The effects of compression ratio and EGR on the performance and emission characteristics of diesel-biogas dual fuel engine

Accepted Manuscript The effects of compression ratio and EGR on the performance and emission characteristics of diesel-biogas dual fuel engine Saket Verma, L.M. Das, S.C. Kaushik, S.S. Bhatti PII: DOI: Reference:

S1359-4311(18)34700-8 https://doi.org/10.1016/j.applthermaleng.2019.01.080 ATE 13260

To appear in:

Applied Thermal Engineering

Received Date: Revised Date: Accepted Date:

30 July 2018 28 December 2018 24 January 2019

Please cite this article as: S. Verma, L.M. Das, S.C. Kaushik, S.S. Bhatti, The effects of compression ratio and EGR on the performance and emission characteristics of diesel-biogas dual fuel engine, Applied Thermal Engineering (2019), doi: https://doi.org/10.1016/j.applthermaleng.2019.01.080

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The effects of compression ratio and EGR on the performance and emission characteristics of diesel-biogas dual fuel engine Authors: Saket Verma*, L.M. Das, S.C. Kaushik & S.S. Bhatti Centre for Energy Studies, Indian Institute of Technology Delhi, India Correspondence: Author Tel.: +917827735765, +919179139159 Postal address: Engines & Unconventional Fuels Lab, Block-5, Centre for Energy Studies, Indian Institute of Technology Delhi, India, Hauz Khas, New Delhi, 110016, India E-mail address of corresponding author: *[email protected]

Abstract In this study, an experimental investigation on diesel-biogas dual fuel (DF) engine is presented based on energy and exergy analyses. The effects of change in compression ratio (CR), exhaust gas recirculation (EGR) and EGR temperature on the performance and emission characteristics of DF engine have been studied. In the first stage, engine was studied with increasing CRs of 16.5, 17.5, 18.5 and 19.5 in stepwise manner. It was found that the higher CRs were not only advantageous to the engine performance from first and second-law point of view but also to the exhaust emissions. In the second stage, DF engine was studied at the highest CR (19.5) and the effects of EGR were analysed. The engine was studied with EGR percentages of 5%, 10% and 15%, which caused slight improvements in engine efficiency at low load and simultaneous decrease in oxides of nitrogen (NOx) emissions. However, high EGR percentages at high loads showed slight decrease in engine efficiency. In the third stage, hot EGR was employed and the results obtained were compared with the cold EGR case. The results showed that the highest efficiencies both at low and high loads were obtained with hot EGR cases and at the same time exhaust emissions could also be kept in check. Key words: Biogas; Compression ratio; Dual-fuel; EGR; Exergy; Irreversibility 1. Introduction The energy demand has escalated sharply in the last few decades mainly driven by the strong economic growth and industrialization [1]. Most of the world’s energy demands are fulfilled by the fossil based energy sources, burning of which has considerably strained the ecological 1

stability and have brought about serious environmental changes [1, 2]. The petroleum and other liquid fuels provide highest energy share of the world’s total energy demand with transportation sector being the largest (~54%) consumer [1]. This indicates that transportation sector is one of the leading contributors in the present environmental instability and will lead in the future due to its fast-growing demand. The finite nature of fossil fuels and ecological damages caused by their consumption question the sustainability of a growth model solely based on them. Therefore, world’s attention has been shifted towards low-carbon fuels in the future fuel mix. The utilization of alternative and renewable fuels in the existing and future prime movers will play a key role in defining a sustainable development model. In this context, biogas is one of the important renewable fuels, especially in the developing countries [3, 4, 5]. Among the number of pathways of energy exploitation from biogas, its utilization in internal combustion engines (ICEs) is of prime interest [6]. The conventional diesel engines can be modified to work with biogas in dual fuel (DF) mode in which gaseous fuel (biogas) acts as the main fuel, i.e. provides major energy share, and liquid fuel acts as the pilot fuel, i.e. used in small quantities [7, 8]. The purpose of the pilot fuel is to ignite the gaseous fuel-air mixture and initiate the combustion so that it can further be sustained by combustion of the main fuel [8]. The DF operation of a conventional diesel engine has many advantages, such as good efficiency, low emission and fuel flexibility depending upon the types of fuel used [9]. In addition to that, with little modifications, it is possible to target the widely accepted diesel engine technology, and therefore maximize the scope of renewable fuel usage. However, there persist many technical challenges in the development of a DF technology. In the case of diesel-biogas DF technology, the major difficulties lie with relatively low conversion efficiency, high hydrocarbon (HC) and carbon monoxide (CO) emissions and limited percentage of gaseous fuel utilization [9, 10]. The biogas, which is typically a mixture of methane (CH4), carbon dioxide (CO2) and small percentages of other gases has very low calorific value. In addition to that the presence of inert gases like CO2 reduces the in-cylinder temperature by absorbing some amount of heat causing poor combustion [10, 11, 12]. Some of these problems, however at the different levels, have been reported with various fuel combinations in a DF engine. In previous studies, various methodologies have been implemented to enhance the performance and emission characteristics of a DF engine. Yang et al. [13] investigated the effects of pilot injection timing on the particle emissions and combustion noise of a diesel2

natural gas dual-fuel engine. It was found that by advancing the injection timing both performance and emission characteristics can be improved. However, combustion noise was found as the limiting factor, which also deteriorated with the advanced injection timing. It was also found that particle emissions were very sensitive to injection timing and can be reduced with advancement. The effects of pilot fuel quantity was studied by Abagnale et al. [14], and it was shown that 50% of methane ratio was optimum for pollutant and CO2 control. Banapurmath et al. [15] analysed the effect of compression ratio (CR) on a biodiesel-CNG DF engine. The experiments were performed by varying the CR from 15 to 17.5 keeping all the other operating parameters (such as injection timing, injection pressure, flow rates etc.) constant. It was reported that increasing the CR resulted in higher cylinder pressure and heat release rates, which rendered improvements in the brake thermal efficiency. By increasing the CR from 15 to 17.5, 1.4% improvement in the brake thermal efficiency was observed. Along with that HC, smoke and CO emissions were considerably decreased, whereas oxides of nitrogen (NOx) emissions were slightly increased. Sayin and Gumus [16] experimentally investigated the effects of CR and injection parameters on a biodiesel-diesel blended directinjection diesel engine. It was found that by increasing the CR from 17 to 18 and 19, brake thermal efficiency was increased by 0.55% and 1.33% respectively. It was reasoned that with higher CRs, the fuel was injected at hotter region leading to improvement in the combustion process. The improvement in thermal efficiency and lower exhaust gas temperature with increased CR have also been reported by Muralidharan and Vasudevan [17] with waste cooking oil as biodiesel and diesel blend in a variable compression ratio diesel engine. This investigation also revealed that at increased CR, biodiesel-diesel blend showed higher combustion pressure, higher rate of pressure rise and lower heat release rate as compared to that with the diesel. The thermodynamic analysis on the effect of varying the CR was studied by Debnath et al. [18] with palm oil methyl ester as the biodiesel fuel. They found that higher CRs showed increase in shaft power, but also increased the energy loss to cooling water and exhaust gases. However, on exergetic terms the variations of these losses were found very small. Therefore, at the higher CRs, exergy destruction was decreased showing better thermodynamic performance of the engine. Bora and Saha [19] studied a biogas run diesel engine in DF mode by varying the injection timing (26°, 29° and 32° before top dead centre (BTDC)) and CR (18, 17.5, 17 and 16). The results indicated that the highest brake thermal efficiency of 25.44% was obtained at the CR of 18 and the injection timing of 29° BTDC. However, at these operating conditions both 3

