Optimization of methanol powered diesel engine: A CFD approach

Applied Thermal Engineering 106 (2016) 390–398

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Applied Thermal Engineering journal homepage: www.elsevier.com/locate/apthermeng

Research Paper

Optimization of methanol powered diesel engine: A CFD approach Dinesh Kumar Soni ⇑, Rajesh Gupta Department of Mechanical Engineering, M.A.N.I.T., Bhopal, India

h i g h l i g h t s
It is possible to reduce emissions further from methanol powered diesel engine.
Different emissions reduction methods can be compare through CFD approach.
Water emulsion method is better than EGR and initial swirl method in terms of emission reduction.
There is no need of any physical modification in diesel engine while using water emulsion method.
Simultaneous, reduction of NO and Soot emission is possible by using water emulsion method.

a r t i c l e

i n f o

Article history: Received 8 January 2016 Revised 1 June 2016 Accepted 4 June 2016 Available online 6 June 2016 Keywords: CFD code Methanol blended diesel engine Initial swirl method EGR method Water emulsion method Emissions

a b s t r a c t A two-stage strategy of emission reduction is applied to a single cylinder diesel locomotive to fulfill more stringent emission norms. First stage includes, the maximum possible emission reduction of diesel - methanol fuel through numerical simulation at low load operating condition. The higher emission reduction blend of diesel - methanol fuel is termed as ‘optimum blend’. Then, in the second stage, further emission reduction is attained by the application of three different methods of emission reduction at same operating condition. These methods are initial swirl ratio, EGR (Exhaust Gas Recirculation) and water addition. The commercially available CFD software AVL FIRE is used to perform simulation on a Kirloskar single cylinder diesel engine (model TV1). After selection of optimum blend from diesel-methanol fuel, effects of swirl ratio (1.0, 1.3, 1.6 and 2), variation in EGR percentage (10% and 20%) and effects of water addition (5%, 10% and 15%) in optimum blend is analyzed in terms of emission parameters. Furthermore, performance parameters (BSFC and BTE) are too analyzed for diesel fuel, diesel – methanol fuel and effective method from three emission reduction methods. The numerical simulation shows that, the water addition method is more efficient than the other two methods because it tends to reduce emissions effectively. Ó 2016 Elsevier Ltd. All rights reserved.

1. Introduction As the realization of the dreadful effects of the global warming phenomenon becomes intense, the governments across the globe are forced to enforce increasingly stringent emission norms to cut greenhouse - emission. To meet a greater degree of stringent emission reduction challenges, it is crucial for researchers to look for adequate technical solutions or alternative fuels which can reduce engine emissions. Diesel engines are identified as the major source of environmental pollution; emitting primarily carbonmonoxide, hydrocarbon, particulate matter (soot) and nitric oxide. While it is much easier to control the former two constituents, it is more difficult to reduce NOx and soot emission simultaneously.

⇑ Corresponding author. E-mail address: [email protected] (D.K. Soni). http://dx.doi.org/10.1016/j.applthermaleng.2016.06.026 1359-4311/Ó 2016 Elsevier Ltd. All rights reserved.

Several researchers have contemplated to reduce engine emission especially focussing on the reduction of soot as well as NOx emission. Owing to low carbon content (37.50%), high oxygen content (49.93%), and high latent heat of vaporization (1178 kJ/kg), methanol has been used as an alternative fuel in diesel engines to reduce NOx and soot emission [1–5]. There are two methods of methanol blending is widely discussed in the literature i.e. Fumigation mode and Emulsion mode. A separate fuel injector used to fumigate methanol with the inlet air in fumigation mode and in emulsion mode, an additive should be used to maintain stability of the mixture because of the poor miscibility of diesel - methanol blend [6,7]. Alternatively, a fuel injector can be utilized to introduce methanol in the combustion chamber [8,9]. Out of the above mentioned modes, emulsion mode is seems to be more effective than the fumigation mode in terms of emission parameters, whereas 10% diesel - methanol emulsion is acceptable [10].

