Investigation of BiodieselâDiesel Fuel Blends on Combustion

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Investigation of biodiesel-diesel fuel blends on combustion characteristics in a light-duty diesel engine using OpenFOAM Harun M. Ismail, Hoon Kiat Ng, Suyin Gan, Xinwei Cheng, and Tommaso Lucchini Energy Fuels, Just Accepted Manuscript • DOI: 10.1021/ef301162d • Publication Date (Web): 12 Nov 2012 Downloaded from http://pubs.acs.org on November 21, 2012

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Investigation of Biodiesel-Diesel Fuel Blends on Combustion Characteristics in a LightDuty Diesel Engine Using OpenFOAM

Harun Mohamed Ismail1, Hoon Kiat Ng1*, Suyin Gan2, Xinwei Cheng1 , Tommaso Lucchini3 1

Department of Mechanical, Materials and Manufacturing Engineering, The University of

Nottingham Malaysia Campus, Jalan Broga, 43500 Semenyih, Selangor Darul Ehsan, Malaysia 2

Department of Chemical and Environmental Engineering, The University of Nottingham

Malaysia Campus, Jalan Broga, 43500 Semenyih, Selangor Darul Ehsan, Malaysia 3

Department of Energy, Politecnico Di Milano, Via Lambruschini 4, 20156 Milano, Italy

*

Corresponding author: Tel.: +603 89248161; Fax: +603 89248017; Email-address:

[email protected]

Abstract This paper reports the differences in combustion characteristics of fossil diesel and the methyl esters of coconut (CME), palm (PME) and soy (SME) over a range of engine conditions. The studies are conducted at a constant engine speed of 2000 rev/min, and at engine load values of 0.5 kW (low), 1.5 kW (mid) and 2.5 kW (high). The investigated fuels are CME, PME and SME at 0% diesel mixture (B100) and 50% diesel mixture (B50) and fossil diesel (B0). Here, the OpenFOAM® open source Computational Fluid Dynamics (CFD) codes are utilised to simulate the in-cylinder events. An in-house model for thermo-physical and transport properties is employed, along with a mechanism comprises 113 species and 399 reactions with integrated NOx

1 ACS Paragon Plus Environment

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kinetics. Good levels of accuracy are achieved in the prediction of the ignition delay (ID) period, peak pressures, pressure traces and heat-release rate (HRR) profiles for all the test fuels. Biodiesel fuels are found to produce larger fuel droplets sizes, longer spray penetration and lower vaporisation rate as compared to that of fossil diesel. In terms of combustion behaviour, the ID decreases as the engine load is increased for neat CME, PME and SME. However, the ID period increases as the load is raised for the B50 mixtures of CME, PME and SME. All neat biodiesels and their blends except neat SME produce shorter ID period than fossil diesel. The changes in the fuel types from diesel to biodiesel alter the physical delay and chemical delay, and hence the overall ID period. As a result, the variations in the combustion phasing between the fuels are recorded.

Keywords: Biodiesel-diesel combustion characteristics, OpenFOAM, multi-component fuel modelling, diesel engine

1.

Introduction In recent years, biodiesel is increasingly used as a substitute to fossil diesel caused mainly

by depleting petroleum reserves and more stringent emission standards1. In addition, due to its physico-chemical characteristics which are similar to that of fossil diesel2, biodiesel can be used with minimal or no modifications on existing compression ignition (CI) or diesel engines. Neat biodiesel and its blends are known to reduce tailpipe concentrations of particulate matter (PM), hydrocarbons (HC) and carbon monoxide (CO)3. Relative increases in nitrogen oxides (NOx) emissions as compared to diesel fuel are observed in some studies when biodiesel fuels are tested in unmodified engines3,4. Nevertheless, there are also studies which demonstrated that biodiesel


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fuels such as coconut methyl ester (CME) reduce the formation of NOx5, 6. In light of these inconclusive findings, more comprehensive studies are required to improve the fundamental understanding of biodiesel combustion and emissions characteristics as well as to conclusively establish the benefits and drawbacks of biodiesel usage in diesel engines7,8. In diesel engines, the type of fuel (diesel/biodiesel/blends) substantially affects both incylinder physical and chemical processes due to the varying physico-chemical properties. These in-cylinder processes dictate the combustion characteristics which in turn influence the engineout pollutants. For instance, at low load conditions, the main part of the combustion is premixed. Here, high energy is being released over a short period of time in the premixed combustion (PMC) phase near the top-dead centre (TDC) which significantly increases the local and average in-cylinder temperatures. As a result, NOx formation rate in this combustion phase is increased leading to higher engine-out NOx9. On the other hand, during the mixing-controlled combustion (MCC) phase, soot formation rate is increased due to fuel rich combustion in the core of the spray jet whereby the equivalence ratio is between 1.5 to 4 and the local in-cylinder temperatures is between 1400 K to 2000 K9. Thus, it is necessary to fully understand the combustion characteristics of various biodiesel fuels in order to better comprehend and improve the emissions of these fuels. To date, the fundamental understanding of spray structure and combustion characteristics of various biodiesel fuels achieved through combustion modelling is yet to be fully established. Early initiatives in this field were carried out by Yuan et al.10, 11 who presented methods for predicting physical properties for biodiesel that can be used for combustion modelling. Chakravarthy et al.7 estimated the thermophysical and transport properties of soy biodiesel for use in computational fluid dynamics (CFD) modelling to evaluate the extent to which


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combustion is modified when biodiesel is substituted for diesel. The preliminary results obtained showed that biodiesel ignition is significantly delayed due to slower liquid evaporation, with the lower vapour pressure and higher liquid heat capacity of biodiesel being two key contributors to this slower evaporation rate. Another CFD modelling effort by Ra et al.8 investigated the effects of soy physical properties on diesel engine combustion characteristics. At various engine operating conditions, the fuel physical properties had significant effects on ignition delay (ID) and burning rates, with no single physical property dominating the differences in ID between diesel and biodiesel fuels. Nonetheless, the simulation results were most sensitive to the liquid fuel density, vapour pressure and surface tension through their effects on the mixture preparation processes. Recent works by Brakora et al.12 as well as Golovitchev and Yang13 concentrated on the development of reduced combustion kinetics mechanism for biodiesel to study the combustion characteristics of soy methyl ester (SME) and rapeseed methyl ester (RME) fuels. As explained here, these fundamental combustion studies focus mainly on SME and RME fuels. From the current available literature detailing on this, investigation into the combustion characteristics of biodiesel fuels such as CME and palm methyl ester (PME) and the complex interactions with the fuel properties has not been extensively conducted as to date. Therefore, the objective of this study is to examine the combustion characteristics of various biodiesel fuel types and blends as compared to diesel fuel. A CFD approach is used together with an experimental engine test-bed study to accurately capture the physical and chemical events of various neat biodiesels and their blends in a light-duty diesel engine. Here, a set of numerical experiments are performed to elucidate the key effects of CME, PME and SME at 0% (B100) diesel mixture, 50% diesel mixture (B50) and diesel (B0) on the combustion characteristics under different engine operating conditions. For this reason, the rates of heat release are


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evaluated under changing operating conditions in order to analyse the combustion characteristics of the biodiesel fuels. The heat release process can largely reflect the combustion and emission characteristics of a CI engine when additional information is available such as the ID period and engine load14. Here, the changes in spray structure, ID, peak PMC and MCC period between the tested fuels are extensively appraised.