NOx and CO2 emission were also highest. Masood et al. [20] investigated on the performance and emission characteristics of a diesel-hydrogen DF engine with different CRs. They reported that the percentage change in brake thermal efficiency when the CR was changed from 16.35 to 24.5 was about 27% at 90% hydrogen substitution and full load. At the same time reductions in HC, CO and particulate matter emissions were 17%, 21% and 16% respectively, however, NOx emissions were increased by 38%. Effect of CR was also investigated by Gnanamoorthi and Devaradjane [21] on a direct-injection diesel engine with various ethanol-diesel blends. They found that the highest compression ratio (19.5) generated the highest efficiency. The increase in CR also improved a range of engine performance parameters along with reduced HC and CO emissions, however, NOx emissions were still reported to be high. Selim [22] examined the effect of CR on combustion noise, knock and ignition limits of a DF engine with LPG, methane and compressed natural gas (CNG) as the main gaseous fuels and diesel as the pilot fuel. It was observed that increasing the CR generally increases the combustion noise, especially with the gaseous fuels of low selfignition temperature. The descending order of combustion noise from DF engine was LPG, methane and CNG. Therefore, even though the literature suggests better engine performance with higher CRs, the high NOx emissions remains a major constraint. In one of the studies, Laguitton et al. [23] even suggested to reduce the CR along with the injection timings to achieve very low NO x emissions in a premixed charge compression ignition (PCCI) diesel engine; however, the fuel consumption was increased, especially at the high loads. One of the methodologies to curb NOx emissions from high CR DF engines could be the use of EGR. Based on a two-zone phenomenological model, Papagiannakis [24] investigated the effect of inlet air preheating and EGR on a natural gas DF engine. It was found that increasing the inlet air temperature improved the engine efficiency, however, the in-cylinder pressure was also significantly high posing the risk of mechanical damage. With the use of EGR not only the in-cylinder pressure was controlled but also the NOx and CO emissions were reduced. Abdelaal and Hegab [25] reported that the use of EGR on a natural gas-diesel DF engine reduced the peak cylinder pressure and hence prolonged the engine life. Hu et al. [26] studied the effect of EGR and hydrogen addition in a natural gas operated spark ignition engine. It was found that increase in EGR first increased and then decreased the thermal efficiency of engine. At high EGR, addition of hydrogen in natural gas was found as a favourable approach to achieve better efficiency along with low emissions. It was explained that at high EGR condition, coefficient 4

of variation of indicated mean effective pressure and cycle by cycle variations were significantly increased, which can be controlled by addition of hydrogen and therefore a stable low-temperature combustion can be realized [27, 28]. A quasi-two-zone combustion model was proposed by Hosseinzadeh et al. [29] to study the second-law analysis of dieselnatural gas DF engine under part load conditions. It was proposed to study the different cases of chemical, radical and thermal EGR on the performance of DF engine at low load. They found that without EGR, 28% of total input chemical exergy was wasted as unburned fuel in the exhaust. Furthermore, radical, thermal and their combined cases of EGR were found to have positive effects on availability terms, whereas, chemical case showed the opposite effect. Similar results were reported by Jafarmadar and Nemati [30] using a simulation model to study the effects of EGR fractions on hydrogen-diesel DF operations. As EGR fractions were increased from 0% to 30%, the exhaust exergy was increased from 14.9% to 56.7% and the exergy efficiency was decreased from 42.4% to 14.1%. Tomita et al. [31] studied the effect of EGR on combustion and emission characteristics of diesel-natural gas DF engine. It was reported that when the EGR rate was increased, the flame area and the luminescence intensity of the pilot fuel flame were decreased. The high EGR rate caused the ignition delay to be lengthened and the start of heat release was delayed. Ghazikhani et al. [32] investigated the relationship between irreversibility and brake specific fuel consumption in an indirect injection diesel engine with various EGR ratios. It was found that at low load and low engine speed, induction of EGR led to considerable decrease in the total in-cylinder irreversibility, however, at high load the effect was opposite. At all the EGR ratios, irreversibility and brake specific fuel consumption showed similar trends. The effect of EGR on a DF (diesel-natural gas) HCCI mode was numerically studied by Jafarmadar et al. [33] based on exergy analysis. They found that with increase in EGR, high specific heat of EGR components and decreased oxygen concentration led to reduction in the heat release and therefore poor combustion. At the 30% EGR condition, exergy efficiency and burned fuel were decreased by 41.3% and 37.7% respectively causing 29% increase in irreversibility. Ismail and Mehta [34] found that as the EGR fraction was increased, the exergy destruction was also raised mainly due to lower product temperature. Though, this was also translated to reduction in product chemical exergy losses resulting in the improvement of exergy efficiency. Based on the literature review, it has been found that the variation in CR and EGR addition are possible techniques to improve the engine performance and control the emissions 5

respectively. In a few reported studies with DF engines, these techniques have been independently investigated with gaseous fuels, such as methane and hydrogen. However, biogas shows significantly different thermo-physical properties as compared to the fuels reported in the literature. Therefore, it is necessary to investigate the effects of these parameters on diesel-biogas DF engine in detail. It has been shown in this work that it is possible to improve the diesel-biogas DF engine performance without compromising with the emissions by optimising the CR and EGR. The results are presented with energy and exergy analyses, which has received little attention in the literature but gives vital information about the scope of improvements. The experimental study has been carried out to encompass the variations in engine parameters, such as CR, EGR fraction and EGR temperature. The emissions analyses for HC, CO, smoke and NOx emissions have also been performed for above engine parameters. 2. Methodology 2.1.Test rig The tests were conducted in a four-stroke, single-cylinder, direct-injection modified diesel engine. The engine comes with the original CR of 17.5, injection timing of 23 °BTDC and constant engine speed of 1500 rpm. It has a bowl-in-piston combustion chamber and MICO BOSCH, 3-hole diesel injector. The major specifications of the test engine are given in Table 1. The engine was coupled to an AC dynamometer, which was connected to an electrical loading panel to control and vary the engine loads. The electrical loading panel consists of a combination of lighting and heating load arrangements, load switches and digital indicators. The pilot liquid fuel (diesel) flows under the gravity from an overhead tank to a high pressure pump and diesel injector. The gaseous main fuel that is biogas is composed of 75.2% CH4 and remaining CO2 and minor amount of impurity gases by volume percentage. The biogas was stored in a high pressure cylinder during the experiments. A two-way pressure regulator was used to reduce the pressure and gas was supplied to the flame trap. The main purpose of the flame trap is to serve as a safety device against backfire, however it also helps in stabilizing the pressure fluctuations. A gas mass flow meter (Proline Promass 80A) based on Coriolis effect was used to measure the biogas flow rates. An air flow meter (Endress + Hauser (EH)) along with an air-box was used to precisely measure the air flow rates to the engine. An EGR system was used for the test conditions with variable EGR rates. It is a double-pipe heat exchanger in which cold water was used to reduce the exhaust gas temperature and the flow rate was regulated from a control valve. The intake air, gaseous fuel 6

and EGR were mixed in the intake manifold of the engine and being supplied to the engine. The schematic diagram of the test rig is shown in Fig. 1. The exhaust gas analysis was done using AVL DIGAS exhaust analyser. The emission measurements of HC, CO and CO2 is based on non-dispersive infrared (NDIR) technique, while O2 and NOx emissions are measured based on electrochemical technique. The smoke emission presented in this work was measured using the smoke meter (AVL 437) and shown in smoke opacity percentages. The in-cylinder pressure data was obtained using Kistler’s (6613CQ18) piezo-electric pressure transducer mounted on engine cylinder head. A combustion analysis kit (Legion Brothers) consists of TDC angle encoder, rpm counter, charge amplifier and software interface was also used to analyse the obtained data. Table 1. Engine specifications Parameters