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The experimental analysis of diesel - methanol emulsion is a well liked approach. Many researchers contributed through experimental analysis of diesel - methanol emulsion fuel. Diesel - methanol emulsion from 0% to 15% by volume is assessed at different RPM and result shows the higher NO emission, whereas CO and HC emission are lower than pure diesel [11]. On the other hand, methanol is fumigated in the combustion chamber and mixture is fired up by diesel fuel. The results indicated that, NO and soot emissions are reduced, while HC and CO emission are higher at varying load and speed condition [12]. The numerical modeling approach of diesel - biodiesel blend is also possible [13]. Besides of biodiesel, alcohols can be simulated numerically by using the appropriate CFD tool [14]. An et al. [15] performed numerical simulation of methanol fuel with different proportions and at varying load conditions. The result shows the lower CO and soot emission; whereas NOx remained same at full load. Li et al. [16] studied the reactivity controlled compression ignition engine (RCCI) by numerical simulation approach and revealed that the HC and soot emission decreases with moderate methanol quantity and advanced injection timing. Methanol is also used as additive in diesel - biodiesel blends. However, emission results remain unchanged after the application of methanol as an additive [17]. Whereas; methanol blended with biodiesel could reduce soot and NOx emission at the same time [18]. Additionally, ethanol blended biodiesel fuel can also be employed to reduce emissions, but it is less efficient, as it contributes to higher NOx and soot emission compared to methanol blended biofuel [19]. Nowadays, it is noticeable fact that the methanol is applied to achieve higher reduction of emissions, especially NOx and soot emission. Close to other way of simultaneous NOx and soot emission reductions are by applying high pressure fuel injection [20] and advanced injection strategies [21]. In spite of emission reduction techniques, these are pricy and hard to manage. To overcome those issues, an optimization of the methanol blended diesel engine is performed by using much simpler and worthwhile methods such as initial swirl, exhaust gas recirculation and by water emulsion. Significant research is proceeding on to reduce emissions from diesel engines by going through initial swirl method. Air fuel mixing and burning rate is greatly influenced by turbulence, squish and swirl. Improvement in inlet design port can generate higher swirl inside the combustion chamber [22]. Reentrant piston geometries can produce high swirl and tumble, which leads to low NOx and soot emissions [23]. Furthermore; the combination of initial swirl and piston generated swirl can also be useful to bring down emissions from combustion chamber [24]. One of the well established emission reduction methods is the EGR (Exhaust gas recirculation). Mechanism of emission reduction through EGR method comprises a low flame temperature and oxygen concentration. Even though, it is an effective emission reduction method, but not suited for long term use (increases engine wear and soot emissions) [25,26]. The EGR method can be utilized to a dual fuel diesel engine to reduce further emissions from diesel - biodiesel blends [27–29]. Importantly, increasing amount of low pressure EGR tends to reduce NO emission effectively than high pressure EGR, due to low temperature [30] and moreover, with the combination of other methods, the combined effect of EGR and advanced injection strategies are useful for emission reduction from diesel engines [31]. In general, EGR method reduces combustion temperature, which leads to low NO emissions. The water emulsion method is effectively reduces NO emissions due to micro combustion or heat sinking phenomena of water particles. The water particles absorb heat of vaporization in the form of latent and sensible heat, which results to low incylinder temperature. Thus, Low incylinder temperature produces less NO emis-

sions [32]. In some studies, NO and soot emissions can be brought down simultaneously by water emulsion method [33–35]. The basic cause of emission reduction is the better air fuel mixing due to micro combustion of water particles followed by spray optimization [36–38]. Although, water emulsion has been used widely with pure diesel for emission reduction, but it will be interesting to analyze the micro explosion property of water in diesel - methanol blended diesel engine. Overall, higher latent heat of vaporization of methanol reduces combustion temperature; causes in reduction of NOx formation. Moreover, the soot formation can be reduced by methanol addition due to the absence of carbon-carbon bond in the methanol. Methanol has an extra oxygen atom, which also helpful for complete combustion and reduces CO and HC formation. As a result, to find a solution for energy shortage and environmental issues in the future, methanol has required strength and adaptability. The combustible properties of methanol over other possible alcohols are listed in Table 1. A naturally - aspirated single cylinder diesel engine is numerically simulated at operating condition of constant speed, at constant load (1 kW) and fixed mass of the fuel. Conferring to literature survey, it is essential to adopt a suitable emission reduction method to obtain minimum possible emissions from diesel engine. Present paper explores the impact of different emission reduction methods applied to a methanol blended diesel engine by means of numerical simulation and helps to optimize the diesel engine by selecting a suitable method from all of them. The optimization of diesel -methanol blended diesel engine is performed to evaluate the effect of emission reduction methods. Whereas, three emission reduction methods involved in the optimization process are Initial swirl, EGR (Exhaust gas recirculation) and water addition. It is beneficial to simulate these methods through CFD tool to economize time and cost, before moving towards the experimental approach. 2. Methodology To find an optimum blend from diesel - methanol blend, methanol is blended with diesel at three different proportions such as D + M10 (Diesel and 10% methanol), D + M20 (Diesel and 20% methanol), and D + M30 (Diesel and 30% methanol), and numerical simulation is carried out to find an optimum blend. There is a proper method to introduce methanol quantity in the simulation through AVL FIRE ESE diesel module. The selection of base fuel (optimum fuel) from three blends of diesel - methanol is executed on the basis exhaust emission. Then, the base fuel is further processed for next stage of investigation by means of application of different emission reduction methods. Firstly, initial swirl method is introducing with the variation in the swirl ratio as 1, 1.3, 1.6 and 2 and then, EGR method is brought in simulation approach with variation in percentage of exhaust gas recirculation as 10% and 20%. At the last, base fuel is emulsified by water in the proportion of 5%, 10% and 15% by volume. Table 1 Properties of alcohols [56]. Properties