2.

Heat-release rate (HRR) analysis This section explains the method of calculating the HRR profile and the assumptions made.

The heat release profile varies in terms of duration, magnitude and shape/phasing when the engine operating conditions changes. In this study, the engine speed is maintained at a constant value of 2000 rev/min. The engine load is varied from 0.5 kW to 1.5 kW and 2.5 kW. In addition, a total of 7 different fuel blends are investigated. As such, significant variations in the heat-release profile are expected throughout the test cases. Variations in the HRR profiles are classified under three distinct categories as shown in Figure 1. In each category, the combustion is divided into two main features: the PMC and MCC. The PMC phase is where the combustion rate is limited by the chemical reaction rate15. The MCC phase is where the combustion process is controlled by the locally available oxygen concentration, where diffusion flame is observed due to extended fuel injection timing. Here, the burn rate is controlled by the mixing process which is dictated by the turbulent intensity. Typically, fuels with long ID period tend to have more fuel mass burnt in the PMC phase. Similarly, fuels tend to undergo longer PMC phase relative to the MCC phase at low load due to the shorter fuel injection duration which leads to longer ID period as shown in Figure 1. Here, a higher rate of energy release takes place during the PMC phase as indicated by the high peak


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HRR value. The MCC phase is short and for some cases, it is negligible. Fuels with shorter ID period or high load conditions typically result in longer MCC phase. Due to the shorter time for fuel and air to mix (short ID period) or longer fuel injection timing to accommodate the higher fuel mass required, the PMC phase is negligible as compared to the MCC phase as can be seen in Figure 1. Finally, the third category of combustion falls in between these two extreme features, with balanced PMC and MCC phases. 55 PMC-Dominated PMC/MCC-Balanced MCC-Dominated

High peak PMC

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Heat-Release Rate / [J/CAD]

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Medium peak PMC 35 Balanced PMC and MCC phases

Low peak PMC

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Short or negligible diffusion-burn period

15

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Longer diffusion-burn period

-5 -15

-5

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35

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Crank Angle / [degrees]

Figure 1: Definition of combustion phasing utilised in this study.

The heat release due to combustion is determined using the First Law of Thermodynamics. Using the change in pressure and volume per unit of crank angle degree, the net rate of heat release is calculated using expression 1. dQnet dv 1 dp γ = p + V dθ γ − 1 dθ γ − 1 dθ

(1)


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where is the adiabatic exponent with a value of 1.35 employed in this study9, is the incylinder pressure and is the volume as a function of crank angle (CA). For the HRR computations, the following assumptions are made: •

The HRR analysis is limited to within the closed cycle of the engine cycle between intake valve closing (IVC) and exhaust valve opening (EVO).

3.

•

The effect of blow-by is negligible.

•

The cylinder wall temperature is constant within the analysis interval.

•

The in-cylinder pressure is uniform within the control volume.

Engine experimental setup and computational fluid dynamics (CFD) model formulations

3.1

Light-duty diesel engine experimental setup

Simulations are carried out for a quiescent, bowl in-piston light-duty diesel engine, which has a compression ratio of 19.1:1 and a displacement volume of 347cm3. Four equally spaced injector holes deliver the required fuel. The air cooled engine operates at speeds between 1500 to 3500 rev/min and loads between 0.5 to 2.5kW. The engine bore and stroke dimensions are given as 80 mm and 69 mm, respectively. Table 1 summarises the specifications of the research engine used in this study. The test engine is coupled to an asynchronous dynamometer mounted on an engine test stand. For the determination of pressure trace, the Kistler type 601A quartz pressure transducer is used. A K-type thermocouple with an upper measurement limit of 1000 oC and a maximum error of ±4 oC is used for the acquisition of the exhaust gas temperature5, 6. Other engine data such as torque, specific fuel consumption, equivalence ratio and volumetric efficiency are collected from the respective sensors contained within the engine test stand. Bosch


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BEA350 gas emissions analyser is employed to measure CO, CO2, HC and NO levels in the exhaust stream, whereas a Bosch RTM 430 smoke opacimeter is used to determine the smoke capacity level5, 6. For the emissions analyser, accuracy levels of ±0.001 vol.%, ±0.001 vol.%, ±1 ppm vol. and ±1 ppm vol. are achieved for the measurements of CO, CO2, HC and NO, respectively5, 6. A photodiode receiver in the smoke opacimeter is able to determine the smoke opacity up to an upper limit of 100% with an accuracy of ±0.1%5, 6. All experimental data is recorded at fully warm, steady state engine conditions. Further details on the experimental measurements and tests can be found in the work reported by Ng et al. 5, 6.

3.2

Computational mesh and initial conditions A 90o sector mesh is generated to take advantage of the symmetry imposed by the four

equally-spaced injector nozzle holes installed centrally. Hexahedral mesh type is employed to construct the engine mesh as shown in Figure 2. The initial conditions of intake pressure and temperature at IVC are 1.23 bar and 300 K, respectively. Here, the dynamic mesh layering utility of mesh topological change library (lib-ICE) 16, 17, 18 is utilised for the modelling of the compression and expansion strokes. The mesh, with cell size of 2.0 (radial) x 2.5 (axial) x 2.0 (azimuthal) mm, is used throughout this study as it is found to give similar accuracy but with shorter computing time as compared to finer resolution mesh19. Further details on the mesh configurations and grid independency study have been reported by Ismail et al.19. All simulations reported here are performed for the closed part of the engine cycle only.


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Figure 2: 90o sector mesh of the Nottingham Test Engine.

3.3

Computational model formulations The description of the computational model used which was implemented in OpenFOAM

open source CFD code under the Lib-ICE library developed by Tommaso et al.18 is provided in this section. Here, the in-cylinder flows and turbulence effects are calculated using the widely used RNG k-ε turbulence model19, 20. The heat transfer from the in-cylinder gasses to the chamber walls is modelled using the Han and Reitz correlation19, 21. The liquid spray structures are predicted using the Huh injector model to represent the injector and fuel droplets primary spray break-up18, 19. Fuel spray atomisation is modelled using the modified Huh and Gosman model as detailed in references18, 19, 22. Secondary break-up is modelled using the Kelvin-Helmoltz and Rayleigh-Taylor model19. Heat transfer between the discrete and gas phases is modelled using the Ranz-Marshall correlation16, 17. The Bai and Gosman correlation is used to predict the dynamics characteristics of droplets wall interactions on the piston bowl wall surface19. Further details about the spray model constants as well as model implementations and validations can be found in16, 17, 18, 19, 23. The spray characteristics of diesel fuels are defined by the thermo-physical and transport properties of C14H30 surrogate. The 9 ACS Paragon Plus Environment