Technical specifications

Make & model

Kirloskar TAF1

Type

Single cylinder

Bore & stroke

87.5 × 110 mm

Swept volume

661 cm3

Compression ratio

17.5:1

Rated brake power

6 bhp / 4.4 kW

Rated speed

1500 rpm

Injection pressure

200 bar

Inlet valve open

4.50 °BTDC

Inlet valve closed

35.50 °ABDC

Exhaust valve open

35.50 °BBDC

Exhaust valve closed

4.50 °ATDC

Standard fuel injection timing

23.0 °BTDC

7

Fig. 1. Schematic diagram of the experimental test rig 2.2.Test cases and conditions The study presented in this article has been divided in three parts. In the first part, only the effect of change in CR on a DF engine was studied. The performance and emission characteristics of the engine were studied at four different CRs of 16.5, 17.5, 18.5 and 19.5 with varied engine loading conditions. The CR of the engine was varied by changing the thickness of the metal gasket placed between the cylinder head and engine block. The reduction in the thickness of metal gasket reduces the clearance volume and consequently increases the CR. Similarly, by increasing the thickness of metal gasket, the lower CR can be obtained. In the second part, the CR of the engine was fixed to 19.5 and the effect of various percentages of cold EGR addition was studied at different engine loads. In case of cold EGR, the recirculated gas temperature was controlled at 35±2 °C. It was achieved with the help of a carefully designed double-pipe counter flow heat exchanger in which the exhaust gas and cooling water flow in opposite directions. The desired range of exhaust gas temperature was achieved by controlling the rate of cooling water supply in the heat exchanger while the EGR operations were studied. Here, three cases of cold EGR percentages viz. 5%, 10% and 15%

8

were analysed and presented at different engine loads, and also compared with the no EGR case (shown as 0% EGR). In this study, the EGR percentage is calculated on mass basis as the ratio percentage of mass flow rate of EGR ( m EGR ) to the mass flow rate of total intake charge ( m in ):  m  EGR (%)   EGR   100  m  in 

(1)

where, the mass flow rate of total intake charge ( m in ) includes the mass flow rate of air ( m air ), mass flow rate of gaseous fuel ( m gas ) and mass flow rate of EGR ( m EGR ). In the third part, the effects of hot EGR on the DF engine performance and emission features were studied. In order to supply hot EGR, the recirculated gas temperature was controlled at 65±2 °C. Here three different cases (i.e. 5%, 10% and 15%) of hot EGR were analysed and compared with the cold EGR cases at two different engine loading conditions of low (BMEP = 1.16 bar) and high (BMEP = 5.32 bar) loads. Here, the high load also signifies the full load condition that corresponds to rated peak power of the present engine. All the other test conditions such as injection parameters, valve timing and engine speed were kept constant across all the cases studied. 2.3.Experimental method The DF operation of the engine was studied with diesel as the pilot fuel and biogas as the main fuel at variable engine loads. Initially the engine was started with the conventional single fuel diesel direct injection. Thereafter, the biogas flow was started by gradually opening the control valve to achieve DF operation of the engine. As biogas flow rate was increased, diesel flow rate was simultaneously decreased by the engine governing system to maintain the engine speed. The DF operation requires a minimum amount of pilot fuel and therefore, biogas flow rate can only be increased up to a certain point. The condition where maximum amount of biogas and minimum amount of diesel fuel is used in the DF operation is termed as maximum diesel substitution. All the tests were performed at steady state, constant engine speed (1500±20 rpm) and maximum diesel substitution conditions. The maximum diesel substitution is defined as:  m D, D  m D, DF  DS (%)     100 m D, D  

(2)

9

where, m D, D is the diesel flow rate in conventional single fuel diesel only operation and m D, DF is the diesel flow rate in dual fuel operation.

At this point of engine operation, the observations were recorded for diesel flow rate, biogas flow rate, air flow rate, engine speed, power output, exhaust emissions and combustion data. The CR of the engine was changed in stepwise manner and the same experimental procedure was repeated. In the second part of cold EGR operation, the amount of EGR was controlled with the help of an EGR valve. The flow rate of EGR was calculated with the help a differential U-tube manometer. The cooling water valve was adjusted to attain the required temperature of cold EGR. The exhaust gas, EGR, oil, air and ambient temperatures were measured with the help of K-type thermocouples. The hot EGR cases were achieved by reducing the cooling water flow rate in the EGR system. The uncertainty values associated with various measured and calculated results are shown in Table 2. The uncertainty in the measured quantities was evaluated based on root-mean-square deviation technique, while propagation of uncertainty method was applied for the evaluation of calculated quantities [35]. Table 2: Uncertainty of measured and calculated quantities Measured quantities

Uncertainty (%) Resolution

Flow rate of air

1.6%

±0.1 m3/h

Flow rate of diesel (diesel only mode) 1.8%

±0.2 cm3

Flow rate of diesel (DF mode)

2.7%

±0.2 cm3

Flow rate of biogas

1.7%

±0.25%

Brake power

1.0%

1W

Temperature

0.2%

1 °C

NOx emission

2.7%

1 ppm by vol.

HC emission

3.1%

1 ppm by vol.

CO emission

2.9%

0.01% by vol.

Smoke emission

2.5%

± 0.05%

Calculated quantities

Uncertainty (%)

Diesel substitution

1.3%

Energy efficiency

1.7%

Exergy efficiency

2.1%

10

Exergy destruction

2.7%

Exhaust gas exergy

1.9%

Table 3. Properties of diesel and biogas used in dual fuel operation Properties

Diesel Biogas

Lower heating value (MJ/kg)

43.0

27.22

Specific chemical exergy (MJ/kg)

45.8

28.64

Stoichiometric A/F ratio

14.5

8.93

Flame speed (cm/s)

30

25 [36]

Auto-ignition

257

600-650 [36]

(kg of air/kg of fuel)

temperature (°C) Density at 1 atm and 15 °C (kg/m3) 848

0.928

Specific heat (Cp – kJ/kgK)

1.35

-

2.4.Energy and exergy analyses The energy and exergy analyses presented in this work have been performed under the following assumptions: 1. The engine along with the dynamometer is considered the control volume. 2. The reference atmospheric conditions are: pressure ( P 0 ) is 1 atm (101.325 kPa) and temperature ( T 0 ) is 25 °C (298.15 K). The composition of elements in the reference atmosphere is taken as suggested by Kotas [37]. 3. The intake air, air-fuel mixture and exhaust gases are considered ideal gas mixtures. 4. The kinetic and potential energies of incoming and outgoing masses are neglected being very small. In the presented thermal system, the energy and exergy exchange take place in various forms. The energy/exergy enters the system in the form of chemical energy/exergy of fuels (pilot + main fuels). A part of this energy/exergy is converted to useful work and remaining is transferred to the surrounding in different forms, such as heat transfer through cylinder walls, exhaust gases and unaccounted energy/exergy.

11

The rate of energy input ( E in ) to the system is given as the sum of energy carried by diesel and biogas:  D  LHVD )  (m  BG  LHVBG )] E in  [( m

where,

 m

is the mass flow rate and

(3) LHV

is the lower heating value of the incoming fuel. The

lower heating values of diesel and biogas is given in Table 3. Similarly, the rate of exergy input ( X in ) to the system is given by:  D  xch, D )  (m  BG  xch, BG )] X in  [( m

(4)

where, xch, D and xch, BG are the specific chemical exergies of incoming diesel and biogas fuels respectively. The specific chemical exergy of liquid diesel fuel was taken from the correlation given by Kotas [37] as:  H O S   H     xch, D  LHV  1.0401  0.1728   0.0432   0.2169   1  2.0628   C C C C             

(5)

where, hydrogen (H), carbon (C), oxygen (O) and sulfur (S) are given by their molecular masses in the diesel fuel. The specific chemical exergy of biogas was calculated considering it an ideal gas mixture and given as suggested by Kotas [37]: x ch, BG   y i x ch,i   R T 0  y i ln y i i

where,

yi

(6)

i

and

xch, i

are the mole fraction and standard specific chemical exergy of the ith

species in the ideal gas mixture of biogas respectively;

R

shows the universal gas constant.