Methanol

Ethanol

Butanol

Chemical formula Molecular weight Oxygen content, wt% Hydrogen content, wt% Stoichiometric AFR Lower heating value, MJ/kg Heat of evaporation, kJ/kg Research octane number Motor octane number Carbon content, wt% Vapor pressure (psi at 37.7)

CH3OH 32.04 49.93 12.5 6.43 20 1178 112 91 37.5 4.6

C2H%OH 46.07 34.73 13.1 8.94 27 840 111 92 52.2 2

C4H9OH 74.12 21.59 13.5 11.12 33 578.4 113 64.8 0.33

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2.1. Numerical simulation

Tþ ¼

The numerical simulation is performed on hemispherical bowl shaped piston geometry of a single cylinder direct injection diesel engine. The major specifications of the diesel engine are listed in Table 2. The PISO (Pressure implicit with splitting of operator) algorithm [39] is used to solve Navier-Stokes equations by applying pressure- velocity calculation procedure and it is found to be an improved algorithm than SIMPLE (Semi- implicit method for pressure linked equations) algorithm [40,41]. To start the simulation process, the initial conditions are selected according to experimental data. The initial pressure and temperature are set equal to 0.65 MPa and 300 K respectively. The initial density is calculated by using ideal gas equation. The turbulence kinetic energy and length scale are set to 15 m2/s2 and 0.003 m respectively. For boundary conditions, the momentum boundary conditions of cylinder head and liner are specified as fixed wall, and boundary condition of piston is specified as moving wall. Whereas, thermal boundary conditions for piston, liner and cylinder head are taken as 550 K, 425 K and 475 K. The present optimization strategy used a k-zeta -f turbulence model to solve energy transport equations leading to enhance the numerical stability of the simulation approach stated by Hanjalic et al. [42]. This model is more accurate than the standard two - equation eddy viscosity k-e model. The k-zeta -f turbulence model is widely accepted for computational meshes and flow conditions of any dimensionless distance near the wall (Y+). In the presented model, the hybrid wall treatment is integrated near the wall by using standard wall function. The following wall function equations are used to represent velocity and temperature profiles near the wall [43]. In the laminar boundary layer (Y+ 6 11.63)

Uþ ¼ Y þ

UT ¼





 Tw Cp T ln qw T

qU T

ð7Þ

pffiffiffiffiffiffiffiffiffiffiffi sw =q

ð8Þ

where UT is the friction velocity and sw is local wall shear stress. In the above equations k (Von Karman constant), E (Empirical constant for wall function) and re,t (turbulent Prandtl number) are coefficients and their values are equal to 0.41, 9.0 and 0.90 respectively. Furthermore, Tw, T, q, re, t, qw, Cp, Y and U  have their usual meaning as wall temperature, temperature, density, laminar Prandtl number, kinematic viscosity, wall heat flux, constant pressure specific heat, distance from the wall and velocity component parallel to the wall respectively. The compression ratio, volume of the hemispherical bowl, mass of fuel injected and speed of the engine are kept constant for whole set of simulations. Only one - third section of the piston bowl geometry is meshed in AVL FIRE ESE diesel module as shown in Fig. 1. A grid independency test is performed to choose the appropriate mesh (independent of grid size) from three mesh sizes of 29,516, 40,868 and 50,208 cells. Fig. 2 shows the recorded results of the grid independency test and it can be seen that, pressure curves of different mesh sizes are within the acceptable range and holds roughly 6 h of simulation time. Thus, mesh size with 40,868 cells consider for further simulation. 2.2. Computational models and validation In the present research, to figure out the complex combustion process of the diesel engine, commercially available CFD software

ð1Þ

Tþ ¼ r e Uþ

ð2Þ +

In the turbulent boundary layer (Y > 11.63)

1 lnðEþy Þ k 
 re Tþ ¼ re;t Uþ þ P

Uþ ¼

re;t

ð3Þ ð4Þ

In the above equation, P is viscous sub layer thermal resistance factor, whereas Y+, U+ and T+ are dimensionless parameters known as distance from the wall, non dimensional tangential velocity and temperature respectively. These symbols are described as follows:

Uþ ¼

YU T

t

U UT

29516

ð6Þ

Table 2 Engine specifications. Make Number of cylinders Bore Stroke Swept volume Type of piston bowl Compression ratio Rated output Connecting rod length No. of nozzle holes Rated speed

Fig. 1. Geometrical grid at TDC.