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properties of CME, PME and SME developed by Ismail et al.19 are utilised for neat biodiesel fuels. For the B50 biodiesel-diesel blends, the properties of neat biodiesel are mixed with that of C14H30 using multi-component modelling approach at 50% blend ratio. For the combustion process, each computational cell is assumed to be a homogeneous system and a reduced chemical mechanism is used to estimate the PMC and MCC combustion phases. The homogeneous (PSR) approach is used instead of the non-homogeneous (PaSR) approach because PaSR requires considerably higher computing time and produces greater level of numerical instability. It is also found that there is no significant increase in the prediction accuracy when the PaSR approach is used. As such, the PSR approach is adopted here where the chemical reaction rates are calculated by an ordinary differential equation (ODE) stiff solver using an operator-splitting technique. For each time-step ∆t, the ODE solver maps the initial composition 0 0 to the reacted value Δ which is a unique function of 0 called the reaction mapping, . The composition array is defined by , … , , , , where Yi is the specie mass fraction, Ns the number of species, T the temperature and p the pressure. However, direct-integration of detailed chemistry is computationally demanding. For this reason, the proposed combustion model works in combination with the TDAC algorithm17, 19, 24. Such approach combines mechanism reduction and tabulation techniques that are both performed onthe-fly in each cell during the simulation. Significant speed-up factors compared to directintegration are ensured (>300), allowing semi-detailed mechanism to be used for simulations of practical devices such as CI engines. For the combustion chemistry, the 113 species and 399 reactions mechanism (BOS-V2) with integrated NOx reaction kinetics proposed by Ismail et al.25 is utilised to represent the CME, PME and SME fuel chemistry. In all the simulations, the different fuels are modelled by varying


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the mass composition ratio of saturated and unsaturated component with respect to individual fuel type19. When the biodiesel fuel changed from CME to PME and SME, the level of unsaturation for the fuel component is varied from 20% to 50% and 80%, respectively to match the typical unsaturation levels present in the biodiesel fuel. A modified Chalmers n-heptane mechanism19, 25, 26 is utilised to represent the combustion of diesel fuel. Finally, a phenomenological model is used to predict soot emissions. The Hiroyasu approach is used to describe the formation of soot, while its oxidation is estimated by the Nagle-Strickland and Constable model. In order to obtain comparable soot data set, both the experimental and simulated data set is normalised using expression 2, where is the input or output variable and

(λ1 , λ2 ) is the ranges of the function27. The data is normalised between the range of 0 and 1 such that the experimental and computed data are in the same scale.  Z i − Z i min xi = λ1 + (λ2 − λ1 ) max min  Zi − Zi

   

( 2)

4.

Results and discussion

4.1

Validation of simulated engine data against experimental results A total of 18 engine test points are utilised to validate and investigate the combustion

characteristics of a total of six biodiesel fuel type and blends in comparison with fossil diesel. The validated fuels are CME, PME and SME at 0% (B100) diesel mixture and 50% diesel mixture (B50). The selection of B50 biodiesel-diesel mixture is to allow a feasible number of test points to be conducted while still allowing sufficient differences to be observed in the test results. The engine speed is maintained at 2000 rev/min and the engine load is varied from low (0.5kW) to mid (1.5kW) and high (2.5kW) loads as given in Table 1. A medium engine speed is


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chosen because it lies within the typical operating range of on-road vehicles and also within the engine’s drivability limits. Meanwhile, the chosen engine loads are between the idling loads and maximum loads achievable for all the test fuels6. Figure 3 (a) - (f) show the comparison of in-cylinder pressure trace and HRR profile between experimental and simulated results of six test fuels at three different engine loads. Generally, good levels of accuracy are achieved in the prediction of the ID period, peak pressures, pressure traces and HRR profiles for all the test fuels. The change in combustion phasing from PMC to MCC phase due to prolonged fuel spray injection period is also clearly captured for all the simulations. Additionally, the ID period and peak pressure are accurately predicted for the entire tested loads within the range of ±0.2o CA. The maximum error between the experimental and computed peak pressure values across the test range is limited to below 4.5%. Furthermore, the timing of the peak pressure varies within ±1.5o CA of the experimental data for all the tested fuels and engine loads. Figure 4 (a) - (f) illustrate the calculated and measured NOx and soot emissions at EVO time. The majority of the test cases have a maximum percentage error between predicted and measured emissions data within an acceptable limit of 8.7% for NOx and 13.5% for soot. Both the NOx and soot trends for all the tests are also well predicted. Based on the analysis conducted as shown in Figure 4, it is found that the highest soot emission is given by CME, followed by PME and lastly SME for both the neat and B50 blends of the biodiesel fuels. This is mainly attributed to the unique combustion behaviour displayed by each fuel, the local in-cylinder temperatures and the concentrations of freely available O and OH radicals for soot oxidation. At low load, the influence of combustion characteristic on soot emissions is more apparent. Therefore, fuels undergoing mainly PMC such as neat SME have less soot emission since soot


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formation rate is low as typically observed in premixed flame. On the other hand, neat and B50 blend of CME consistently produce a reduction of exhaust NOx level, as well as giving the highest percentage of NOx reduction across all the engine loads tested. Among the B50 blends of biodiesel fuels, the use of B50 SME results in the highest NOx emission, followed by B50 PME and finally B50 CME. This trend is also applicable with the various neat biodiesel fuels for different engine load operations. Figure 4 also shows that an increase in NOx level at EVO is observed when the engine load is decreased, with highest NOx emission recorded at low engine load for different test fuels. At low load, the fuel injection period is short where fuel delivery is mostly completed before the start of combustion (SOC). Here, the combustion process is predominantly PMC, which burns more completely and results in elevated local in-cylinder temperature. This leads to greater NOx formation for all the test fuels at low load as compared to mid and high load operations. Further details of the CFD engine validations against experimental results are reported by Ismail et al.19. Based on the validation results, it can be concluded that the selected CFD models and model constants are able to accurately predict the in-cylinder combustion and emission characteristics with acceptable error tolerance limits.


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80 (a) CME(B50)

(c) SME(B50)

(b) PME(B50)

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-10

-10 -60 -50 -40 -50 -60 -30 -20 -10 -400

-30 -20 -10 00 -20 -50 -40 -10 10 20-3030 40 -60

Crank Angle Degrees [CAD]

Calc-0.5kW

Measured-0.5kW

10 201030 40 -60 -30 -20 -1040 0 20 -50 -40 30


Calc-1.5kW

Measured-1.5kW

10 20 30 40


Calc-2.5kW

Measured-2.5kW

Figure 3: Pressure trace and HRR for low, medium and high load at 2000 rev/min for (a) CME (B50), (b) PME (B50), (c) SME (B50), (d) CME (B100), (e) PME (B100) and (f) SME (B100).


