Template for the Preparation of Abstract

8th International Exergy, Energy and Environment Symposium

Second Law Analysis of a CI Engine Fuelled With Biodiesel-Diesel Blends 1*

1

2

Abdulvahap ÇAKMAK, Atilla BILGIN

Ondokuz Mayıs University, Kavak Vocational School, Department of Motor Vehicles and Transportation Technologies Samsun, 55850, TURKEY 2

Karadeniz Technical University, Faculty of Engineering, Mechanical Engineering Department Trabzon, 61080, TURKEY * E-mail: [email protected]

Keywords: Energy analysis, exergy analysis, exergy destruction, biodiesel-diesel blends

Abstract In this study, first law analysis (energy analysis) and second law analysis (exergy analysis) were applied to a single cylinder, air cooled and direct injection diesel engine using biodiesel-diesel blends. The fuel blends contain corn oil methyl ester by volume ratios of 10, 20 and 50 %. By using experimental data which were collected at different engine speeds for full load condition, energy and exergy components of the engine were calculated and compared with each other. According to results the tested engine can run more efficiently, by using B10 for full load condition at 2000 rpm where maximum exergetic efficiency occurs. I. Introduction Today, diesel engines are commonly used in many fields due to their higher efficiency which reaches up to 40%. This situation has lead to an increase in the share energy consumption of diesel engines. It is estimated that increasing energy demand can not be supplied only by fossil fuels. In addition, environmental problems resulting from the use of fossil-based fuels has necessitated use of renewable and environmentally friendly fuels. In fact, there are many alternative fuels which can be substituted for fossil-based diesel fuel (Tat, 2011). However, biodiesel is the best alternative to the fossil-based diesel fuel in terms of fuel properties, engine performance and emisssions. It is widely accepted that biodiesel can reduce HC, CO and PM emissions but slight performance loss in diesel engines (Ozsezen and Canakcı, 2010). Therefore, the use of pure biodiesel or biodiesel-diesel fuel blends in diesel engines increases in all over the world. This kind of development makes attractive to study on using biodiesel-diesel fuel blends in CI engines. Former studies on using biodiesel as an alternative fuel in diesel engine are extensively based on the first law of thermodynamics. Although the first law analysis is sufficient for general performance computations, the insight provided by a second law analysis is invaluable in understanding the details of the overall thermodynamics of engine operation (Caton, 2000). Furthermore, without second law analysis, it is not possible to determine the energy degradation that causes a reduction in work output of the engine (Alkidas, 1988). The second law of thermodynamics provides a distinction between the quantity and quality of energy by taking into account the irreversibilities (Moran and Shapiro, 1988). Therefore, the application of exergy analysis to internal combustion engines is very useful to provide quantitative information on irreversibilities and various exergy losses, and in recent decades, it has been widely applied to internal combustion engines (Sezer and Bilgin, 2013).

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In literature, there are a lot of papers which apply exergy analysis to internal combustion engines. For application of exergy analysis to internal combustion engines, some researchers used a thermodynamics-based engine cycle model with incorporate second law analysis and the others used experimental data. Zheng and Caton employed an engine cycle simulation incorporating the second law of thermodynamics to evaluate the energy and exergy distribution of various processes associated with injection timings and EGR levels in a low temperature combustion diesel engine (Zheng and Caton, 2012). Sekmen and Yilbası employed the first and second law of thermodynamics to analyze the quantity and quality of energy in a four cylinder, direct injection, water cooled diesel engine operated with diesel fuel and soybean oil methyl ester as biodiesel fuel at constant engine speed (Sekmen and Yilbası, 2011). Azoumah et al. performed an exergy analysis combined with gas emissions analysis to optimize the performance of a compression ignition engine using biofuels such as cottonseed and palm oils, pure or blended with diesel for different engine loads (Azoumah et al., 2009). Rath et al. applied the energy and exergy analysis on a single cylinder four stroke diesel engine using karanja methyl ester under varying compression ratios (Rath et al., 2014). Tat determined the effect of cetane number and ignition delay on the energy and exergy efficiencies of an internal combustion engine fuelled with biodiesel (Tat, 2011). This paper aims to contribute in this field by assessing diesel engine running on biodiesel-diesel fuel blends at full load and various engine speeds. For this purpose, fuel energy and exergy, effective power, exhaust heat and exhaust exergy, total heat loss and total heat loss exergy, exergy destruction, and first and second law efficiencies were calculated and the results were compared for all test fuels and operation conditions.