The rate of useful work output ( EW ) from the system can be written as: EW  X W  BP

(7)

here, BP represents the brake power output from the engine. It can be noted that as the work output is a high grade energy, it also represents the equivalent amount of exergy as shown in Eq. (7). The energy and exergy efficiencies are given by Eq. (8) and Eq. (9) respectively.  I  ( EW / E in ) 100

(8)

 II  ( X W / X in ) 100

(9)

12

A significant amount of exergy is also carried away by hot exhaust gases leaving the engine system. The rate of exergy flow ( X ex ) with outgoing exhaust gases is given by [38]:







  T   P  X ex  m ex  C P,ex  Tex  T 0  m ex  T 0  C P,ex  ln ex   R  ln ex  0 T   P 0   

(10)

It can be noted that Eq. (10) represents only the physical exergy of exhaust gases. The chemical exergy part of the engine exhaust gases has been found irrecoverable by the researchers [39, 40] and therefore it has been considered as exergy destruction in this work. In Eq. (10),

m ex , Tex

and

Pex

are mass flow rate, temperature and pressure of exhaust gases

respectively. The rate of exergy transfer with wall heat losses ( X Q ) from the engine is given by [38]:  T0 X Q   1   Tg 

 E Q  

(11)

where, Tg is the instantaneous in-cylinder temperature and E Q is the instantaneous heat transfer rate from the engine. The instantaneous heat transfer rate was calculated using Annand’s correlation [41]. Based on the second-law of thermodynamics, the exergy balance can be written to find out the rate of exergy destruction ( X des ) in the system as [38]: X des  ( X in )  ( X W  X ex  X Q )

(12)

In the IC engine system, there are various irreversible processes which are carried out with some amounts of exergy destructions. One of the major processes is combustion [42, 43, 44], which converts the chemical energy of the fuel to the thermal energy. The entropy generation during in-cylinder combustion process ( S gen,comb ) can be given by [40]:  i, p si, p   m  j ,r s j ,r   ( E Q / Tg ) S gen,comb   m

(13)

where, the first term in the right hand shows the sum of entropies of burned gas product species, whereas, the second term shows the sum of entropies of reactant species. The product of combustion is comprised of exhaust gas species, which was determined from the exhaust gas analysis. The reactant mixture is comprised of air, fuels and EGR species. The last term shows the entropy generation due to heat exchange between the system and the surrounding. The specific entropies of various species were calculated using the polynomial 13

curve fit obtained from JANAF table [41, 45]. Consequently, the exergy destruction, also called irreversibility ( Icomb ) due to combustion process can be given by [38]: Icomb  T 0  S gen,comb

(14)

The irreversibility also takes place due to mechanical friction in IC engine processes. The frictional irreversibility ( I fri ) was calculated by subtracting the brake power output from the indicated power output of the engine. Another source of irreversibility is related to mixing process, owing to its inherent nature of entropy generation. The mixing of air with gaseous fuel and EGR in the intake manifold and also with the liquid fuel causes entropy generation. The mixing irreversibility ( Imix ) was calculated by [37]: n

Imix   R T 0  ni ln xi

(15)

i

where,

ni

is the number of moles in the

i th

species. The unaccounted irreversibility ( Iun ) was

calculated by subtracting the accounted sources of irreversibilities from the total exergy destruction as given by: Iun  X des  ( Icomb  Imix  I fri )

(16)

3. Results and discussion 3.1.Effect of compression ratio (CR) Variation in the fuel flow rates (both diesel and biogas) with engine load is shown in Fig. 2 for different CRs. The fuel flow rates increase with increase in engine load to satisfy the high energy demand. It was also found that the biogas flow rate was always higher than the diesel flow rate at the same operating conditions because of the high pilot fuel substitution and lower calorific value of biogas. The effect of increase in CR was found positive on fuel demand as it lowered the fuel flow rates. It also improved the diesel substitution causing decrease in pilot fuel requirements. The effect of change in CR on diesel substitution is depicted in Fig. 3. Significant improvement in diesel substitution was found when the CR was increased from 16.5 to 19.5. In DF mode, minimum quantity of pilot fuel is required to initiate the combustion that is sustained by burning of the main fuel. In the present case with biogas as the main fuel, its low flame speed limits the flame propagation making it susceptible to misfire. Therefore, slightly higher amount of pilot fuel is required to improve the combustion characteristics of a biogas DF engine. It was found that increasing the CR 14

increases in-cylinder temperature, which improves the flame speed of combustible mixture. Moreover, higher CR causes increased temperatures at the end of compression process, which results in lower ignition delays and helps in lowering the pilot fuel requirements. The improvements in diesel substitutions were higher at the higher loads as compared to that at the lower loads. The diesel substitutions were increased by 7.19% and 11.66% by increasing the CR from 16.5 to 19.5 at the BMEPs of 1.16 bar and 5.32 bar respectively. At these operating conditions, the highest diesel substitutions were observed at the CR of 19.5 that were 80.24% and 67.28% respectively. 0.5 0.45 Diesel- CR=16.5

Fuel flow rates (g/s)

0.4

Diesel- CR=17.5

0.35 0.3

Diesel- CR=18.5

0.25

Diesel- CR=19.5

0.2

Biogas- CR=16.5

0.15

Biogas- CR=17.5

0.1

Biogas- CR=18.5

0.05

Biogas- CR=19.5

0 1

2

3

4

5

6

BMEP (bar)

Fig. 2: Fuel flow rates of diesel and biogas with engine loads for various compression ratios. CR=16.5

CR=17.5

CR=18.5

CR=19.5

4

5

85

Diesel substitution (%)

80 75 70 65 60 55 50

45 40 1

2

3

6

BMEP (bar)

Fig. 3: Diesel substitution with engine loads for various compression ratios. 15

The effect of CR variation on energy efficiency is shown Fig. 4. Increase the CR was found to be favourable, which showed improvements throughout the engine load spectrum. The CR of 19.5 showed highest energy efficiencies as compared to other CRs at all the engine loads. The increase in CR from 16.5 to 19.5 resulted in 2.63% and 3.08% improvements in energy efficiencies at BMEPs of 1.16 bar and 5.32 bar respectively. One of the key reasons for this improvement is the higher average cycle-temperature caused by high compression ratio. The higher cycle-temperature indicates higher Carnot efficiency and therefore, better energy conversion capability of the thermal system. In addition to that the improved flame speed at higher CRs also assists in combustion of the end-gas-mixture. The increase in compression temperature with higher CRs renders lower ignition delay and shorter combustion duration enhancing the combustion process. The improved combustion process at elevated temperature results in better conversion efficiency. Variations in exergy efficiency and exergy destruction with various CRs are depicted in Fig. 5. Exergy efficiency increases with increase in both the engine load and CR. It is also evident that the increase in exergy efficiency is accompanied by decrease in exergy destruction. At low load, low in-cylinder temperature results in poor combustion and poor conversion of input fuel energy to shaft work. This is because at low load small amount of diesel is required, which is further reduced with DF operation as replaced by the main fuel. This small amount of diesel may not create sufficient number of ignition centres to ignite the surrounding gas-air mixture. The consequence of this effect is incomplete combustion and poor exergy efficiency. Nevertheless, at high load, high amount of pilot diesel fuel is needed, which yields better combustion and improved exergy efficiency. Highest exergy efficiency and lowest exergy destruction were observed at the CR of 19.5 that were 10.09% and 71.84% at the BMEP of 1.16 bar and 26.22% and 41.73% at the BMEP of 5.32 bar respectively. It was also found that the hot exhaust gases carry significant amount of input fuel exergy, especially at high loads, which indicates as a potential heat recovery option. The variations in exhaust gas exergy with engine loads for different CRs is shown in Fig. 6. It was found that at the high load, combustion temperature was significantly higher than that at the low load. Therefore, at the end of expansion process, burned gases leave at high temperatures leading to high exhaust gas exergy. Effect of CR on exhaust gas exergy was insignificant at low load. However, at high load, exhaust gas exergy was slightly decreased with higher CRs. As higher CRs improve the flame propagation speed and render shorter duration of combustion, these cause more of the combustion process to take place near TDC and lead to lower temperature of burned exhaust gases. On the other hand, increased CR also lead to higher peak temperature, 16

which may cause slight increase in exhaust gas temperature. Therefore, exhaust gas temperature is governed by the relative dominance of these phenomena. At the full load, exhaust gas exergies of 10.81%, 10.7%, 10.17% and 9.94% were calculated for CRs of 16.5, 17.5, 18.5 and 19.5 respectively. CR=16.5