100

0.5

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Emission Index NOx / [(g/s)/(kg/s)]

0.4 0.3

Normalised Soot / [-]

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Measured-NOx

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CME(B100) PME(B100) SME(B100)

Measured-Soot

Calculated-Soot

Figure 4: Emission index of NOx and normalised soot for (a) B50 blend at low load (0.5 kW), (b) B50 blend at mid load (1.5 kW), (c) B50 blend at high load (2.5 kW), (d) B100 at low load (0.5 kW), (e) B100 at mid load (1.5 kW), and (f) B100 at high load (2.5 kW).

4.2

Effects of biodiesel fuel types and blends on mixture preparation process One of the important factors that determine the mixture preparation process before the start

of combustion (SOC) is the start of injection (SOI) timing. In this study, the SOI timing is advanced by 0.2o and 1.2o CA when the diesel fuel is substituted with the B50 and B100 blends,


Energy & Fuels

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respectively for CME, PME and SME. In this engine set-up, the fuel injection system is based on a rotary-type fuel injection system. Here, the actuation of fuel injector valve is controlled by the fuel line pressure which is governed by the fuel pump. Typically, the system is calibrated for diesel fuel. However, diesel fuel has lower bulk modulus, surface tension and viscosity than biodiesel fuels. Therefore, the fuel line pressure increases at a higher rate when biodiesel fuels are used instead. As a result, the fuel injector valve opens at an earlier CA with biodiesel fuels as compared to when diesel is used. When the SOI is advanced, the fuels tend to undergo early SOC before TDC during the compression stroke. This leads to higher local and average incylinder temperatures and pressures due to both the combustion process and the reduction in control volume size during compression stroke. This condition is ideal to promote NOx formation. Fuel spray and vapour structure are also significantly affected when diesel fuel is replaced with B50 or neat CME, PME and SME during the mixture preparation stage. At low load, almost all the fuel mass is injected before SOC. From the simulation data, neat biodiesel fuels generally have longer spray penetration lengths than diesel which follow the trend shown in Table 2 and Figure 5 (a). Similar trends are also observed at mid and high engine loads as depicted in Figures 5 (b) and 5 (c). However, when CME, PME and SME are blended with 50 vol% of diesel fuel, the general trend as discussed above is not reproduced. Here, B50 blends of PME and CME exhibit lower liquid penetration length as compared to diesel at low engine load. At mid load operation, all the biodiesel fuels have lower liquid penetration length. Finally, all the biodiesel fuels produce longer liquid penetration length as compared to diesel at high load. The factors governing the observed trend are the ID, engine load, fuel thermo-physical properties and the timing at which the liquid length reaches stabilisation point. After the liquid length stabilisation


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Energy & Fuels

point, the majority of liquid droplets are rapidly fully vaporised due to combustion. A fuel that reaches this stabilisation point earlier tends to have shorter liquid penetration length.


Energy & Fuels

0.020

(a)

Liquid Spray Penetration Length / [m]

0.018

Most of the fuel fully injected before SOC

0.016 0.014 0.012 0.010 0.008 0.006 0.004 0.002 0.000 -12

-10

-8

-6

-4

-2

0

Crank Angle / [degrees] 0.020

(b)


0.018 0.016 0.014 0.012 0.010 0.008 0.006 0.004

Liquid length stabilisation point, where, after this point most of the liquid droplets fully vaporised rapidly due to combustion

0.002 0.000 -12

-10

-8

-6

-4

-2

0

2

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8

10

Crank Angle / [degrees] 0.020

(c)

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1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

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0.016 0.014 0.012 0.010 0.008 0.006 0.004 0.020 0.018 0.016 0.014 0.012 0.002 0.010 0.008 0.006 0.004 0.002 0.000 0.000

Liquid length stabilisation point, where, after this point most of the liquid droplets fully vaporised rapidly due to combustion

-12 -10 -12

-8

-10 -6

-4

-2 -8 0

2

4-6

6

8

-4 10

12

14 -2 16

18

200

Crank Angle / [degrees] Diesel(B0) CME(B100)

CME(B50) PME(B100)

PME(B50) SME(B100)

SME(B50)

Figure 5: Simulated liquid spray penetration lengths for the different test fuels and blends, with an engine speed of 2000 rev/min at (a) 0.5 kW, (b) 1.5 kW and (c) 2.5 kW.


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Energy & Fuels

The important thermo-physical properties which significantly affect the spray penetration length are determined to be surface tension, viscosity, liquid density, and vapour pressure19. Here, it is found that the atomisation quality of biodiesel fuels is poorer compared to diesel. Biodiesel fuel droplets are less affected by the in-cylinder aerodynamics forces due to the higher surface tension and viscosity that lead to larger droplets being formed. In addition, the normalised vaporisation rate of a single droplet of B100 and B50 biodiesels is much slower than that of diesel fuel as a result of lower vapour pressure19, given that the fuel mass injected into the system is identical. Consequently, a biodiesel fuel droplet loses mass at slower rate due to vaporisation as opposed to its diesel counterpart. As a result of the combined effects of the larger droplet size and lower vaporisation rate, biodiesel fuel droplets have greater momentum to traverse across the combustion chamber before SOC. Once auto-ignition takes place however, the spray penetration length is now governed by the enhanced vaporisation process due to the higher charge temperature produced from combustion. Here, the influence of ID and the time at which the liquid length reaches stabilisation point are more dominant in determining the spray penetration length of the fuel. As for vaporisation trends illustrated in Figure 6, neat biodiesel fuels have lower levels of vaporised fuel mass as compared to the B50 blends at low and mid loads. At high load however, all the fuels have similar levels of vaporised fuel mass. The reason for this is that at low load, almost all the fuel is injected before SOC. Hence, the vaporisation rate is highly dictated by the fuel mass flow-rate, individual fuel properties and saturated evaporation point. On the other hand, the fuel injection period is prolonged at high load and therefore, the vaporisation rate is greatly enhanced by the combustion process for all the test fuels. The fuel vaporisation rate is predominantly controlled by thermo-physical properties prior to the SOC event19. However, the


Energy & Fuels

vaporisation rate is mainly regulated by the rate of heat exchange between the fuel droplets and the high temperature gas mixtures after auto-ignition. 75 65

10

55 8

45

6

35 25

4

15 2

5

0

Vaporised Fuel Mass @ SOC / [%]

12

Ignition Delay / [degrees]

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

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-5

ID-0.5kW

ID-1.5 kW

ID-2.5kW

0.5 kW

1.5 kW

2.5 kW

Figure 6: Computed ID period and vaporised fuel mass at SOC for all the test fuels at three different engine loads.

