Investigation on combustion process and emissions characteristic in direct injection diesel engine powered by wet ethanol using blend mode

Fuel Processing Technology 149 (2016) 86–95

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Fuel Processing Technology journal homepage: www.elsevier.com/locate/fuproc

Research article

Investigation on combustion process and emissions characteristic in direct injection diesel engine powered by wet ethanol using blend mode Wojciech Tutak a, Arkadiusz Jamrozik a, Michał Pyrc a, Michał Sobiepański b a b

Czestochowa University of Technology, Institute of Thermal Machinery, Armii Krajowej 21 Av., 42-201 Czestochowa, Poland Czestochowa University of Technology, Institute of Mechanical Technology, Al. Armii Krajowej 21 Av., 42-201 Czestochowa, Poland

a r t i c l e

i n f o

Article history: Received 23 January 2016 Received in revised form 12 March 2016 Accepted 5 April 2016 Available online xxxx Keywords: Combustion Biofuel Ethanol Heat release Supply system

a b s t r a c t This paper presents results of investigation of the engine powered by diesel–ethanol fuel blends with wide range of ethanol fraction. Two supply systems of fuel pump are investigated. It is used two supply systems of fuel pump: low pressure 0.1 bar and high pressure 1 bar. Researches are conducted on 1-cylinder direct injection compression ignition engine with constant rotational speed 1500 rpm. Thermodynamic parameters and exhaust emission of engine powered by fuels blend are taken into account. Heat release rate and pressure rise have been analyzed as well. Test results confirmed that the combustion process occurring in the diesel engine powered by blend takes place in a shorter time than in the typical diesel engine. High pressure supply system (HPSS) allows to combustion blends with higher ethanol fraction, up to 45%. With high pressure supply system the value of indicated mean effective pressure (IMEP) up to 40% of ethanol fuel fraction kept almost constant with value near to 6 bar. For 15% ethanol fuel fraction in HPSS case the COVIMEP was 2.5-times lower then in low pressure supply system (LPSS) case. In both supply systems emission of THC increased or was on near a constant level. With the increase of ethanol fraction in blend the NOx increased as far as combustion process started to deteriorate. © 2016 Elsevier B.V. All rights reserved.

1. Introduction Due to the fact the climatic changes and the slow, but continuous, increase in the average global temperature growing interest in renewable fuels. The main reason for the negative impact of combustion, which takes place in engines, on the climate is the extensive use of fossil fuels [1,2]. The CO2 is the most unfriendly to the environment and it contributes to global warming. The European Parliament passed Directive 2009/28/EC on the promotion of the use of energy from renewable sources. This provision requires EU member states to use 10% of renewable fuels in transport by 2020. A similar concern as compared to liquid fuels as solid fuels attempts to eliminate fossil fuels [3]. There is a growing number of works devoted to the production of biofuels in the fermentation process or the use of algae [4,5]. Diesel engines are widely used due to their high combustion efficiency, reliability, adaptability and cost-effectiveness [6]. One of the main challenges faced by internal combustion engines is to operate with reduced fuel consumption and pollutant emissions. The main pollution problem for diesel engines is Abbreviations: CA, crank angle, deg; CD, combustion duration, deg; CI, compression ignition; COV, Coefficient of Variation, %; EF, ethanol fuel; HPSS, high pressure supply system; IC, internal combustion; ID, ignition delay, deg; IMEP, indicated mean effective pressure, bar; ITE, indicated brake thermal efficiency, %; LHV, lower eating value, MJ/kg; LPSS, low pressure supply system; MFB, mass fraction burned, %; SFC, specific fuel consumption, g/kWh; TDC, top dead center; i, sequence number of engine cycle; p, in cylinder pressure, bar; φ, crank angle, deg; κ, ratio of specific heats; δ, measurement error. E-mail address: [email protected] (W. Tutak).

http://dx.doi.org/10.1016/j.fuproc.2016.04.009 0378-3820/© 2016 Elsevier B.V. All rights reserved.

nitrogen oxides (NOx) and soot. One of the methods to improve combustion process which leads to reduction of toxic components of exhaust gas form diesel engine is to use of oxygenated fuels such as alcoholic fuels: ethanol or methanol. In the first solutions of dual fuel diesel engines used natural gas [7–9]. The gas was delivered into the engine intake system through the mixer [8]. These solutions still work where gas is obtained from landfills, sewage treatment plants, biogas plants and others. If the calorific value of gas is enough high it is itself burned in engines but if its calorific value is to low it is burned in the dual-fuel engines [8,10]. In order to reduce nitrogen oxides emissions especially in industrial gas engines used the two-stage combustion systems [11–13]. In these combustion systems very lean mixtures are combusted with excess air ratio up to 2.0 [11]. Currently, there are using a variety of biofuels in liquid form to power of the conventional engines or engines carrying out dual fuel combustion system. Co-combustion fuels can be realized using dual fuel mode or by using blends of fuels [14–16]. Power to diesel engine in mixture of fuels is realized by using a typical supply system. The difficulty of this mode is limited mixing capability of fuels having different properties. Additionally the fuel blends are not stable and separate in the presence of small amounts of water. This solution is not flexible enough to change the ratio of diesel/other fuel. The major benefit of this solution is no need to introduce design changes to the engine supply system. Small additions of other fuels can be easily burned in the engine. The change usually requires only injection timing control system or ignition timing of engine. Second mode is dual fuel technology which is