CR=17.5

CR=18.5

CR=19.5

30

Energy efficiency (%)

25 20 15 10 5 0 1

2

3

4

5

6

BMEP (bar)

CR=16.5

CR=17.5

CR=18.5

CR=19.5

CR=16.5

CR=17.5

CR=18.5

CR=19.5

30

90

25

80 70

20

60

15

50

10

40

5

30

1

2

3

4

5

Exergy destroyed (% of input fuel exergy)

Exergy efficiency (%)

Fig. 4: Energy efficiency with engine loads for various compression ratios.

BMEP (bar)

Fig. 5: Exergy efficiency (solid lines) and exergy destruction (dotted lines) with engine loads for various compression ratios.

17

Exhaust gas exergy (% of input fuel exergy)

CR=16.5

CR=17.5

2

3

CR=18.5

CR=19.5

12 10 8 6 4 2 0 1

4

5

6

BMEP (bar)

Fig. 6: Exhaust gas exergy with engine loads for various compression ratios. The effect of engine load and CR on hydrocarbon emissions (HC) is shown in Fig. 7. Significant reduction in HC emissions were recorded at low load with higher CRs. At high loads, HC emissions were considerably lower than that at the low load, which further decreased with higher CRs. High HC emissions at low load is caused by incomplete combustion of fuel-air mixture as a result of over lean condition. In addition to that, as discussed above, small amount of pilot fuel may cause poor combustion. In case of biogas as the main fuel, the presence of CO2 significantly lowers the flame speed [11, 46]. The slowly propagating flame may terminate before reaching to the end-gas mixture, which could produce high HC emissions. The high ignition delay at low load has also found to contribute in high HC emissions [11]. The increase in compression ratio elevates the temperature in compression process that reduces the ignition delay period. Furthermore, high combustion temperature also helps in better combustion and reduces HC emissions. Increasing the CR from 16.5 to 19.5 showed 27.8% and 37.4% decrease in HC emissions at the BMEPs of 1.16 bar and 5.32 bar respectively. Similar effects as that of HC emissions were observed for CO emissions with increase in CRs. Fig. 8 shows the variation of CO emissions with engine load for various CRs. Significant reductions in CO emissions were observed with increased CRs both at low and high loads. By increasing the CR from 16.5 to 19.5, CO emissions were reduced by 38.2% and 35.6% at the BMEPs of 1.16 bar and 5.32 bar respectively. CO emissions are the result of partial oxidation of carbonaceous fuels. At fuel rich regions, insufficient availability of oxygen leads to freezing of oxidation reaction resulting in high levels of CO generation. The 18

further oxidation of CO depends on the reaction temperature that governs the affinity of carbon-oxygen combination. Increase in CR results in high in-cylinder pressure and temperature, which promote this reaction and further oxidation of CO to CO2. Perhaps, relatively low combustion temperature results in higher CO emissions at low load as compared to that with the high load. CR=16.5

CR=17.5

CR=18.5

CR=19.5

4

5

16

Hydrocarbon (g/kW h)

14 12 10 8 6 4 2 0 1

2

3

6

BMEP (bar)

Fig. 7: Hydrocarbon emissions (HC) with engine loads for various compression ratios. CR=16.5

CR=17.5

CR=18.5

CR=19.5

4

5

Carbon monoxides (g/kW h)

80 70 60 50 40 30 20 10

0 1

2

3

6

BMEP (bar)

Fig. 8: Carbon monoxide (CO) emissions with engine loads for various compression ratios. Another positive effect of increasing the CR with diesel-biogas DF engine is lowered smoke emissions. Variation in smoke emission with engine load is shown in Fig. 9 for various CRs. 19

It is evident that improved combustion with higher CRs help in reduction of smoke emissions. The high smoke generation takes place near diesel fuel spray zone, where fuel particles are clustered to form very small globules. Higher diesel substitution, especially at low load, results in smaller spray zone and reduces smoke emission as compared to high load. By increasing the CR from 16.5 to 19.5, smoke emissions were reduced by 52.5% and 38.3% at the BMEPs of 1.16 bar and 5.32 bar respectively. Variation of nitrogen oxides (NO x) emissions with engine load is shown in Fig. 10. It was found that NOx emissions increase with increase in both the engine load and CR. The incylinder NOx generation is governed by the chemical kinetics of interaction between oxygen and nitrogen molecules present in both the air and fuel [41]. The rate of chemical kinetics depends on reaction temperature, molecular concentrations and residence time. As the combustion temperature increases with the increase in CR, it also stimulates the rate of NO x formation. Therefore, CR of 19.5 resulted in highest NOx emissions throughout the engine load spectrum. The engine load showed great impact on NO x emissions as it increased from 3.6 g/kW h to 17.3 g/kW h, when the BMEP was increased from 1.16 bar to 5.32 bar. The major reason for low NOx emissions at low load is relatively lean fuel-air mixture as compared to that at the high load. Furthermore, high diesel substitution at low load leads to small amount of pilot fuel injection and consequently a small pilot spray zone. The combustion near this narrow spray zone creates relatively lower localized temperature and therefore, lower NOx emissions. On the other hand, at high load, high amount of fuel requirement results in a rich fuel-air mixture, which burns much rapidly and produces high combustion temperature. CR=16.5

CR=17.5

CR=18.5

CR=19.5

3

4

5

9

Smoke opacity (% )

8 7 6 5 4 3 2 1 0 1

2

BMEP (bar)

20

6

Fig. 9: Smoke emissions with engine loads for various compression ratios. CR=16.5

CR=17.5

CR=18.5

CR=19.5

4

5

20

Nitrogen oxides (g/kW h)

18 16 14 12 10 8 6 4 2 0 1

2

3

6

BMEP (bar)

Fig. 10: Nitrogen oxides (NOx) emissions with engine load for various compression ratios. 3.2. Effect of EGR The effects of using EGR on performance and emission characteristics of diesel-biogas DF engine have been illustrated from Fig. 11 to Fig. 14. It was found that at all the loading conditions, increasing the percentage of EGR renders lower diesel substitution. Variation in the diesel substitution with different EGR percentages is shown in Fig. 11. At small percentages of EGR of 5%, there was little variation in diesel substitution, especially at low loads. However, higher percentages of EGR led to sizable changes (decrease) in the diesel substitution. At low load (BMEP = 1.16 bar) and no EGR condition, the diesel substitution was 80.2% that was reduced to 79.6%, 76.0% and 70.0% with 5%, 10% and 15% EGR cases respectively. Similarly, at high load (BMEP = 5.32 bar), the diesel substitution was decreased from 67.3% to 64.5%, 63.0% and 55.0% respectively for above cases. This may be attributed to the well-known dilution effect of EGR, which increases the specific heat capacity of the incylinder charge mixture (air-biogas-EGR). In DF mode, a minimum amount of pilot fuel is needed to initiate the combustion in the premixed phase, which creates the favourable incylinder condition of sufficient temperature and pressure. The next stage of combustion is mainly sustained by propagation of flame through the premixed gas-air mixture. The increased heat capacity of charge mixture with EGR addition reduces the in-cylinder 21