4.3

Effects of biodiesel fuel types and blends on combustion characteristics

4.3.1 Effects on ID The type of fuel utilised in a CI engine significantly influences the ID. In this study, it is found that the ID period is shortened when diesel is substituted with neat CME and PME, while ID is increased when diesel is substituted with neat SME as illustrated in Figure 6. This trend is observed for all three engine loads. In contrast, ID periods shorter than that of diesel are obtained for B50 blends of CME, PME and SME. The observed trend can be attributed to two factors; first is the fuel thermo-physical properties and second is the fuel chemical composition. The thermo-physical properties typically control the droplet break-up rate, vaporisation rate, spray length, vapour length and fuel/air mixing prior to SOC. This is collectively known as the physical delay. Meanwhile, the chemical composition dictates the onset of energy release for the 20 ACS Paragon Plus Environment

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Energy & Fuels

ignition process. It is typically measured using cetane number, where a higher cetane value indicates a shorter ID period. The chemistry effect is commonly known as the chemical delay. Normally, the effects of chemical delay are more prominent than physical delay9. The chemical delay is controlled by the molecular composition of each fuel. Among the biodiesels, fuels with lower unsaturation levels tend to have shorter IDs. For example, neat CME with a 20% average unsaturation level by mass has the shortest ID while neat SME has the longest ID with an average of 80%19. Fuels with higher levels of unsaturation have longer ID periods because the carbon double-bond is relatively stronger than that of single bond. Hence, more energy is required to break this bond such that the oxidation process can be initiated. As a result, fuels with higher levels of unsaturation require longer time to gain sufficient energy to break all the doublebonds as compared to those with lower levels of unsaturation. Apart from fuel type, the fuel blend level also affects ID. Among the tested biodiesel fuels, the magnitude of reduction in the ID period is enhanced when B50 blends of CME, PME and SME are used instead of their B100 counterparts as shown in Figures 6 and 7. Here, the combined effects of physical and chemical delays are responsible for the observed trends. Firstly, the B50 blends have much finer fuel droplets compared to their B100 counterparts due to better overall atomisation. Secondly, the 50% lighter diesel component in the B50 blends undergoes preferential evaporation. This leads to a much higher vaporisation rate of B50 blends compared to B100 fuels. Consequently, the overall physical delay of B50 blends is reduced significantly compared to B100 fuels resulting in reduced overall ID period for all B50 blends. Furthermore, variations in ID are observed when the engine load is changed. The greatest reduction in ID period occurs at high load, while the lowest occurs at low load when diesel fuel is substituted with neat biodiesel fuels as can be seen in Figure 7. This means that at high load,


Energy & Fuels

the use of neat biodiesel enhances the magnitude of reduction in the ID period whereas at low load, the reduction in ID is moderated. Both Figures 7 and 8 illustrate this point. These observations are mainly due to the biodiesel fuel thermo-physical properties and fuel injection pressure. Firstly, as stated before, neat biodiesel fuels have higher viscosity and surface tension compared to diesel. The fuel spray penetration length, atomisation, droplets break-up rate and vaporisation are highly dependent on these properties. At high load, all the tested biodiesel fuels have longer liquid spray penetration lengths due to higher fuel injection pressure. The fuel droplets also tend to undergo greater atomisation rate when exiting the fuel injector due to the higher injection pressure. The combined effects of improved atomisation and longer spray penetration length result in a higher vaporisation rate at high load. Furthermore, better air utilisation of the surrounding air is achieved as a result of longer spray penetration length. Consequently, the physical delay is reduced significantly at high load. The overall magnitude of reduction in the ID period is thus enhanced at high load compared to low load when diesel is substituted with neat biodiesel fuels. 30 20

Ignition Delay Variations / [%]

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

10

Positive value indicates increase in ID as compared to diesel

0 -10 -20 -30 -40 -50 -60 -70

Negative value indicates reduction in ID as compared to diesel B0

CME(B50)

PME(B50) 0.5kW

SME(B50) CME(B100) PME(B100) SME(B100) 1.5kW

2.5kW

Figure 7: Variations in computed ID period when the base fuel (diesel) is substituted by biodiesels at different engine loads.


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Energy & Fuels

Nevertheless, the abovementioned trend is reversed when diesel is substituted with the B50 blends of CME, PME and SME. In these cases, the greatest magnitude of reduction in ID is achieved at low load while the lowest is achieved at high load as shown in Figure 7. This means that at low load, the use of B50 fuel blend enhances the reduction of ID while at high load, the reduction in ID is moderated as seen in Figure 8. Here, the surface tension and viscosity of B50 blends are lower than neat biodiesels due to the mixing with diesel. At any load, the B50 blends tend to have higher atomisation rate, fuel droplet break-up rate and vaporisation rate than neat biodiesels leading to shorter spray penetration length. When the fuel is injected at high injection pressure at high load, the B50 fuel droplets undergo enhanced atomisation and droplet breakup rate and thus lose the momentum to travel further into the combustion chamber. Poorer fuel/air mixing occurs as a result, which eventually increases the physical delay compared to neat biodiesel. Thus, a lower magnitude of reduction in ID is achieved at high load when the diesel is substituted with B50 biodiesel-diesel blends. To summarise, the engine load has significant influence on the level of changes to the ID period when the fuel blend level is varied, with opposing trends for B100 and B50 fuels.


Energy & Fuels

11 ID Period / [degrees]

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

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9 7 5 3 1 0.0

0.5

1.0

1.5 Load / [kW]

Diesel PME (B50) CME(B100) SME(B100) Linear (CME (B50)) Linear (SME (B50)) Linear (PME(B100))

2.0

2.5

3.0

CME (B50) SME (B50) PME(B100) Linear (Diesel) Linear (PME (B50)) Linear (CME(B100)) Linear (SME(B100))

Figure 8: Variations of computed ID period with engine load for all the test fuels at 2000 rev/min.

4.3.2

Effects on combustion phasing

HRR profile For all loads, combustion phasing changes when diesel is substituted with CME (B100) and PME (B100). Figure 9 (a) - (c) shows that longer periods of MCC phase are obtained with more fuel mass burned in the diffusion flame when diesel is substituted with CME (B100) and PME (B100). However, the combustion phasing is not changed significantly when diesel is substituted with neat SME. Here, most of the fuel mass is burned in the PMC phase similar to that of diesel fuel. Similar trends are also observed when diesel fuel is substituted with B50 blends as illustrated in Figure 9 (a) - (c). The trends are mainly caused by the differences in ID period of the biodiesel fuels. Due to higher unsaturation levels, neat SME and B50 SME blend tend to have longer ID periods than neat CME and PME as well as their B50 blends. Neat SME also has longer ID period than that of diesel as shown in Figure 6. As a result, SME fuels tend to have longer fuel injection periods before SOC which leads to most fuel mass being burnt during the PMC phase and thus, longer PMC phases. Conversely, neat CME and PME as well as their B50 24 ACS Paragon Plus Environment

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Energy & Fuels

blends have significantly shorter ID periods than that of diesel as shown in Figure 6. This leads to earlier SOC which results in substantial fuel mass burnt during the MCC phase and thus, longer MCC phase at all loads. In addition to fuel type, the engine load also significantly influences the HRR profile. From Figure 9 (a), it is found that all the fuels have significantly longer PMC phases at low load compared to mid and high loads. Negligible MCC periods are observed for neat SME and diesel fuel at low load. However, when the load is increased from low to high, there are clearly changes in the combustion profiles whereby a significant portion of the fuel is now burned during the MCC phase for all the fuels. The dominance of the MCC phase as the load is raised is illustrated in Figure 9 (b) and (c). The reason for these observations is that at low load, the fuel injection period is shorter. Almost all the fuel mass is fully injected before SOC, hence a significant portion of the fuel mass is burned during the PMC phase. On the other hand, the fuel injection period is much longer at high load to accommodate the greater fuel mass required. This causes the fuel to be burnt in the diffusion flame of the MCC phase. Additionally, the change in combustion phasing from PMC to MCC is enhanced at low load when diesel is replaced with neat CME and PME as well as their B50 blends. The change in combustion phasing is not significant at higher load since all the fuels tend to exhibit MCC phase due to the longer fuel injection period.

