0.02 0.01 DEE

0

Thermodynamic model Experimental [52]

-0.01 -10

0

10

20

30

40

50

Crank angle, CAD

e Fig. 1. Comparison of various predicted and experimental data.

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1599

Fig. 2. Variation of thermodynamic parameters for various conditions.

injection (MFI) is fixed by changing equivalence ratio, which means that no changes in fuel injection rate (or fuel injection equipment). As seen in the figure, the slightly narrower trace reaching a lower peak pressure and lower temperature values are obtained with DME and DEE due to lower heating value of DME and DEE, as in literature [55]. DME gives the lowest pressure and temperature values; the reductions in peak pressure are about 15.4% for DME and 9.6% for DEE in comparison with diesel. The reductions in cylinder pressure result in lower work output about 27.5% and 16.9% for DME and DEE. Similarly, the decrease in temperature cause to reduction

of heat loss about 35.1% and 22.8% for DME and DEE, respectively. At the second case in Fig. 2(b), the equivalence ratio is fixed, therefore MFI increases 64% for DME and 32% for DME due to having of DME and DEE the lower stoichiometric airefuel ratio (AFRs) than diesel fuel as seen in Table 1. In that situation, cylinder pressure and temperature significantly increase compared to the first case and slightly higher values than diesel fuel, as in literature [51]. The highest pressure and temperature values are obtained with DME; the increments in peak pressure are about 5.6% and 2.4% for DME and DEE. That increment in pressure generates improvement in

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1600

Table 3 Energy distribution values for various conditions (N ¼ 2600 rpm, qs ¼ MBT). Parameter

Diesel @ f ¼ 0.73

DME @ f ¼ 0.45

DEE @ f ¼ 0.56

DME @ f ¼ 0.73

DEE @ f ¼ 0.73

Fuel energy, J Heat loss, J Indicated work, J Friction work, J Exhaust energy, J

1168.6 282.4 496.8 80.5 308.9

774 183.2 (23.6%) 360.3 (46.5%) 80.5 (10.4%) 150 (19.5%)

934 217.9 412.4 80.5 223.2

1244.1 314.8 547.6 80.5 301.2

1217.4 295.5 519.2 80.5 322.2

(24.1%) (42.5%) (6.9%) (26.5%)

work output about 10.2% and 4.5% for DME and DEE. The increments in heat loss are the level of 11.5% and 4.6% for DME and DEE, respectively. For the both conditions given above, energy distribution values and their percentages into fuel energy are given in

(23.3%) (44.1%) (8.6%) (24%)

(25.3%) (44%) (6.4%) (24.3%)

Table 3. Work output values of DME and DEE are lower than those of diesel for the same mass fuel injection as given in Table 3 but the portion of work output higher and parts of heat and exhaust energy losses are lower compared to diesel fuel. Moreover, work output

0.8

38

DME@ ER 1

Brake thermal efficiency, %

DME@ ER 2

Equivalence ratio

0.7 ER 1 0.6

ER 3

0.5

0.4

ER 2

0.3

36 DEE@ ER 1 DIESEL@ ER 1

34

DEE@ ER 3

32

30

28 1000

1800

2600

3400

4200

5000

1000

Engine speed, rpm

1800

2600

DME@ ER 1

160

5000

DME@ ER 1

DEE@ ER 1

DEE@ ER 1 140

8

Brake torque, Nm

Brake mean effective pressure, bar

4200

b

DIESEL@ ER 1

DEE@ ER 3 6

Catalog value

120 DIESEL@ ER 1 100

DEE@ ER 3

80

DME@ ER 2

DME@ ER 2

60

4 1000

1800

2600

3400

4200

1000

5000

1800

DME@ ER 1

50

DEE@ ER 1 Catalog value

DIESEL@ ER 1

40

30 DEE@ ER 3

20

3400

4200

5000

d

10 DME@ ER 2 0

Brake specific fuel consumption, g/kWh

c 60

2600

Engine speed, rpm

Engine speed, rpm

Effective power, kW

3400

Engine speed, rpm

a 10

450

400

350

DME@ ER 2

DME@ ER 1 DEE@ ER 3

300

DEE@ ER 1

250

DIESEL@ ER 1 200

1000

1800

2600

3400

Engine speed, rpm

4200

5000

(24.2%) (42.6%) (6.6%) (26.6%)