temperature and thus reduces the flame propagation speed. This increases the chances of misfire and therefore, higher amount of diesel is required to sustain and improve the combustion characteristics. It is also important to mention that in the present configuration of DF combustion, EGR replaces some amount of air during induction and may affect the availability of oxygen for combustion. This effect is pronounced at the high load condition, where rich fuel-air ratio exists and may lead to poor combustion. Therefore, EGR should be limited to counter both the effects of reduced diesel substitution and poor combustion characteristics. The effect of EGR on energy efficiency of DF combustion is shown in Fig. 12. It was found that at low load, EGR showed slight improvements in the energy efficiency. The energy efficiencies of 10.91%, 11.58% and 11.08% were calculated with EGR percentages of 5%, 10% and 15% respectively as compared to 10.64% with no EGR case. This increase in energy efficiency of DF combustion may be attributed to the chemical effect of EGR [29]. The recirculated exhaust gas carries unburned fuel particles, which may contribute in combustion of the next cycles and improve the engine performance. Furthermore, by replacing some amount of air, EGR lowers the air-fuel ratio. The result of which could be improved combustion and better energy efficiency. This effect is particularly favourable at low load conditions, where very lean air-fuel mixture exists. However, at high load, high EGR percentages significantly dilute the charge and lead to poor combustion. At the full load, slight improvement in energy efficiency was observed with 5% EGR, whereas, further increase in EGR percentages reduced the engine performance. EGR=0%

EGR=5%

EGR=10%

EGR=15%

85

Diesel substitution (%)

80 75 70 65 60 55 50 45 40 1

2

3

4

5

6

BMEP (bar)

Fig. 11: Diesel substitution with engine loads for various EGR percentages.

22

EGR=0%

EGR=5%

EGR=10%

EGR=15%

30

Energy efficiency (%)

25 20

15 10 5 0 1

2

3

4

5

6

BMEP (bar)

Fig. 12: Energy efficiency with engine loads for various EGR percentages. Table 4: Combustion parameters for dual fuel engine operating at different EGR percentages. Test

Peak cylinder

Peak heat release

Ignition delay

conditions

pressure (bar)

rate (J/°CA)

(°CA)

Low Load

High load

Low Load

High load

Low Load

High load

EGR–0%

66.82

81.47

21.58

94.34

21.4

13.2

EGR–5%

67.82

81.80

19.42

64.61

21.0

13.4

EGR–10%

66.85

78.02

17.32

35.58

21.6

13.8

EGR–15%

65.92

74.51

16.69

28.74

22.3

14.5

The exergy efficiency and exergy destruction are shown in Fig. 13 for different EGR percentages. The high percentage of EGR negatively affects the exergy efficiency, especially at high loads. The addition of EGR leads to increased ignition delay and lowered in-cylinder temperature as shown in Table 4. This affects the pre-mixed phase of combustion, which is the primary reason for poor combustion. Moreover, EGR lowers the burn rate in diffusion phase of combustion by absorbing some amount of heat. At high EGR percentages, these effects dominate over the positive chemical effect of EGR and reduce the exergy efficiency. Simultaneously, it also leads to higher exergy destruction and the highest exergy destruction is found at 15% EGR both at low and high loads. At the full load, exergy destructions were increased by 3.77%, 7.7% and 10.96% with 5%, 10% and 15% EGRs respectively as compared to that with the no EGR case. 23

The exhaust gas temperature and exhaust gas exergy were also significantly affected by the EGR addition. The effect of EGR on the exhaust gas exergy is shown in Fig. 14 with the engine loads. It can be seen that as the EGR percentage increases, the exhaust gas exergy decreases with the greater effect of EGR at higher loads as compared to lower loads. As discussed in the preceding sections, the lower combustion temperature achieved with EGR causes lower charge temperature after the expansion stroke. Therefore, exhaust gases leave at lower temperature relating to the lower exhaust gas exergies. At the BMEP of 1.16 bar, exhaust gas exergies were decreased by 0.44%, 0.74% and 1.04% with 5%, 10% and 15% EGRs respectively as compared to the no EGR case. Similarly, at the BMEP of 5.32 bar, 0.65%, 2.03% and 2.99% decrease in exhaust gas exergies were calculated as compared to

EGR=0%

EGR=5%

EGR=10%

EGR=15%

EGR=0%

EGR=5%

EGR=10%

EGR=15%

30

90

25

80 70

20

60 15

50

10

40

5

30 1

2

3

4

5

6

Exergy destroyed (% of input fuel exergy)

Exergy efficiency (%)

the no EGR case.

BMEP (bar)

Fig. 13: Exergy efficiency (solid lines) and exergy destruction (dashed lines) with engine loads for various EGR percentages.

24

Exhaust gas exergy (% of input fuel exergy)

EGR=0%

EGR=5%

EGR=10%

EGR=15%

12 10 8 6 4 2 0 1

2

3

4

5

6

BMEP (bar)

Fig. 14: Exhaust gas exergy with engine loads for various EGR percentages. On the emissions side, the effects of using EGR in diesel-biogas DF engine are shown in Fig. 15 to Fig. 18. It was also found that some reduction in HC emissions is possible with EGR in diesel-biogas DF engine at the low load. The effect of EGR on HC emissions is shown in Fig. 15 with the engine loads. At the BMEP of 1.16 bar, HC emissions were reduced by 11.45%, 17.7% and 14.58% using EGR percentages of 5%, 10% and 15% respectively. Perhaps, with the use of EGR, some of the unburned hydrocarbons present in the exhaust gases from the previous cycles could burn in the next cycle. The favourable condition of excess air is present at low load conditions for proper combustion of any residual hydrocarbons along with the main charge. On the other hand, availability of air is less at the high loads, which further decreases with higher percentages of EGR. This possess the possibility of increased HC emissions at higher EGR percentages as can be seen from Fig. 15.

25

EGR=0%

EGR=5%

EGR=10%

EGR=15%

12

Hydrocarbon (g/kW h)

10 8 6 4 2 0 1

2

3

4

5

6

BMEP (bar)

Fig. 15: Hydrocarbon emissions (HC) with engine loads for various EGR percentages. Variations in CO emission with engine loads with various EGR percentages are shown in Fig. 16. The effect of EGR on CO emission of DF engine was similar to that on HC emissions. It reveals that the use of EGR reduces CO emissions at low loads, whereas, little variation was observed at high loads. Furthermore, the reduction in CO emission decreases with increase in EGR percentage. At the BMEP of 1.16 bar, CO emissions were reduced by 22.8%, 18.6% and 13.5% using 5%, 10% and 15% EGRs respectively. The EGR can recycle and burn the CO contain in the exhaust and may help in reducing it. Furthermore, the active free radicals present in the exhaust gases promote the oxidation process and help in complete combustion. However, it appears that inert nature of EGR has the dominating effect and therefore, higher percentages of EGR acts negatively and increases the emissions. The rate of formation of CO inside the engine cylinder depends on air-fuel ratio, combustion temperature and mixture formation. At high load, addition of EGR causes fuel richness (due to replacement of air), which promotes the CO formation. Furthermore, its inert nature renders lower combustion temperature, which dampen the further oxidation process leading to higher CO emissions.