Peak PMC Value The peak PMC is affected by the vaporisation rate of the individual fuel and also by the ID. At low load, since almost all the fuel is injected before SOC, the PMC phase is mainly influenced by the quantity of the available combustible mixture. Fuels with higher vaporisation


Energy & Fuels

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rate (i.e. high volatility) have higher peak PMC and vice-versa as shown in Figures 9 (a) and 10. All B50 blends are also observed to have marginally higher vaporised mass up to the point of SOC compared to neat biodiesel fuels at low load as depicted in Figure 6. As such, B50 blends typically have higher peak PMC compared to their B100 counterparts. At mid and high load conditions, the fuels which exhibit longer ID have higher mass of fuel burnt (MFB) during the PMC phase. As the ID period is increased, more combustible mixture is available to be consumed over a short period of time thus giving rise to the high peak PMCs observed in Figure 9 (b) and (c).


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55 (a)-Load 0.5kW

50

Heat-release Rate / [J/CAD]

45 40

At low load B50 fuels have higher peak PMC than neat biodiesel due to higher vaporised mass at EOI

Peak PMC increases as ID prolonged

ID period of B50 blends are shorter than neat biodiesel fuels at low load

35 30 25 20 15

EOC CA between diesel and biodiesels is similar

10 Fully PMC

5

A

0 -5 (b)-Load 1.5kW

50 45


At higher load B50 fuels have lower peak PMC than neat biodiesel


40 35 30 25 20 15

Biodiesels have earlier CA of EOC as compared to diesel

10 5

-5 65

A

Combustion phasing changing from MCC to PMC

0 -15 (c)-Load -10 -52.5kW 0

5

10

15

B

20

25

30

35

40

45

50

55

60

65

55

60

65

At higher load B50 fuels have lower peak PMC than neat biodiesel

60 55


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Energy & Fuels

50 45 40 35 30


25 20

Biodiesels have earlier CA of EOC as compared to diesel

15 10

A

Combustion phasing changing from MCC to PMC

5 0

B

-5 -15

-10

-5

0

5

10

15 20 25 30 35 Crank Angle / [degrees]

40

Diesel(B0)

CME(B50)

PME(B50)

CME(B100)

PME(B100)

SME(B100)

45

Figure 9: HRR profile for the test fuels at (a) 0.5 kW, (b) 1.5 kW and (c) 2.5 kW.


50

SME(B50)

Energy & Fuels

Generally, there is no real correlation between engine load and peak PMC values. There is also no clear correlation between engine load and the ability of biodiesel fuel to affect the peak PMC. Nevertheless, there are clear trends when the fuel type and blend level are varied. Firstly, the highest reduction in peak PMC value is achieved by neat CME, followed by neat PME and finally, neat SME as shown in Figure 11. Similar trends are also observed when diesel fuel is substituted with B50 blends of CME, PME and SME. This can be attributed to the chemical composition of the biodiesel fuels. Both neat and B50 blends of SME have the highest unsaturation level which therefore produces a longer ID period. A longer ID period leads to a higher quantity of vaporised fuel mass available before SOC. At auto-ignition point, a higher quantity of combustible mixture is burned over a short time period giving rise to the high peak PMC seen in Figure 9 (a) - (c), Figure 10 and Figure 11. In contrast, neat and B50 blends of CME tend to have shorter ID period due to the low unsaturation levels. Consequently, a lower amount of vaporised fuel mass is available for combustion at SOC resulting in lower peak PMC.

65 60 55

Peak PMC / [J/θ]

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

50 45 40 35 30 25 20 15

Fuel Type 0.5kW

1.5kW

2.5kW

Figure 10: Computed peak PMC value for the test fuels at 0.5 kW, 1.5 kW and 2.5 kW.


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30

10

Peak PMC Variations / [%]

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Energy & Fuels

Positive value indicates increase in peak PMC as compared to diesel

-10

-30

-50

-70 Negative value indicates reduction in peak PMC as compared to diesel -90

B0

CME(B50)

PME(B50)

SME(B50) CME(B100) PME(B100) SME(B100) Fuel Type

0.5kW

1.5kW

2.5kW

Figure 11: Variations in computed peak PMC value when the base fuel (diesel) is substituted by biodiesels at different engine loads.

Combustion efficiency In CI engines, the highest energy efficiency is achieved when most of the heat release from combustion occurs near TDC9. The combustion efficiency of different fuels at various engine operating conditions can be identified using data extracted from the HRR profile, namely the combustion duration, end-of-combustion (EOC), crank-angle of 50% MFB (CA50) and crankangle of 90% MFB (CA90). Fuels with short combustion duration, high MFB near TDC, with early CA50 and CA90 are regarded as fuels with higher combustion efficiency. Prolonged heat release in the expansion stroke leads to more heat loss through heat transfer between the hot gases and the combustion chamber walls. In the expansion stroke, the hot gas mixture is exposed to the expanding control volume. This results in higher contact surface between the hot gases and the chamber walls. With higher wall surface exposed to the hot gases, a higher rate of heat loss occurs. Therefore, the combustion process is regarded as less efficient if a significant amount of energy is released at a relatively later stage of the expansion stoke. 29 ACS Paragon Plus Environment

Energy & Fuels

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In order to facilitate the subsequent discussion, the definitions and conventions used to analyse the combustion efficiency are detailed first. The combustion duration is defined as the CA period between SOC and 90% of MFB. The combustion duration is utilised to determine the combustion speed of the fuels. Meanwhile, the EOC is defined as the CA position with 90% of MFB (abbreviated here as CA90). The timings when 50% of MFB (CA50) and 90% of MFB (CA90) are utilised to map the combustion efficiency of each test fuel at different engine loads. In this study, neat biodiesels and B50 biodiesel blends are compared against the combustion efficiency of diesel fuel at different loads. Referring to Figure 12 (a) - (c), the timings of peak PMC, end-of-premixed-burn (EOPMB) and CA50 are advanced when diesel is substituted with neat CME, PME and SME as shown in for all loads. Similar observations are attained when diesel is substituted with B50 blends of CME, PME and SME. These trends are mainly due to the varying ID periods and SOIs of the fuels. Firstly, all the biodiesel fuels (neat and B50) except neat SME have shorter ID periods than diesel. Secondly, all the biodiesel fuels (neat and B50) have advanced SOI timings compared to diesel. On the other hand, the CA90 is function of both engine load and fuel type. Here, CA90 is advanced at mid load but at high load, CA90 is delayed for in all cases when diesel is substituted with biodiesel fuels. However, CA90 is delayed at low load when diesel is substituted with biodiesel fuels except for neat SME as illustrated in Figure 12 (d). This may be due to the dependency of the burn speed on the combustion phasing of each fuel. In terms of the overall combustion duration or speed, all neat biodiesels and their blends except neat SME have slower combustion speed compared to diesel when the combustion process is predominantly PMC. Furthermore, biodiesel fuels burn at a lower rate than diesel when the combustion process is