ð5Þ

Kirloskar engine, TV-1 1 87.5 mm 110 mm 661 cc Hemispherical shape 17.5 5.2 kW 234 mm 3 1500 rpm

Pressure (bar)

Yþ ¼

40868

50208

50 45 40 35 30 25 20 15 10 5 0 620

670

720 Crank angle (Deg)

Fig. 2. Grid independency test.

770

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The validation of experimental and simulation results are figure out in Fig. 3, alongside the pressure and heat release rate at different crank angles for the same engine speed under same operating condition of 1 kW load. Additionally, NO emission is also validated at different diesel - methanol blends. Even though, the qualitative trend between predicted and experimental results is found to be descent, still some inconsistencies are observed at TDC. This may be due no availability of input parameter such as injection duration. All the used model of present optimization procedure is well validated by Tatschl [51].

AVL FIRE is used, which worked on a finite volume approach. The premixed and diffusion combustion phases of diesel methanol blends can simulate with the use of a coherent flame model (CFM); however, an advanced form of CFM model, Extended coherent flamlet model - three zone (ECFM-3Z) is used to combine spray model to EGR and NO formation [44,45]. The wave breakup model is applied for diesel fuel spray simulation [46]. A Spray wall interaction model ‘walljet1’ is used for the simulation of non - evaporated fuel particles those striking on the wall of the combustion chamber [47]. The extended zeldovich mechanism is used to calculate the NO emissions and this model can be simulated with the ECFM-3Z combustion model based on equilibrium approach [48]. The kinetic soot model is used to calculate soot formation and oxidation [49]. In this work, both Dukowicz model [50] and multicomponent evaporation model allows the usage of dissimilar types of fuel in combustion process.

Before using a specific method to reduce emissions from diesel methanol blended diesel engine by experimental analysis, its aptness in terms of emissions should be adjudicated through simula-

Experiment-Pressure Simulaon-HRR

60

Simulaon Pressure Exeperiment HRR 70

60

50

60

50

50

40

30

20

20

10

40 30

20

20 0

0 710 760 Crank Angle (Deg)

10 660

710

760 Crank Angle (Deg)

(b)

(a) Simulaon-Pressure Exeperiment-HRR 100

60

80

80

40

60

30 40

20 10

20

0

0 660

710

Simulaon-Pressure Exeperiment-HRR 90

50

HRR (J/Deg)

Pressure (bar)

50

Experiment-Pressure Simulaon-HRR

Pressure (bar)

Experiment-Pressure Simulaon-HRR

0

40

70 60

30

50 40

20

30

10

20

0

0

HRR (J/Deg)

660

50

30

10

10

0

70

HRR (J/Deg)

40

30

Simulaon-Pressure Exeperiment-HRR 80 60

Pressure (bar)

40

HRR(J/Deg)

Pressure (bar)

Exeperiment Pressure Simulaon HRR

3. Result and discussion

10

760

660

710 760 Crank Angle (Deg)

Crank Angle (Deg)

(c)

(d) Experiment-NO emission

Simulaon-NO emission

NO Mass Fracon

0.0004 0.00035 0.0003 0.00025 0.0002 0.00015 0.0001 0.00005 0

Diesel

D+M10

D+M20

D+M30

Diesel and Diesel-Methanol blends (e) Fig. 3. Comparison of experimental and simulated results at low load with (a) Pure Diesel, (b) D + M10, (c) D + M20, (d) D + M30, and (e) NO emission at all blends.

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tion approach. Therefore, the emission characteristics of diesel methanol fuel and impact of different emission reduction methods on diesel methanol fuel is analyzed and discussed in this segment.