Nomenclature 𝐸𝑓̇ 𝑚̇𝑓 𝐻𝑢 Ne

M ω 𝑛 𝑄̇𝑒𝑥ℎ 𝑚̇𝑤 𝑄̇𝑙𝑜𝑠𝑠 𝐸̇ 𝑥𝑓 𝑒𝑥𝑓𝑐ℎ 𝐸̇ 𝑥𝑒𝑥ℎ 𝐸̇ 𝑥𝑑𝑒𝑠𝑡 0 𝑒̅𝑥𝑐ℎ ℎ̅ 𝑠̅ w1 , w2 , w3 , … . , wn

all the relevant temperatures and pressures. Engine tests were performed at 1000 and 3000 rpm, with intervals of 500 rpm at full load. Test data were recorded by versatile data acquisition system with 0.5 second interval during 10 seconds and the mean values of the data were used for energy and exergy analysis. All measurements are conducted under steady state conditions.

: Fuel energy enters the cylinder per unit time (kW) : Mass flow rate of the fuel (kg/s) : Lower heating value (kJ/kg) : Effective power (kW) : Torque (Nm) : Angular velocity (rad/s) : Engine speed (rpm) : Exhaust heat per unit time (kW) : Mass flow rate of calorimeter cooling water (kg/s) : Total heat loss per unit time (kW) : Fuel exergy enters the cylinder per unit time (kW) : Specific chemical exergy of the fuel (kJ/kmol) : Exhaust exergy per unit time (kW) : Exergy destruction rate (kW) : Standard chemical exergy (kJ/kmol) :Specific molar enthalpy (kJ/kmol) : Specific entropy (kJ/kmol/K)

𝑥1 , 𝑥2 , 𝑥3 , . . . , 𝑥𝑛

: Uncertainties of independent variables : Independent variables

Greek letters ɳI ɳII

: First law efficiency : Second law efficiency

Subscript 0 exh dest ch tm f i

: Reference environment : Exhaust : Destruction : Chemical : Thermomechanical : Fuel : Component

Abbreviations LHV CI rpm

: Lower heating value : Compression ignition : Revolutions per minute

Fig. 1: Schematic view of the experimental setup

Biodiesel (corn oil methyl ester) was produced in the laboratory from corn oil by transesterification reaction with methanol (CH3OH), where potassium hydroxide (KOH) was used as a catalyst. The reaction parameters giving the lowest kinematic viscosity were chosen as 1.1% catalyst concentration, 60℃ reaction temperature, 60 minutes reaction time and 9:1 alcohol/oil molar ratio (Gülüm and Bilgin, 2015). Diesel fuel (Eurodiesel) was purchased from a commercial supplier with its typical formula C14.09H24.78 (Canakci, 2007). Produced corn oil biodiesel was then mixed with diesel fuel by ratios of 10, 20 and 50% by volume designated as B10, B20 and B50, respectively. Some important properties of the test fuels such as density (ISO 4787), kinematic viscosity (DIN 53015) and lower heating value (DIN 51900-2) were measured in the Prof. Dr. Saadettin GUNER Fuel Research and Application Center at Karadeniz Technical University. Average molecular mass and typical formula of the corn oil biodiesel calculated from fatty acid distribution. After that the typical formula of B10, B20 and B50 was determined as C14.55H25.82O0.2, C15H26.85O0.4 and C16.38H29.96O1, respectively. Some fuel properties of the test fuels are shown in Table 1.

II. Experimental Methods Experiments were carried out on existing test unit at the Mechanical Engineering Department of Karadeniz Technical University produced by Tecquipment. A schematic view of the test unit is shown in Fig. 1. The test unit includes a single cylinder, natural aspirated, four-stroke, air cooled and direct injection diesel engine. Maximum power of the test engine is 6.6 kW (@3600 rpm) and its compression ratio is 20.5:1. The experimental setup enables accurate measurement of engine torque, engine speed, air mass flow rate, fuel volume flow rate and exhaust gas calorimeter cooling water volume flow rate and -2-

Tab. 1: Some fuel properties of test fuels Properties Diesel B10 B20 B50 Density at 826.379 831.767 836.357 852.322 20 °C, kg/m3 Kinematic 2.8475 2.9326 2.9781 3.2342 viscosity at 40°C, mm2 /s Average 194.213 203.987 213.629 242.939 molecular mass, kg/kmol Lower heating 42797 42224 41697 40048 value, kJ/kg Lower heating 35367 35119 34874 34134 value, kJ/L


III. Energy and exergy analysis

by heat transfer. Because friction power is dissipated between the piston assemblies (about half of the total friction), valve mechanism, bearings and engine driven accessories, and it ends up heating the oil or surrounding environment (Pulkrabek, 1997), friction losses evaluated in the total heat loss (Çakmak, 2014). Brake thermal efficiency (first law efficiency) of the control volume is usually defined as the ratio of effective power output to the fuel energy input rate and determined by Eq. (6).