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Energy & Fuels

predominantly MCC. However, when the combustion process is balanced between PMC and MCC (typically at mid-load for this study), the biodiesel fuels tend to burn at a faster rate than diesel fuel. This is because each fuel has different combustion phasing at different loads due to its chemical composition and thermo-physical properties. Additionally, amongst the biodiesel fuels, the timings of peak PMC, EOPMB and CA50 are advanced the most when diesel is substituted with CME (B100) followed by PME (B100) and finally, SME (B100) as seen in Figure 12 (a) - (c). The trends are also similar for the B50 counterparts. Here, the ID is responsible with SME having the longest ID period followed by PME and finally, CME. It can be deduced that the CME fuels have the highest ability to reduce the peak PMC values. This attribute of CME may be one of the contributing factors to its ability to reduce exhaust NOx.


8

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13.8

EOPMB / [CAD, ATDC]

(a) 6 4 2 0 -2

(b) 8.8 3.8 -1.2 -6.2

-4 -6

-11.2

8

65

(c)

(d)

EOC,CA90 / [CAD, ATDC]

60

6

CA50 / [CAD, ATDC]

1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

CA Peak PMC / [CAD, ATDC]

Energy & Fuels

4 2 0

55 50 45 40 35 30 25 20

-2

15 -4

10 0.0

0.5

1.0

1.5

2.0

2.5

3.0

0.0

0.5

1.0

Load / [kW] Diesel CME(B100)

1.5

2.0

2.5

3.0

Load / [kW] CME (B50) PME(B100)

PME (B50) SME(B100)

SME (B50)

Figure 12: Variations in combustion characteristics of the test fuels: (a) timing of peak PMC, (b) timing of EOPMB, (c) CA50 and (d) CA90.

5.

Conclusions This paper reports the mixture preparation process and combustion characteristics of neat

and B50 blends of CME, PME and SME biodiesels as compared to diesel fuel. The differences in fuel properties have significant effects on the in-cylinder mixture preparation process thus affecting combustion and emission characteristics. It is found that the SOI timing is advanced for biodiesel due to the higher bulk modulus, surface tension and viscosity, which collectively increases the in-cylinder temperatures and promotes higher NOx. Biodiesel fuels are found to have larger fuel droplets size, longer spray penetration and lower vaporisation rate as compared to diesel. The liquid penetration length is observed to be highly influenced by the timing of the 32 ACS Paragon Plus Environment

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Energy & Fuels

SOC event. The vaporisation rates of CME, PME and SME are observed to be enhanced when B50 blends of the respective fuels are used compared to the neat biodiesels. The B50 blends of CME, PME and SME have higher vaporisation rates due to the faster evaporation of the lighter 50% diesel component in these multi-component fuel blends. The changes in fuel type from diesel to biodiesel change both the physical delay and chemical delay, and hence the overall ID period. As a result, variations in the combustion phasing between the fuels are seen. Generally, fuels with longer ID periods tend have higher fuel mass burned in the PMC phase and fuels with shorter ID period have more fuel mass burned in the MCC phase at a specific engine load. With the exception of neat SME, the ID is reduced when diesel is substituted with neat biodiesels or biodiesel blends at all loads. The reduction in ID magnitude is enhanced when B50 blends of CME, PME and SME are used compared to their B100 counterparts. The use of neat biodiesel enhances the magnitude of reduction in ID at high load while at low load, the magnitude of reduction in ID is moderated. This trend is reversed when diesel is substituted with B50 blends of CME, PME and SME. The ID period is reduced when the load is increased for neat CME, PME and SME. However, the ID increases as load is increased for the B50 blends of CME, PME and SME. Even though changes in engine load alter the combustion phasing, the combustion phasing is still dependent on the fuel type at each load. Fuels with long ID periods (i.e. neat and B50 blend of SME, diesel) have more mass of fuel burned during the PMC phase. Conversely, neat and B50 blends of both CME and PME have more fuel mass burned during the MCC phase. It is found that neat and B50 blend of CME have the highest magnitude reductions in peak PMC compared to diesel, whereas neat and B50 blend of SME have the lowest magnitude reductions in peak PMC. Fuel burn duration or speed is found to be highly dependent on combustion


Energy & Fuels

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phasing, which is dictated by fuel type and engine load. Biodiesel fuels burn at a slower rate compared to diesel when the combustion is either predominantly PMC or MCC phase. However, when the combustion periods of the PMC and MCC phases are similar, the burn speed is faster for biodiesel fuels compared to diesel.

Acknowledgements This work was supported by the Ministry of Science, Technology and Innovation (MOSTI) Malaysia under the grant no. 03-02-12-SF0045. The Faculty of Engineering at the University of Nottingham Malaysia Campus is also acknowledged for its support towards this project.

References (1) Ng, J.H.; Ng, H.K.; Gan, S. Recent trends in policies, socioeconomy and future directions of the biodiesel industry. Clean Technol. Environ. Policy. 2010, 12(3), 213-238. (2) Ng, H.K.; Gan, S. Combustion performance and exhaust emissions from the non-pressurised combustion of palm oil biodiesel blends. Appl. Therm. Eng. 2010, 30(16), 2476-2484. (3) Ayhan, D. Biodiesel production from vegetable oils via catalytic and non-catalytic supercritical methanol transesterification method. Prog. Energy Combust. Sci. 2005, 31, 466–487. (4) Charles, J.M.; Andre, L.B.; Glen, C.M. An Experimental investigation of the origin of increased NOx emissions when fueling a heavy-duty compression-ignition engine with soy biodiesel. SAE Papers. 2009-01-1792, 2009. (5) Ng, J.H.; Ng, H.K.; Gan, S. Engine-out characterisation using speed–load mapping and reduced test cycle for a light-duty diesel engine fuelled with biodiesel blends. Fuel. 2011, 90, 2700–2709. (6) Ng, J-H.; Ng, H.K.; Gan, S. Characterisation of engine-out responses from a light-duty diesel engine fuelled with palm methyl ester (PME). Appl. Energ. 2012, 90 (1), 58-67. (7) Chakravarthy, K.; McFarlane, J.; Daw, S.C.; Ra, Y. Reitz RD. Physical properties of soy bio-diesel & implications for use of bio-diesel in diesel engines. SAE Paper. 2007-01-4030, 2007. (8) Ra, Y.; Reitz, R.D.; McFarlane, J.; Daw, S.C. Effects of fuel physical properties on diesel engine combustion using diesel and bio-diesel fuels. SAE Paper. 2008-01-1379, 2008. (9) Heywood J.B, Internal Combustion Engine Fundamentals. New York: McGraw Hill Internaitonal, 1988. 34 ACS Paragon Plus Environment