Methanol is proved to be advantageous over diesel fuel in terms of emissions. It absorbs latent heat of vaporization to evaporate and atomized with the air. Thus reduces combustion temperature. Moreover, high concentration of oxygen atoms in the fuel serves to complete combustion in the combustion chamber. Additionally, the low total equivalence ratio of diesel methanol blend increases heat capacity of the methanol fuel, which in turn reduces combustion temperature and emits less NO emissions [52]. Diesel - methanol blend forms lean mixture as methanol quantity increasing in the blend. A Lean mixture of blend tend to low flame temperature, caused low NO formation. It can be seen from Fig. 4(a) that, NO mass fraction is monotonically reducing with higher percentage of methanol in the blend. Therefore, follows expected trend. Whereas, maximum reduction of 27% is achieved at D + M30 blend. An increasing trend of soot emission is observed at higher diesel methanol blend because of low combustion temperature. The lower heating value of methanol is about half of diesel and latent heat of vaporization of methanol is about 4 times of diesel fuel, which produces cooling effect in the engine and leads to a reduction in peak temperature [53]. On the other hand, methanol contains 34% oxygen and lower cetane number than diesel fuel. The combined effect of these two increases peak temperatures in the cylinder. At low engine load, the in cylinder temperature is low and it further decreased as quantity of methanol increased. Low in cylinder temperature increases soot formation due to incomplete combustion. The working fuel of the present analysis is below than stoichiometric ratio and low load condition in which cooling effect is more influential over other effects in case of soot formation. Additionally, high load condition of the engine produces the less effective cooling effect [54]. The combustion speed at low load is lower than high load. Therefore, at slow combustion speed, oxygen part of the methanol is consumed to reduce CO and HC emission in the premixed combustion phase. Whereas, soot formation mainly results from diffusion combustion (Mixing controlled combustion phase). So, absence of sufficient oxygen at the time of diffusion combustion will lead to high soot formation at low combustion speed. Finally, the effect of high oxygen of methanol is diminished at low load combustion. The CO and HC emissions are simulated by using the ECFM-3Z combustion model. Whereas, CO is calculated through following equations [43].

NO


Cn Hm Ok þ

3.1. Emission characteristics of diesel-methanol fuel

0.00025


 m k m O2 ! n CO2 þ H2 O Cn Hm Ok þ n þ  4 2 2

ð9Þ

 n k m  O2 ! n CO þ H2 2 2 2

ð10Þ

where n, m and k denote carbon, hydrogen and oxygen atoms. AVL FIRE Engine does not directly calculate HC. AVL Fire after treatment module treated above reaction further to produce HC as standard output in result. Fig. 4(b) indicates a higher reduction of 58% in CO emission which is achieved by D + M30. This reduction may be attributed to the high oxygen content in methanol fuel, brings complete combustion inside cylinder. Whereas, formation of unburned hydrocarbons (HC) is greatly influenced by improper atomization of fuel. Methanol has a low quantity of hydrogen and carbon atoms than diesel fuel [11], additionally, 34% of oxygen concentration of methanol fuel leads to complete combustion of the mixture. This likely to reduce HC emissions and it can be noted from Fig. 4(b) that, higher reduction of HC emissions is achieved at D + M30 than conventional diesel fuel. Hence D + M30 is selected as an optimum blend. 3.2. Impact of initial swirl on optimum blend Swirl is defined as the motion provided to the air inside combustion chamber. Local equivalence ratio inside the combustion chamber is greatly affected by swirl, as swirl distributed the mixture inside combustion chamber. High swirl ratio reduces the concentration of fuel on the cylinder wall and avoids accumulation of too much fuel on the wall [22]. Fuel rich zones are decreasing as the swirl ratio increasing from SR0 to SR2. Furthermore, there may be local fuel rich and fuel lean zones inside the combustion chamber at constant mass of fuel injected. However, better air-fuel mixing is accomplished at low swirl ratios [22], due to two turbulent kinetic energy zones observed during simulation process. Accordingly, SR1.3 is likely to be more uniform inside the piston chamber. On the other hand, high swirl ratio accelerates the rotation of air inside the combustion chamber. The higher rotation of air brings spray bending angle downward, caused to separation of spray lines. Consecutively, results in incomplete combustion [22]. A scattered pattern of NO mass fraction is observed at different swirl ratios as shown in Fig. 5(a). The trend of NO mass fraction seems to be increasing for SR1.3 to SR2. The soot mass fraction increases with almost constant rate of increment at higher swirl ratios after speedily drift up to SR1.3. It can be seen from Fig. 5

CO

SOOT

HC

0.04 0.035

Mass fracon

Mass fracon

0.0002 0.00015 0.0001 0.00005

0.03 0.025 0.02 0.015 0.01 0.005

0

0

D

D+M10

D+M20

D+M30

D

D+M10

(a) Fig. 4. Emission characteristics of diesel and methanol blends.

D+M20

(b)

D+M30

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NO

0.00018

SOOT

CO

0.025

HC

0.00016 0.02

0.00012

Mass fracon

Mass fracon

0.00014 0.0001 0.00008 0.00006 0.00004

0.015 0.01 0.005

0.00002 0

0

1

1.3 Swirl rao

1.6

0

2

0

1

1.3 Swirl rao

(a)

1.6

2

(b)

Fig. 5. Emission characteristics of swirl ratio applied on optimum blend.