For energy and exergy analysis the following assumptions were made; the entire engine was considered as a steady-state control volume, fuel and air enter, and effective power, heat loss and exhaust gases leave the control volume at a constant rate as depicted in Fig. 2. The combustion air and exhaust gases are assumed to be ideal gas mixtures and potential and kinetic energy effects of the incoming and outgoing fluid streams are ignored to simplify calculations. Complete combustion is also considered. The reference environment is defined as P0 = 1 atm and N ɳI = ̇ e 𝑇0 = 25 ℃. Ef

(6)

Input exergy rate, including only chemical exergy of the fuel calculated as Ėxf = ṁf exfch

(7)

Where exfch denotes specific chemical exergy of the fuel and it evaluated using the following expression (Kotaş, 1995): h o s exfch = [1.0401 + 0.1728 + 0.0432 + 0.2169 (1 − c

h

2.0628 )] Hu c

c

c

(8)

Fig. 2: Engine as a thermodynamic open system

Because the incoming air stream is very close to reference environment state the amount of the energy brought into control volume with it can be ignored. Then the only energy input to the control volume is fuel energy is given by Eq. (1). Ėf = ṁf Hu

In Eq. (8) c, h, o and s are mass fractions of carbon, hydrogen, oxygen and sulfur content of the fuel, respectively. Effective power exergy, meaning the net exergy work rate, is also equal to the effective power: ĖxNe = Ne

(1)

(9)

Exhaust exergy rate is the sum of the thermomechanical Where Hu is the lower heating value (kJ/kg) and ṁf is and chemical exergy of each component of the the mass flow rate of the fuel(kg/s), respectively. The exhausted gases in a unit time and it was calculated effective power of the engine Ne (kW) is determined by using the following equation. Eq. (2). ṁ Ėxexh = ∑ni=1 i (e̅xtm,i + e̅xch,i ) (10) μi Ne = Mω10−3 (2) When thermomechanical and chemical exergy terms are Where ω = πn⁄30 is angular velocity of the crankshaft, written in open forms Eq. (10) becomes: and M and n are torque (Nm) and engine speed (rpm), ṁ respectively. Energy loss rate due to exhaust gas can be ̇ Exexh = ∑ni=1 i {[h̅i (T) − h̅i (T0 )] − T0 [s̅ i (T) − s̅ i (T0 )] + μi calculated as follows: 0 ̅xi lnxi )} (xi e̅xch,i +R (11) (Tw2 −Tw1 ) (Texh2 − T0 ) Q̇ exh = ṁw cp,w (T (3) exh2 −Texh3 ) Where ṁi is the mass flow rate of the i-th component (kg/s), μi is the molecular mass of the i-th component Where ṁw (kg/s) and cp,w (kJ/kg/K) mass flow rate (kg/kmol) , h̅i (T) is the specific molar enthalpy and specific heat of the calorimeter cooling water, Tw1 and (kJ/kmol) at the exhaust temperature T(K) , h̅ (T ) is i 0 Tw2 cooling water inlet and outlet temperatures, Texh2 and the specific molar enthalpy (kJ/kmol) at the ambient Texh3 exhaust gas temperatures at inlet and outlet of the temperature T (K) , s̅ (T) is the specific entropy 0 i calorimeter, and T0 ambient air temperature. Total heat (kJ/kmol/K) at the exhaust temperature T(K), s̅ (T ) is i 0 loss per unit time for the control volume was determined by the absolute entropy (kJ/kmol/K) at the ambient applying energy conservation principle to control volume. temperature T0 (K) , xi is the molar ratio of the i-th 0 component in the exhaust gas, e̅xch,i is the standard ̅ Ėin = Ėout (4) chemical exergy of the i-th component (kJ/kmol) and R is the universal gas constant (kJ/kmol/K). Q̇ loss = Ėf − Ne − Q̇ exh (5) Total heat loss exergy rate was calculated using Eq. (12). Total heat loss comprises all energy lost to the environment ĖxQ̇loss = (1 − T0 ⁄Te )Q̇ loss -3-