1000

1800

2600

3400

Engine speed, rpm

e Fig. 3. Variation of engine performance parameters.

f

4200

5000

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portions of DME and DEE are better than diesel even when the MFI is increased for DME and DEE. These variations can be attributed to suitable combustion characteristics of DME and DEE, i.e. lower auto ignition temperature, rapid evaporation, simple chemical structure, lower carbon fraction and oxygen content. Fig. 3 shows the variation of engine performance parameters such as brake thermal efficiency, mean effective pressure, brake torque, brake power and brake specific fuel consumption for two different injection scenes. The variation of equivalence ratio respect to engine speed is given in Fig. 3(a); equivalence ratio is fixed as ER2 and ER3 for DME and DEE to be equal mass fuel injection at first scene. The second scene equivalence ratio is fixed as ER1 for all of the fuels. Fig. 3(b) shows the variation of brake thermal efficiency (bte); bte is slightly higher than diesel for DME and DEE up to 3000 rpm and then it sharply decreases at the first scene. However, at the second scene DEE and especially DME give higher bte than diesel for overall engine speed, as in literature [51]. The maximum gains in bte with DME and DEE are about 5.1% and 1.1% at 2200 rpm. The main reason of improvement in bte is thought to be better combustion efficiency of DME and DEE; a more part of fuel energy is converted to work with DME and DEE as given in Table 3 because of the improved combustion efficiency sourced from their good combustion characteristics as cited above, although they have a lower LHV than diesel fuel. Additionally, the lower heat loss in Fig. 2(a) for the first scene and the higher work output in Fig. 2(b) for the second scene are considered an additional contribution to development of bte. However, when the fueleair mixture becomes too lean to auto-ignite or support a propagating flame as in the first scene the hydrocarbon remains unconsumed in a compression ignition engine, which results in reduction in bte. Fig. 3(c) shows the variation of brake mean effective pressure (bmep) according to engine speed; bmep values belongs to DME and DEE are dramatically lower those of diesel for the equal injection mass. That reduction can be attributed to cylinder pressure drop of these fuels given in Fig. 2(a). The decrements are about 31.2% and 19.9% for DME and DEE at the engine speed of 2200 rpm. On the other hand, bmep of DME and DEE for ER1 become higher than that diesel by growing significantly. The increments for DME and DEE are about 11.9% and 5.3% compared to diesel at 2200 rpm. Improving of bmep can be attributed to enlarged fuel energy with more fuel injection and better combustion efficiency. The oxygen content of DME and DEE might also partially contribute on the increase of bmep, because oxygen presence in the fuel improves the combustion efficiency by helping to homogenization of the fueleair mixture in the cylinder. Very similar variations in brake torque to bmep appear in Fig. 3(d). The catalogue brake torque value is also given in the figure to indicate the reality of the computation. The variations in brake power are shown in Fig. 3(e); brake power increases up to 4200 rpm then slightly decrease for all fuels. The brake power of DME and DEE for ER2 and ER3 are very lower about the level of 32.1% and 19.4% at 4200 rpm than that of diesel due to the lower calorific value and smaller fuel delivery amount of DME and DEE comparison with diesel, as seen in Fig. 2(a). However, they are located at higher level than diesel especially at high speeds (2600rpm and above) for ER1 by enlarging fuel supply as seen in Fig. 2(b), as in literature [55]. The increments in brake power for DME and DEE are about 13.6% and 6% at the engine speed of 4200 rpm. As shown in Fig. 3(f), brake specific fuel consumption (bsfc) reaches the minimum values between 1800 and 2600 rpm then increases for all fuels. This can be explained by the following reason: when the speed increases, the fuel supply increases due to the more working cycles in a specific period of time. DME gives the maximum bsfc values while diesel fuel has the minimum bsfc values for both of two cases, similar to literature [55]. The increments in for the first scene are about 47.1% and 24.7% for DME and DEE at 2200 rpm. The bsfc is slightly decreases for ER1 for