(a), that the NO and soot mass fraction is displaying comparably better results at SR1.3 than other swirl ratio. Methanol is oxygenated fuel and provides sufficient amount of oxygen, which heads to complete combustion and reduces CO and HC emissions in premixed combustion zones. Thus, results recorded at 80°ATDC are enough to show the actual trend of CO and HC emissions. Fig. 5 (b) shows that, the peak of CO and HC emissions can be observed at SR2 due to higher local fuel rich zone in the combustion chamber which results to incomplete combustion after burning. 3.3. Impact of Exhaust gas recirculation (EGR) on optimum blend The EGR method is expressively to reduce emissions from diesel engines [23–26]. The execution of EGR methods is primarily approached to lean air fuel ratio in premixed combustion mode i.e. before compression stroke. The lean combustion phenomena shift the whole combustion process towards the expansion stroke side. This shifting is due to lower ignition delay. The shifting of combustion process results to low combustion pressure inside the cylinder. As a result, peak pressure decreases and volume increases as the piston moves from TDC to BDC in the expansion stroke side [30]. Furthermore, the EGR method at sufficient in cylinder temperature causes formation of CO2 and H2O instead of CO and HC respectively. Fig. 6(a) displays low NO formation at high EGR rates. Low NO formation is attributed to low in cylinder pressure and temperature inside the combustion chamber, which is due to shifting of

NO

the combustion process [27]. A dipping trend of soot emission is observed in Fig. 6(a). Soot formation decreases because of low combustion temperature, as EGR rate is changed from 0% to 10%. Soot emission follows same trend such reported in literature [54]. Although, soot emission increases as EGR rate is changed from 10% to 20% owing to lack of oxygen atoms. EGR leads to incomplete combustion in low surrounding temperature due to dilution of intake air and another reason could be the poor fuel utilization inside combustion chamber [16,27]. Fig. 6(b) shows that, the high rate of EGR tends to high CO formation due to lack of oxygen atoms at lean combustion phenomena [27]. It can also be seen in Fig. 6(b) that, HC emissions are increasing continuously from 0% EGR rate to 20% EGR rate. This is because of the incomplete and long combustion process [55].

3.4. Impact of water addition on optimum blend The following reaction is active while water is added to the optimum blend;

CH3 OH þ H2 O <¼> CO2 þ 3H2

The conversion of methanol to carbon di oxide and hydrogen is occurring at high flame temperature inside combustion chamber, as shown in the above equation. The formation of Carbon di oxide as a product helps to cut CO emissions from diesel engine, whereas the extra amount of hydrogen in the product is advantageous for

CO

SOOT

0.00016

0.02

0.00012

Mass fracon

Mass fracon

HC

0.025

0.00014 0.0001 0.00008 0.00006 0.00004

0.015 0.01 0.005

0.00002 0

ð11Þ

0 0%

10% EGR rate

20%

0%

(a) Fig. 6. Emission characteristics of EGR rate applied on optimum blend.

10% EGR rate

(b)

20%

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SOOT

CO 0.02

0.00014

0.018

0.00012

Mass fracon

Mass fracon

NO 0.00016

0.0001 0.00008 0.00006

0.016 0.014 0.012 0.01 0.008 0.006

0.00004

0.004

0.00002 0

HC

0.002 0%

5%

10%

0

15%

0%

5%

10%

15%

Water emulsion

Water emulsion

(a)

(b)

Fig. 7. Emission characteristics of water emulsion applied on optimum blend.

Table 3 Response of different methods of emission reduction. Emission (mass fraction)

D

D + M30a (optimum blend)

SR 1.3b

EGR 20%b

Water additionb (15% volume)

NO Soot CO HC

0.000204 0.000025 0.0353 0.0279

0.000149 (27% ;) 0.000063 (175% ") 0.0148 (58% ;) 0.018 (65% ;)

0.000135 (10% ;) 0.000095 (46% ") 0.0133 (11% ;) 0.0221 (22% ")

0.000093 (36% ;) 0.000061 (almost same) 0.0162 (9% ") 0.0207 (15% ")

0.000039 (74% ;) 0.000057 (10% ;) 0.0126 (15% ;) 0.0151 (16% ;)

" Emission increases. ; Emission decreases. a Comparison with diesel fuel. b Comparison with optimum blend (D + M30).