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1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47 48 49 50 51 52 53 54 55 56 57 58 59 60

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(10) Yuan, W.; Hansen, A.C.; Zhang, Q. Predicting the physical properties of biodiesel for combustion modelling. Trans. ASAE. 2003, 6(46),1487-1493. (11) Yuan, W.; Hansen, A.C.; Zhang, Q. Vapor pressure and normal boiling point predictions for methyl esters and biodiesel fuels. Fuel. 2005, 84,943-950. (12) Brakora J.; Ra Y.; Reitz RD.; McFarlane J.; Daw CS. Development and Validation of a Reduced Reaction Mechanism for Biodiesel-Fueled Engine Simulations. SAE Paper. 200801-1378, 2008. (13) Junfeng, Y.; Golovitchev, V.I. Construction of combustion models for rapeseed methyl ester bio-diesel fuel for internal combustion engine applications. Biotechnol. Adv. 2009, 27, 641– 655. (14) Usman A.; Zheng M. Fast heat release characterization of diesel engine," Int. J. Therm. Sci. 2008, 47, 1688-1700. (15) Uwe H.; Rofl E.; Johansson B. Detailed Heat Release Analyses with Regard to Combustion of RME and Oxygenated Fuels in an HSDI Diesel Enigne. SAE Paper. 07PFL-353, 2007. (16) Ettorre, D.; Lucchini, T.; D'Errico, G. Comparison of combustion and pollutant emission models for DI diesel engines. SAE Paper. 07NAPLES-90, 2007. (17) D'Errico, G.; Lucchini, T.; Atzler, F.; Rotondi, R. Computational fluid dynamics simulation of diesel engines with sophisticated injection strategies for in-cylinder pollutant controls. Energ. Fuel 2012; 26, 4212–4223. (18) Lucchini, T.; D'Errico, G.; Ettorre, D. Numerical investigation of the spray/mesh/turbulence interactions for high-pressure, evaporating sprays at engine conditions. Int. J. Heat Fluid Flows. 2011; 32, 285-297. (19) Ismail, H.; Ng, H.K.; Cheng, X.; Gan, S.; Lucchini, T.; D'Errico, G. Development of thermophysical and transport properties for the CFD simulations of in-cylinder biodiesel spray combustion. Energ. Fuel 2012; 26, 4857–4870. (20) Ismail, H.; Ng, H.K.; Gan, S. Evaluation of non-premixed combustion and fuel spray models for in-cylinder diesel engine simulation. Appl. Energy. 2012, 90(1), 271-279. (21) Han, Z.; Reitz, R.D. A temperature wall function formulation for variable-density turbulent flows with application to engine convective heat transfer modeling. Int. J. Heat Mass Transfer. 1997, 40, 613-625. (22) Bianchi, G.M.; Pelloni, P. Modeling the diesel fuel spray breakup by using a hybrid model. SAE Paper. 1999-01-0226, 1999. (23) Karrholm, F.P.; Nordin, N. Numerical investigation of mesh/turbulence spray interaction for diesel applications. SAE Paper. 2005-01-2115, 2005. (24) Contino, F.; Jeanmart, H.; Lucchini, T.; D’Errico, G. Coupling of in situ adaptive tabulation and dynamic adaptive chemistry: An effective method for solving combustion in engine simulations. Proc. Combust. Inst. 2011, 33, 3057–3064. (25) Ismail, H.; Ng, H.K.; Gan, S.; Lucchini, T.; Onorati, A. Development of a reduced biodiesel combustion kinetics mechanism for CFD modelling of a light-duty diesel engine. Fuel, 2012, doi: 10.1016/j.fuel.2012.10.015. (26) Golovitchev VI.; Atarashiya K; Tanala K.; Yamada S.;. Towards Universal EDC-Based Combustion Model For Compression Ignited Engine Simulation. SAE Papers. 2003-011849, 2003. 35 ACS Paragon Plus Environment

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(27) Ismail, H.; Ng, H.K.; Queck, C.W.; Gan, S. Artificial neural networks modelling of engineout responses for a light-duty diesel engine fuelled with biodiesel blends. Appl. Energy. 2012, 92, 769-777.

Nomenclature ALMR B0 B50 B100 CA CAD CA50 CA90 CFD CO CO2 CME CI DAC HC EOC EOPMB EVO HRR ID ISAT IVC MCC ME MEDM MFB N2 NO NOx O2 PM PMC PME SMD (d32) SME SOC TDAC TDC Xi Zi λ1, λ2

adaptive local mesh refinement 100% diesel 50 % biodiesel-diesel mixture 100 % biodiesel crank angle crank angle degree crank angle 50% mass fraction burnt crank angle 90% mass fraction burnt computational fluid dynamics carbon-monoxide carbon dioxide coconut methyl ester compression ignition dynamic-adaptive chemistry hydrocarbon end-of combustion end-of premixed-burn exhaust valve opening heat-release rate ignition delay in-situ adaptive tabulations intake-valve closing mixing-controlled combustion methyl esters modified eddy-dissipation model mass fraction burnt nitrogen nitrogen-monoxide nitrogen oxides oxygen particulate matter peak-premixed combustion palm methyl ester sauter-mean diameter soybean methyl ester start of combustion tabulation of dynamic-adaptive chemistry top-dead centre normalised data maximum/minimum data normalisation range 36 ACS Paragon Plus Environment

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Tables Table 1: Engine specifications and test conditions for CFD simulation study. Specifications of Nottingham Test Engine Engine Type Light-duty diesel Piston Type Bowl-in-piston Cylinder Head Type Flat cylinder head Displacement Volume per-cylinder 347 cm3 Compression Ratio 19.1 : 1 Stroke 69 mm Bore 80 mm Connecting Rod Length 114.5 mm Intake Valve Closing (IVC) -140o ATDC Exhaust Valve Opening (EVO) 140o ATDC Engine Speed 1500 – 3500 rev/min Validation test range Load (kW) Speed (rev/min) Fuel (B100) Fuel (B50) 0.5 2000 CME CME 0.5 2000 PME PME 0.5 2000 SME SME 1.5 2000 CME CME 1.5 2000 PME PME 1.5 2000 SME SME 2.5 2000 CME CME 2.5 2000 PME PME 2.5 2000 SME SME

Table 2: Trends of liquid spray penetration length for various test fuels. Rank/Load 1 (Highest) 2 3 4 5 6 7 (Lowest)

0.5 kW PME (B100) SME (B100) CME (B100) SME (B50) Diesel PME (B50) CME (B50)

1.5 kW PME (B100) SME (B100) CME (B100) Diesel SME (B50) PME (B50) CME (B50)

2.5 kW PME (B100) SME (B100) CME (B100) SME (B50) PME (B50) CME (B50) Diesel


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