26

EGR=0%

EGR=5%

EGR=10%

EGR=15%

50

Carbon monoxides (g/kW h)

45 40 35 30 25 20 15 10

5 0 1

2

3

4

5

6

BMEP (bar)

Fig. 16: Carbon monoxide (CO) emissions with engine loads for various EGR percentages. Variations in smoke emission with engine loads are shown in Fig. 17 for various EGR percentages. It reveals that EGR has negative impact over smoke emissions for all the loading conditions. Also, smoke emission drastically increases with higher percentages of EGRs. The addition of EGR results in both lower in-cylinder temperature and reduced oxygen availability, which reduce the smoke oxidation mechanism [41]. Furthermore, higher percentage of EGR results in lower diesel substitution (see Fig. 11) and therefore higher liquid fuel also causes increased smoke emissions. It is also evident that smoke emission increases with the increase in engine loads. Variation in NOx emissions is depicted in Fig. 18, which shows significant improvement over no EGR case. The use of EGR showed monotonic reduction in NOx emissions at both low and high loads. At the BMEP of 1.16 bar, reduction in NO x emissions were calculated to be 11.1%, 27.8% and 52.8% against the no EGR case. Whereas, at the BMEP of 5.32 bar, the reductions were 3.4%, 14.7% and 22.9% respectively for the above cases. It shows that higher reductions in NOx emissions are possible at low load as compared to that with the high load. Nevertheless, reductions are significant at all the loading conditions. The effect of EGR is to reduce the combustion temperature, which brings down the rate of NOx formation. It may also be attributed to slight reduction in oxygen availability with EGR addition, which could also reduce the rate of NOx formation. It can be noted that using the smallest proportion 27

of 5% EGR, it is possible to compensate the increased NO x emissions by employing the higher CR without compromising with the engine efficiency. EGR=0%

EGR=5%

EGR=10%

EGR=15%

25

Smoke opacity (% )

20

15

10

5

0 1

2

3

4

5

6

BMEP (bar)

Fig. 17: Smoke emissions with engine loads for various EGR percentages.

Nitrogen oxides (g/kW h)

EGR=0%

EGR=5%

EGR=10%

EGR=15%

20 18 16 14 12 10 8 6 4 2 0 1

2

3

4

5

6

BMEP (bar)

Fig. 18: Nitrogen oxide emissions (NOx) with engine loads for various EGR percentages. 3.3. Effect of EGR temperature The effects of EGR temperature on the performance characteristics of diesel-biogas DF engine are presented from Fig. 19 to Fig. 22. The hot EGR affected the diesel substitution, however, in opposite manners at low and high loads. At low load, diesel substitution was increased with hot EGR, whereas at high load, it was slightly decreased as compared to that 28

with the cold EGR. The hot EGR rises the intake charge temperature causing some reductions in ignition delay, which may help in slight reduction of pilot diesel at low load. On the other hand, hot EGR also decreases the volumetric efficiency and therefore, the air intake capacity of the engine. This effect could dominate at the high load condition due to rich fuel-air mixture and may restrict the amount of gaseous fuel utilization. The effect of increasing EGR percentage was found similar for both hot and cold EGRs that was to decrease the diesel

High load (Hot EGR)

High load (Cold EGR)

53.55

60 55

55.03

79.04

73.65 63.02

65

62.13

70

64.50

75 64.20

Diesel substitution (%)

80

70.06

Low load (Cold EGR)

76.05

Low load (Hot EGR) 79.58

85

81.14

substitution.

50 45 40

5% EGR

10% EGR

15% EGR

Fig. 19: Diesel substitution at low (BMEP = 1.16 bar) and high (BMEP = 5.32 bar) loads with cold and hot EGR. The energy and exergy efficiencies are depicted in Fig. 20 and Fig. 21 respectively for both hot and cold EGRs and different EGR percentages. It shows that hot EGR slightly improves the energy and exergy efficiencies at low load condition as compared to cold EGR. Due to hot EGR, the charge temperature is high after compression that causes lower ignition delay, higher combustion temperature and improved flame velocity. These factors lead to improved combustion of the gas-air charge and consequently, slight improvements in the efficiencies are observed with hot EGR. The improvement in effective air-fuel ratio and active free radicals from hot exhaust gases may also help in better combustion at low load conditions. On the other hand, induction of hot EGR also replaces significant amount of intake air causing reduced volumetric efficiency. This condition is critical especially with high percentages of EGR and high engine loads. Such conditions could lead to the incomplete combustion of fuel-air charge and adversely affect the combustion process. Perhaps, this could be reasons for slightly lower efficiencies at the high load with high percentages of hot 29

EGR. The highest energy and exergy efficiencies were observed with 10% of hot EGR at the low load and 5% of hot EGR at the high load. Nevertheless, it was found that the use of hot EGR reduces the exergy destruction at both low and high loads as shown in Fig. 22. It reflects that hot EGR lowers the irreversibility in various engine processes and consequently

High load (Cold EGR) 25.14

26.84

27.06

High load (Hot EGR) 27.99

Low load (Cold EGR)

28.30

30 25

11.08

11.24

11.58

10.91

15

11.97

20 11.23

Energy efficiency (%)

Low load (Hot EGR)

25.65

lead to better utilization of the energy from the second-law of thermodynamics point of view.

10 5 0

5% EGR

10% EGR

15% EGR

Fig. 20: Energy efficiency at low (BMEP = 1.16 bar) and high (BMEP = 5.32 bar) loads

High load (Cold EGR) 23.78

25.40

25.61

High load (Hot EGR) 26.5

Low load (Cold EGR)

26.79

30 25

10.50

10.65

11.0

11.35

15

10.3

20 10.65

Exergy efficiency (%)

Low load (Hot EGR)

24.27

with cold and hot EGR.

10 5 0

5% EGR

10% EGR

15% EGR

Fig. 21: Exergy efficiency at low (BMEP = 1.16 bar) and high (BMEP = 5.32 bar) loads with cold and hot EGR.

30

52.69

52.04

49.43

48.83

50

45.50

60

44.12

70

73.97

High load (Cold EGR)

72.65

High load (Hot EGR)

72.86

Low load (Cold EGR)

71.28

73.30

72.13

Exergy destroyed (% of input fuel exergy)

80

Low load (Hot EGR)

40 30 20 10 0

5% EGR

10% EGR

15% EGR

Fig. 22: Exergy destroyed at low (BMEP = 1.16 bar) and high (BMEP = 5.32 bar) loads with cold and hot EGR. On the emissions side, hot EGR showed reductions in both HC and CO emissions. Variations in HC and CO emissions at low and high loads with hot and cold EGRs are shown in Fig. 23 and Fig. 24 respectively. The results reflected that significant improvement was possible at the low load, however, little to no variations was observed at the high load. It indicates that hot EGR improves the DF combustion at the low load. Furthermore, higher level of improvements were observed with lower percentages of EGR. In fact, 15% EGR case showed higher CO emissions both at low and high loads. Therefore, higher percentages of EGRs may deteriorate the combustion process specifically at high loads. Variations in smoke emission at low and high loads with hot and cold EGRs are shown in Fig. 25. It shows increased smoke emissions with hot EGR as compared to cold EGR. Furthermore, the increments are higher at the high load as compared to low load. One of the major reasons in this regard could be the reduction in diesel substitution with hot EGR (see Fig. 19). The increase in liquid fuel injection may enhance the soot formation, which further increases with the presence of higher percentages of inert EGR gases. The improved combustion temperature should favour the oxidation of smoke particles, however, this effect shows little impact over the others. The use of hot EGR also showed slight increase in NO x emissions as compared to that with the cold EGR at both low and high loads. NOx emissions are shown in Fig. 26 for hot and cold EGRs. It could be related to slight increase in the combustion temperature. However, the increments are not significant. Perhaps, the reduced intake air and combustion duration slow down the NOx formation. 31

High load (Cold EGR) 7.92

7 6 5 4

0.94

0.97

0.68

1

0.61

2

0.69

3 0.62

Hydrocarbon (g/kW h)

8

8.20

High load (Hot EGR)

7.99

Low load (Cold EGR)

7.25

8.54

8.23

9

Low load (Hot EGR)

0

5% EGR

10% EGR

15% EGR

Fig. 23: Hydrocarbon emissions (HC) at low (BMEP = 1.16 bar) and high (BMEP = 5.32

Low load (Cold EGR)

High load (Hot EGR)

High load (Cold EGR)

35

34.88

34.43

33.06

40

29.67

17.43

18.23

16.12

20

15.68

25

16.46

30

16.05

Carbon monoxides (g/kW h)

45

39.48

Low load (Hot EGR)

37.12

bar) loads with cold and hot EGR.