Then, swirl ratio at SR1.3, EGR rate with 20% and water addition with 15% in diesel methanol blend has chosen as effective cases from respective methods. The responses of these cases are listed in Table 3.

the combustion process. It is clear from the referenced articles that, diesel emulsified with water is conducive for the diesel engines as far as emissions are concerned. A micro combustion phenomenon improves air fuel mixing inside the cylinder due to high heat of vaporization of water particles. Water particles absorb latent heat to vaporize, thus combustion temperature reduces. It can be observed from Fig. 7(a), the lower combustion temperature produces low NO emissions as water quantity increases such as 5% (D + M30 + W5), 10% (D + M30 + W10) and 15% (D + M30 + W15). On the other side, water particles are converted in the shape of tiny particles, which passes to improve air fuel mixing. Simultaneously, soot emissions get also reduced due to proper air fuel mixing [32–38] and enhanced air fuel mixing reduces CO and HC mass fraction, see Fig. 7(b). Thus, addition of 15% (D + M30 + W15) water quantity by volume in diesel methanol blend is selected as an appropriate case from this section.

3.5. Performance parameters On the basis of Table 3, 15% (D + M30 + W15) water emulsified diesel methanol blend is compared with diesel and optimum blend of diesel and methanol (D + M30) in terms of Brake specific fuel consumption (BSFC), Brake thermal efficiency (BTE) and Engine power. The Brake specific fuel consumption (BSFC) is increasing for diesel, D + M30 and D + M30 + W15 fuel continuously, as displays in Fig. 8(a). Any blending in the diesel fuel reduces the lower heating value of the mixture and thus higher amount of the fuel has to be injected to maintain same amount of power. Hence, an increasing trend is observed for BSFC. Due to

BSFC

BTE

500

BTE (%)

BSFC (g/kwh)

600

400 300 200 100 0 Diesel

D+M30

D+M30+W15

Engine power

11.2 11 10.8 10.6 10.4 10.2 10 9.8 9.6 9.4 9.2 9

1.4 1.2 1 0.8 0.6 0.4 0.2 0 Diesel

(a)

D+M30

(b) Fig. 8. Comparison of performance parameters.

D+M30+W15

Power (kw)

700

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lower heating value of the blends, the less amount of heat is liberated during mixture combustion. It causes low brake thermal efficiency (BTE) of the engine, as can be seen from Fig. 8(b). Whereas, transformation of thermo physical properties (lower calorific value and density) is responsible for lower brake power at optimum fuel blends.

Table 4 Emission characteristics of optimum blend and water emulsion.

4. Conclusion

" Emission increases. ; Emission decreases. a Comparison with diesel fuel. b Comparison with optimum blend (D + M30).

The optimization is performed to achieve minimum possible emissions from diesel- methanol optimum blend by the application of initial swirl, EGR percentage and water emulsion methods. The present strategy is one of its kind and can be possible through CFD approach. It is safer to use CFD approach before moving to experimental analysis because of almost no manufacturing risk involved in it. Additionally, results are encouraging to attain minimum emission reduction. The major outcomes from optimization are as follows: 1. The high amount of methanol in the diesel fuel leads to low NO emissions from diesel engines. Increase in methanol up to 30% has achieved a maximum reduction of 27%, 58% and 65% in NO, CO and HC emission respectively compared to conventional diesel fuel. However, soot emissions are exception in result. Therefore, D + M30 blend is selected as an optimum blend and treated by emission reduction methods in next stage of simulation. 2. Influence of initial swirl on emissions has been investigated numerically and the results indicate lower emission in case of SR1.3 whereas soot emission is almost constant for SR1.3 and SR1.6, however SR2 is showing high soot formation. Simultaneously, CO and HC emission is almost same for all swirl ratios. Hence, it can be concluded that SR1.3 is optimum. While results compared to optimum blend, NO and CO reduced by 10% and 11%, respectively, whereas HC and soot increased by 22% and 46% respectively (see Table 3). 3. Two rates of EGR i.e. 10% and 20% have been investigated numerically and result indicates that, NO and soot formation is less for 20% rate of EGR, whereas CO and HC is high on same rate of EGR. CO and HC emission is increased up to 9% and 15%, respectively, whereas NO reduced to 36% when compared to optimum blend as shown in Table 3. However, soot is approximately same as optimum blend. By considering severity of NO and soot emissions, EGR 20% rate is supposed to be an appropriate case. 4. Numerical simulation of water blends shows that NO and Soot formation is lowest while using amount of water is 15% with diesel on volume basis. Similarly, it can be seen from Fig. 7 (b), that CO and HC emission are also low in case of 15% water addition. So, W15 is chosen as optimum case water-diesel blends. It can be seen from Table 3 that, all parameters of emission are reducing comparison to optimum blend (D + M30), as NO, Soot, CO and HC reduced to 74%, 10%, 15% and 16% respectively. 5. In the end, the leading case from diesel methanol blend and water emulsified diesel methanol blend is itemized in Table 4. The optimum cases from all methods show their comparative numerical values than optimum bled (D + M30). The only method in which all emission parameters are reducing is water emulsion method and it helps to concluded that, the water emulsion method has inclination towards minimum emission from diesel methanol blends but there is some sort of compromise has to pay with engine power and brake thermal efficiency.