(12)

neat diesel fuel, while B20 and B50 blend give lower power output, although the differences are slight. Oxygen Where T0 and Te are ambient air temperature and engine content in molecular structure of the biodiesel could be block temperature, respectively. Units are in K for the reason for the improvement in combustion giving a temperatures in Eq. (12). slight increase in power output for B10 blend. However, Under steady-state operation exergy rate balance equation because of the LHV of blend fuels are lower than neat for the control volume can be expressed as: diesel fuel and high viscosity of biodiesel, an increase in biodiesel content in the blend results a decrease in power Ėxf − ĖxNe − Ėxexh − ĖxQ̇loss − Ėxdest = 0 (13) output of the engine. Exergy destruction rate Ėxdest can be solved from this equation to yield: Ėxdest = Ė xf − ĖxNe − Ė xexh − ĖxQ̇loss

(14)

Finally, the second law (or exergetic) efficiency of the engine can be evaluated from the ratio of the power output to the fuel exergy input rate, i.e., ɳII =

Ė xNe

(15)

Ėxf

The experimental data are obtained from measured physical quantities. These quantities have some Fig. 3: Variations of fuel energy input rates versus engine speed uncertainties due to uncertainties of measuring tools and measurement systems. Hence, uncertainty analysis should be applied for demonstration reliability of the calculated results. Uncertainties of the calculated results were determined by using by the method proposed by Kline and McClintock given in (Holman, 2001). wR = [(

∂𝑅 ∂x1

2

w1 ) + (

∂𝑅 ∂x2

2

w2 ) + ⋯ + (

∂𝑅 ∂xn

2 0.5

wn ) ]

(16)

where x1 , x2 ,…,xn are independent variables, w1 , w2 ,…, wn are the uncertainties of each independent variables, and wR is the uncertainty of the result R . It was determined that the uncertainty of the first and second law Fig. 4: Variations of fuel exergy input rates versus engine speed efficiencies varied from 0.6090% to 0.8040% for all the test fuels and engine speeds Therefore, it can be said that the results have high reliability. IV. Result and discussions Fig. 3 and Fig. 4 show the variation of fuel energy and exergy input rates with the change in engine speed and test fuels. Both fuel energy rate and fuel exergy rate increase with engine speed as expected. Fuel energy and fuel exergy are proportional to lower heating value of the fuel. As seen in figures biodiesel-diesel blends provided slightly less energy and exergy to the engine than diesel fuel. Slight differences between tested fuels can be attributed to the smaller differences between volume-based lower heating values (LHVs) of blend fuels and diesel fuel in compared to mass-based heating values of them. Although mass-based LHV of diesel fuel is 1.36%, 2.64% and 6.86% higher than B10, B20 and B50, respectively, its volume-based heating value is only 0.71%, 1.41% and 3.61% higher than those of blends. Effective power variations of the engine are given in Fig. 5 for tested fuels at full load conditions. As seen in the figure the shapes of the power-speed curve are similar to each other. B10 blend gives slightly higher power output than the -4-

Fig. 5: Variations of effective power versus engine speed

Fig. 6 shows variations of the exhaust heat loss rate and exhaust exergy rate with respect to engine speed. Both of them increase with increasing engine speed due to increased fuel energy and fuel exergy entering the cylinder. Variations of the exhaust heat losses for the tested fuels are originated from differences of the exhaust temperatures. It was observed that exhaust temperatures of the blend fuels do not change linearly with biodiesel percentage. This can be attributed to the different


combustion characteristics of each fuel. On average, the exhaust temperature for diesel fuel, B10, B20 and B50 were determined as 419.2 ℃ , 425.2 ℃ , 434.4 ℃ and 428.2 ℃, respectively. B20 gives maximum exhaust heat loss because of higher exhaust temperature while B10 offers minimum exhaust heat loss. In spite of higher energy input with diesel fuel, higher power output of B10 results in lower energy loss to exhaust. After approximately 2500 rpm exhaust heat loss exceeds the effective power output of the engine. This result is compatible with literature (Pulkrabek, 1997). One item, which needs to be mentioned that there is a big difference between exhaust heat loss rate and exhaust exergy rate for all test fuel, especially lower engine speed. For example, when engine fuelled with diesel fuel exhaust exergy rates are lower 550.442% and 312.10% than exhaust heat loss rates at 1000 and 3000 rpm, respectively. This implies that only a small amount of exhaust heat can be utilized for useful work (Yasar, 2008). However, in order to improve efficiency exhaust heat loss should be evaluated.