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12

10

DIESEL @ φ = 0.73 DME @ φ = 0.45 DEE @ φ = 0.56 DME @ φ = 0.73 DEE @ φ = 0.73

N = 2600 rpm θ s= MBT

8

6

4

2

0 CO2 (%)

CO (%)

NO (ppm)

Fig. 4. Exhaust emissions for various conditions.

DME and DEE but bsfc is still higher about 43.5% and 23.6% for DME and DEE in this case. The increments in bsfc can be attributed to lower energy density of DME and DEE. The emission values obtained from chemical equilibrium of combustion products for various conditions are shown in Fig. 4. The CO2, CO and NO emissions are evaluated for diesel, DME and DEE fuels. CO2 for DME and DEE is lower than diesel fuel for all conditions due to simpler chemical structure and lower carbon content of DME and DEE. As seen from the figure, CO and NO emissions become higher for the equal equivalence ratio condition, while they are lower at the equal mass fuel injection condition for DME and DEE. The reduction in CO emission can be attributed to lower carbon fraction and oxygen content of DME and DEE. The lower carbon fraction results in lower incomplete combustion products and oxygen presence in the fuels makes an extra contribution to reduction of incomplete combustion products such as CO. It is estimated that the reduction in NO emission is sourced from lower combustion temperature given in Fig. 2(a). However, increases in CO and NO emissions for equal equivalence ratio condition can be attributed to rises in combustion temperatures of DME and DEE given in Fig. 2(b) and abundant oxygen sourced from increased mass fuel injection of DME and DEE as seen in Fig. 2(b). It is estimated that the higher combustion temperature leads to dissociation reactions and abundant oxygen in the combustion chamber contributes dissociation products such as CO and NO. 5. Conclusions In this study, the use of dimethyl ether and diethyl ether as alternative fuels in a diesel engine is investigated by means of a thermodynamic diesel engine cycle model. The following conclusions can be summarized from the results of the study. 1. Engine performance decreases and specific fuel consumption increases for dimethyl ether and diethyl ether in case of the same fuel injection rate (or unmodified fuel injection system) due to lower heating values of these alternative fuels so the decrements in brake power at 4200 rpm are about 32.1% and 19.4% and also the increments in brake specific fuel consumption at 2200 rpm are about 47.1% and 24.7% for dimethyl ether and diethyl ether, respectively. However, brake thermal efficiency of dimethyl ether and diethyl ether is better than or close to diesel due to their favorable combustion characteristic.

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2. Improvements in engine performance, brake thermal efficiency and brake specific fuel consumption are gained by both dimethyl ether and diethyl ether for the same equivalence ratio conditions while amount of fuel injection of the alternative fuels increases about 64% for dimethyl ether and 32% diethyl ether due to their lower stoichiometric airefuel ratio. The gains obtained are about 13.6% and 6% in brake power at 4200 rpm, 5.1% and 1.1% in brake thermal efficiency at 2200 rpm for dimethyl ether and diethyl ether respectively compared to diesel. However, brake specific fuel consumption is still higher about 43.5% for dimethyl ether and 23.6% for diethyl ether than diesel. 3. In brief, a large amount of dimethyl ether and diethyl ether is injected in order to maintain the same power output so some modification is required in fuel injection system. On the other hand, oxygen content of these alternative fuels improves the combustion process making it more complete. Thus, generally lower pollutant emissions are emitted with these fuels.

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