Emission (mass fraction)

Diesel fuel

D + M30a (optimum/ base blend)

D + M30 + W15b

NO Soot CO HC

0.000204 0.000025 0.0353 0.0279

0.000149 (27% ;) 0.000063 (175% ") 0.0148 (58% ;) 0.018 (65% ;)

0.000039 (74% ;) 0.000057 (10% ;) 0.0126 (15% ;) 0.0151 (16% ;)

5. Future scope of the work Application of different emission reduction techniques on a diesel engine cannot be judged through experimental analysis due to involvement of time and cost factor. Investigation of these methods at a time can be carried out through numerical simulation. Out of the three emission reduction techniques, two techniques (EGR and Initial swirl) employed modification of the internal or external geometry of the diesel engine, which proved to be costly up to some extent. On the other hand, water addition doesn’t need any changes in the diesel engine and it is cost efficient method because of the use of simple tap water. It should be comfortable to use water emulsion method instead of initial swirl (swirl ratio) and EGR method in future for experimental analysis. Additionally, it reduces emission up to lowest possible level, which seems to be more beneficial in terms of environmental issues. In short, this paper supports the water addition method over EGR and initial swirl method to reduce emissions. References [1] L. Zhu, C.S. Cheung, W. Zhang, Z. Huang, Emission’s characteristics of a diesel engine operating on biodiesel and biodiesel blended with ethanol and methanol, Sci. Total Environ. 408 (2010) 914–921. [2] Y. Liu, W. Jiao, G. Qi, Preparation and properties of methanol–diesel oil emulsified fuel under high-gravity environment, Renew. Energy 36 (2011) 1463–1468. [3] C.H. Cheng, C.S. Cheung, T.L. Chan, S.C. Lee, C.D. Yao, K.S. Tsang, Comparison of emissions of a direct injection diesel engine operating on biodiesel with emulsified and fumigated methanol, Fuel 87 (2008) 1870–1879. [4] Z.H. Zhang, C.S. Cheung, T.L. Chan, C.D. Yao, Emission reduction from diesel engine using fumigation methanol and diesel oxidation catalyst, Sci. Total Environ. 407 (2009) 4497–4505. [5] C.S. Cheung, L. Zhu, Z. Huang, Regulated and unregulated emissions from a diesel engine fuelled with biodiesel and biodiesel blended with methanol, Atoms. Environ. 43 (2009) 4865–4872. [6] Z.H. Huang, H.B. Lu, D.M. Jiang, K. Zeng, B. Liu, J.Q. Zhang, Engine performance and emissions of a compression-ignition engine operating on the diesel/ methanol blends, Proc. InstMech. Eng. Part D: J. Automob. Eng. 218 (2004) 435–447. [7] M.R. Chao, T.C. Lin, H.R. Chao, F.H. Chang, C.B. Chen, Effects of methanolcontaining additive on emission characteristics from a heavy-duty diesel engine, Sci. Total Environ. 279 (2001) 167–179. [8] M.G. Popa, N. Negurescu, C. Pana, A. Racovitza, Results obtained by methanol fuelling diesel engine, SAE Tech Paper No. 2001-01-3748, 2001. [9] C.D. Yao, C.S. Cheung, C.H. Cheng, Y.S. Wang, Reduction of smoke and NOx from diesel engines using a diesel/methanol compound combustion system, Energy Fuel 21 (2007) 686–691. [10] C.H. Cheng, C.S. Cheung, T.L. Chan, S.C. Lee, C.D. Yao, K.S. Tsang, Comparison of emissions of a direct injection diesel engine operating on biodiesel with emulsified and fumigated methanol, Fuel 87 (2008) 1870–1879, http://dx.doi. org/10.1016/j.fuel.2008.01.002. [11] C. Murat, K. Huseyin, C. Eyub, S. Ozgur, An experimental investigation on effects of methanol blended diesel fuels to engine performance and emissions of a diesel engine, Sci. Res. Essays 6 (15) (2011) 3189–3199, http://dx.doi.org/ 10.5897/SRE 11.230. [12] C. Yao, C.S. Cheung, C. Cheng, Y. Wang, T.L. Chan, S.C. Lee, Effect of diesel/ methanol compound combustion on diesel engine combustion and emissions, Energy Convers. Manage. 49 (2008) 1696–1704, http://dx.doi.org/10.1016/j. enconman.2007.11.007. [13] M.C. Cameretti, U. Ciaravola, R. Tuccillo, L. De Simio, S. Iannaccone, A numerical and experimental study of dual fuel diesel engine for different

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