15 10 5 0

5% EGR

10% EGR

15% EGR

Fig. 24: Carbon monoxide (CO) emissions at low (BMEP = 1.16 bar) and high (BMEP = 5.32 bar) loads with cold and hot EGR.

32

High load (Hot EGR)

High load (Cold EGR)

5

13.6

13.9

6.1

6.3

2.9

10

7.1

7.6

15

11.3

11.9

20

2.8

Smoke opacity (% )

25

22.7

Low load (Cold EGR) 23.5

Low load (Hot EGR)

0

5% EGR

10% EGR

15% EGR

Fig. 25: Smoke emissions at low (BMEP = 1.16 bar) and high (BMEP = 5.32 bar) loads

High load (Cold EGR)

13.68

15.17

16

14.90

High load (Hot EGR) 16.78

Low load (Cold EGR)

16.65

18 14 12 10

2

1.73

1.80

2.58

4

2.97

6

3.21

8 3.51

Nitrogen oxides (g/kW h)

Low load (Hot EGR)

13.40

with cold and hot EGR.

0

5% EGR

10% EGR

15% EGR

Fig. 26: Nitrogen oxide (NOx) emissions at low (BMEP = 1.16 bar) and high (BMEP = 5.32 bar) loads with cold and hot EGR. 3.4. Exergy balance The exergy balances are shown for both low and high load conditions in Fig. 27 and Fig. 28 respectively. These are at the highest exergy efficiency points corresponding to 10% hot EGR at low load and 5% hot EGR at high load. It can be seen that at the low load, the percentage of input fuel exergy converted to shaft work is much smaller as compared to high load. At the low load, small amount of pilot diesel fuel generates a weak ignition source, and the combustion proceed slowly through lean gas-air mixture. As a result, the heat release rate is 33

much smaller at the low load operation that produces relatively low temperature and poor combustion of DF engine. This also leads to relatively lower heat transfer and exhaust gas exergies at the low load as compared to that at the high load. Therefore, the low load operation of DF engine shows significantly higher exergy destruction. The exergy destruction at low load was 71.28% that was decreased to 44.13% at high load condition. It can be attributed to relatively high pressure and temperature conditions achieved at the high load, which reflects higher quality of energy and therefore its better conversion to work and other engine processes. Among the various sources of exergy destructions, combustion accounts for the highest share, which has also been reported by the other researchers [42, 43, 47]. The combustion is accompanied by thermodynamic irreversibility due to its various sub-processes such as internal heat transfer, mass transfer and chemical reactions. The transport processes of heat and mass transfer occur due to temperature and concentration gradients through irreversible processes and cause entropy generation [47]. It appears from Fig. 27 and Fig. 28 that the combustion irreversibility at low load is slightly lower than the high load condition. However, the combustion irreversibility is decreased with the high temperature of operation or the high temperature of combustion products [42]. The lower combustion irreversibility at low load can be attributed to relatively poor combustion process and therefore the incomplete combustion products. Mechanical friction is another source of irreversibility, which accounts for 11.80% and 6.70% of the total input fuel exergy at low and high loads respectively as shown in Fig. 27 and Fig. 28. The mechanical friction in IC engines is a predominant function of the engine speed and also slightly affected by the engine load. In the present work, as the engine speed remains constant, the magnitude of exergy loss towards mechanical friction is not much affected with the engine loads. However, as a result of higher input exergy at high load, lower percentage of exergy loss with mechanical friction is obtained as compared to low load. Similarly, the magnitude of mixing irreversibility was found to vary insignificantly with engine loads, however, their percentage distributions vary due to differences in input fuel exergies. Furthermore, it was found that mixing irreversibility accounts for small part of the total exergy destruction as compared to other processes.

34

Fig. 27: Sankey diagram showing exergy distribution at low load.

Fig. 28: Sankey diagram showing exergy distribution at high load.

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4. Conclusions In the present article a diesel-biogas DF engine was studied to investigate the effects of variation in CR and addition of EGR on performance and emission characteristics of the engine. The CR was varied from 16.5 to 19.5 in stepwise manner and engine operations were studied at maximum pilot fuel substitution settings. The engine operations were also studied at three EGR percentages of 5%, 10% and 15%. In addition to that the effect of EGR temperature was also investigated on engine performance and emission characteristics. Following conclusion are drawn based on the above results: 1. It was revealed that employment of higher CR improves the pilot fuel substitution and energy/exergy efficiency. The exergy efficiency was increased by 2.94% and exergy destruction was decreased by 4.66% at full load condition as the CR was increased from 16.5 to 19.5. On the emissions side, increase in CR showed significant decrease in HC, CO and smoke emissions, however, NOx emissions were increased with higher CRs. 2. It was found that use of EGR in diesel-biogas DF engine can control the problem of higher NOx emissions at higher CR. The NO x emissions were reduced by 3.4%, 14.7% and 22.9% with the use of 5%, 10% and 15% cold EGRs respectively at the full load condition. 3. The cold EGR also indicated slight improvements in engine performance at the low load condition. Nevertheless, with high EGR percentages and high load, slight deterioration in engine efficiency was noticed. 4. This study also reveals that hot EGR has positive effects over cold EGR on the engine performance and emission characteristics of diesel-biogas DF engine. 5. The highest exergy efficiencies were obtained with hot EGR cases. They were 11.35% with 10% hot EGR at low load, and 26.79% with 5% hot EGR at high load at the CR of 19.5. At these operating points with hot EGR, slight improvements in HC and CO emissions were also recorded. 6. Based on the results obtained from this study, a combined approach of higher CR and hot EGR shows a promising technique for simultaneous performance enhancement and emission reduction from diesel-biogas DF engines.

36

Acknowledgement We would like to acknowledge the support of Council of Scientific & Industrial Research (CSIR), New Delhi, India. Research facility provided by Indian Institute of Technology Delhi (IITD) is also gratefully acknowledged.

Abbreviations AFR BDC BMEP BP CH4 CI CO CR DF DS EGR HC LHV TDC Symbols Cp E I

 m

n P Q R R

S s T X

x y

Subscripts I II BG ch comb D DF

Air-to-fuel ratio Bottom dead center Brake mean effective pressure Brake power Methane/Natural gas Compression ignition Carbon monoxide Compression ratio Dual fuel Diesel substitution Exhaust gas recirculation Hydrocarbon Lower heating value Top dead center Specific heat at constant pressure, J/Kg K Energy, J Irreversibility, J/s Mass flow rate, Kg/s Number of moles or number of species Absolute pressure, Pa Heat, J characteristic gas constant, J/Kg K Universal gas constant J/mole K Entropy, J/K Specific entropy, J/Kg K Absolute temperature, K Exergy, J Specific exergy, J/Kg Mole fraction

First-law Second-law Biogas Chemical Combustion Diesel fuel Dual fuel 37

ex fri g gen in i, j mix p Q r W un Superscripts 0 Greek Symbols 

Exhaust Friction In-cylinder gas Generation incoming Arbitrary number/species Mixture products Heat transfer Reactants Work Unaccounted Reference state Efficiency

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Highlights 1. 2. 3. 4. 5.

Performance and emission features of diesel-biogas dual fuel engine are studied. Effects of compression ratio, EGR and EGR temperature are investigated. Exergy analyses is applied and various sources of irreversibilities are quantified. Higher compression ratio improves the diesel substitution and efficiency. Rate of EGR and EGR temperature also affects the engine characteristics.

43