Fig. 6: Variations of the exhaust heat loss rate and exhaust exergy rate versus engine speed

Fig. 7 shows total heat loss rate and total heat loss exergy rate as a function of engine speed at full load operation. As engine speed is increased, mechanical friction, combustion temperature and fluid movement within cylinder increase and this leads to more heat transfer form the engine to surroundings. As seen in the figure the operations with B10 results in higher rates of heat transfer from the engine because it gives the lowest exhaust heat loss as taking into account the energy conservation principle (it can be seen from Eq. 5). By fueling the engine with B10, on average 25.5% of the fuel energy is rejected into the atmosphere and this may be a sign of better combustion (Canakci and Hosoz, 2006) which results in higher combustion temperatures with B10. Also, it can be stated that fueling the engine with B10 leads to shorter ignition delay time which reduces the exhaust temperature but increases total heat loss due to more time for heat transfer from the hot gases to surroundings. Total heat loss exergy is proportional to the heat rejection from the engine and therefore, both of them present nearly the same variations but different in magnitude. As illustrated in Fig.7, for tested fuels total heat loss rate values change from 1.28 kW to 4.32 kW, total heat loss exergy rate values are below 1 kW at all engine speeds and this indicates the energy degradation. Total heat loss from the engine can be -5-

reduced by thermal barrier coating. But it causes an increase in exhaust gas temperature and if not recovery, the higher exhaust heat will be rejected to atmosphere with exhaust gases (Parlak et al., 2005).

Fig. 7: Variations of the total heat loss rate and total heat loss exergy rate versus engine speed

Exergy destruction rate with respect to the engine speed at full load operation is shown in Fig. 8. In real processes exergy is not conserved and it is destroyed by irreversible processes in the engine, such as combustion, friction, heat transfer along finite temperature gradients, mixing, etc. As seen in the figure for all test fuel, exergy destruction rate increase with increase engine speed due to increasing irreversibilities. It is detected that more than half of the fuel exergy is destroyed and exergy destruction rate are higher than the effective power of the engine at all engine speed. The average value of the exergy destruction rate for diesel, B10, B20 and B50 are calculated as 6.77, 6.56, 6.69 and 6.67 kW, respectively. The most important source of the irreversibility in the internal combustion engine is the combustion process which destroys a significant fraction of the fuel exergy. It is estimated that owing to combined effect of oxygen content and low viscosity of the B10, by using it combustion improve and consequently exergy destruction decreases. Moreover, it can be noted that higher lubricity of biodiesel reduces the mechanical friction particularly in mechanical injection pump which results in less exergy destruction for blend fuels. First law efficiency and second law efficiency of the engine for different engine speeds are shown in Fig. 9 and Fig. 10, respectively. Both efficiencies show the same behavior but different in magnitude. Because the engine consumes a higher amount of fuel exergy compared to the fuel energy values for the same test fuel, the second law efficiencies lower than the corresponding first law efficiencies. For all test fuels, the first law efficiency of the engine varied from 35.23% to 40.41%, whereas the second law efficiency varied from 33.00% to 37.83%. It is noticed that in Fig. 9 and Fig. 10 biodiesel-diesel fuel blends give better first and second law efficiencies compared to diesel fuel, generally. The improvement in both efficiencies when using biodiesel-diesel fuel blends can be attributed to the oxygen content of the blend fuels which promote a better mixture formation and combustion, and to the higher


lubricity of biodiesel which reduces the mechanical friction loss. But, as increasing biodiesel percentage, the viscosity of the fuel also increase which prevent forming a better air-fuel mixture. For this reason combustion may deteriorate and reduce both efficiencies by increasing biodiesel percentage. When using B10 the positive effect of the oxygen content and good lubricity property become more dominant on the negative effect of the high viscosity. Therefore, maximum first and second law efficiency is obtained at 2000 rpm as 40.41 and 37.83%, respectively by using B10.

V. Conclusions A single cylinder, natural aspirated, four-stroke, air cooled and direct injection diesel engine fuelled with diesel, B10, B20 and B50 was operated at different engine speed and maximum load. Using data gathered from the experiments, energy and exergy components of the engine was calculated and compared with each other for all fuel operation. All test fuels result approximately similar fuel energy, fuel exergy and effective power. But with B10 it is observed slightly higher effective power output than the other fuels particularly at high engine speeds. Although, the sum of energy value of the exhaust heat loss and total heat loss are 59.66-64.77% of the fuel energy, the availability of them are less than 10% of the fuel exergy for all tested fuels. Exergy destruction rate was found to be higher for diesel fuel as compared to blend fuels, on average. For all tested fuel, exergy destruction rate is higher than effective power output at all engine speed. Maximum first and second law efficiency is obtained at 2000 rpm as 40.41% and 37.83%, respectively by fuelling the engine with B10. References

Fig. 8: Variations of the exergy destruction versus engine speed

Fig. 9: Variations of the first law efficiency versus engine speed

.

Fig. 10: Variations of the second law efficiency versus engine speed

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