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Fig. 1. Diagram of the experimental setup. 1—engine, 2—fuel injector, 3—constant pressure fuel supply system, 4—fuel tank, 5—in cylinder pressure sensor, 6—charge amplifier, 7—PC with data acquisition system, 8—exhaust gases temperature sensor, 9—cooling fan, 10—pulsation damping tank, 11—intake air flowmeter, 12—air filter, 13—excess air ratio sensor, 14—fuel pressure gauge, 15—pressure of fuel supply system.

realized by additional supply system and it is used in diesel engines. Additional fuel is delivered into the intake port of the internal combustion engine. Such solution guarantees close to homogeneous structure of airfuel mixture delivered into engine cylinder. The ignition process is controlled by the injected dose of diesel fuel. This requires the addition of the injector, along with a separate fuel tank, lines and controls [14, 15]. The biggest advantage of this solution is the flexibility to change the ratio of fuel burned. Nowadays renewable fuels are increasingly used in dual fuel engines. Ethanol is a biomass based renewable energy source, which can be produced with relatively low cost. Ethanol for a long time is considered as curious engine fuel. At first ethanol is associated with spark-ignition engines. In recent years ethanol is used as a fuel for diesel engines [14,15]. It is utilized in both methods: blend or dual fuel cocombustion [17–19]. Rakopoulos et al. [20] performed an experimental investigation on the effects of ethanol–diesel blends on the performance and emissions of a diesel engine using 5% and 10% ethanol rates by volume in diesel fuel. They used six-cylinder, turbocharged and after-cooled, heavy

duty, direct injection (DI), Mercedes–Benz engine. The tests conducted using each of the fuel blends, with the engine working at two speeds and three loads. They stated that the use of ethanol mixtures reduced CO and NOx emissions while increased THC emissions and brake thermal efficiency [20]. Sayin and Canakci [21] investigated the effects of injection timing and mixture rates on the engine performance and exhaust emissions of a single cylinder diesel engine. They used ethanol blended diesel fuel from 0% to 15%. The engine load was 15 and 30 Nm. The tests conducted at five different injection timings (21°, 24°, 27°, 30° and 33° CA bTDC). Authors stated that NOx and CO2 emissions increased while brake thermal efficiency, CO and HC reduced with increasing ethanol rates in the mixture. In case of to late injection timing NOx and CO2 emissions decreased and HC and CO emissions enhanced. In case of advanced injection timings, NOx and CO2 emissions increased and HC and CO emissions decreased. Oliveira et al. [22] investigated the effects of fuel blends containing 5, 10 and 15 wt.% of anhydrous ethanol in diesel

Table 2 Properties of diesel fuel and ethanol, [14,15]. Table 1 Main engine parameters. Parameter Number of cylinders Displacement volume Bore Stroke Compression ratio Rated power Crankshaft rotational speed Injection pressure Injection timing Maximum rated power

Unit 1 0.573 90 90 17:1 7 1500 21 17 7.4

– dm3 mm mm – kW rpm MPa deg bTDC kW

Properties

Diesel

Ethanol

Molecular formula Molecular weight Cetane number Lower heating value, (MJ/kg) Density at 20 °C, kg/m3 Viscosity at 25 °C, (mPa s) Heat of evaporation, (kJ/kg) Stoichiometric air fuel ratio Autoignition temperature, (°C) Flame speed, (m/s) Flame temperature, (°C) Carbon content, wt.% Oxygen content, wt.%

C14H30 198.4 51 41.7 856 2,8 260 14.7 230 0.86 2054 87 0

C2H5OH 46 ~11 26.9 789 0.983 840 9.0 425 ~3 2120 52.2 34.8

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Table 3 Composition of fuels blends. Properties

D100

DE05

DE10

DE15

DE20

DE25

DE30

DE35

DE40

DE45

Diesel, (%vol) Ethanol fuel, (%vol) Density, (kg/m3) Lower heating value, (MJ/kg) Heat of evaporation, (kJ/kg) Oxygen content, (%)

100 0 856 41.7 260 0

95 5 852.7 40.9 289 1.74

90 10 849.3 40.2 318 3.5

85 15 845.9 39.5 347 5.2

80 20 842.6 38.7 376 7.0

75 25 839.3 37.9 405 8.7

70 30 835.9 37.2 434 10.4

65 35 832.6 36.5 463 12.2

60 40 829.2 35.7 492 13.9

55 45 825.8 35.0 521 15.7

fuel with 7% of biodiesel on performance, emissions and combustion characteristics of a diesel power generator. The engine tested with the fuel blends directly injected into the combustion chamber, and the applied load varied from 5 to 37.5 kW. The results compared with standard biodiesel operation indicated that in-cylinder peak pressure and heat release rate decreased at low loads and increased at high loads with the use of ethanol. Increasing ethanol concentration causes increases ignition delay, decreases combustion duration and reduced of exhaust gas temperature. The use of ethanol decreased carbon dioxide (CO2) emissions. Carbon monoxide (CO), total hydrocarbons (THC) and oxides of nitrogen (NOx) emissions showed different behavior, depending on load and ethanol concentration. Some researches used to supply diesel engines other liquid fuels such as n-butanol or gasoline. Lennox et al. [23] presented results of burning of 5%, 10%, and 20% shared volume of n-butanol with diesel dual fuel engine. The aim was to compare the effects of the blends on the engine combustion characteristics and regulated emissions namely nitrogen oxides (NOx), unburned hydrocarbon (UHC), carbon monoxide (CO) and soot. Using n-butanol shared volume significantly improved the reduction of regulated emissions. They reached reduction in soot 55.5%, 77.8%, and 85.1% respectively; CO reduction was 35.7%, 57.1% and 71.4%; NOx increase was 10.3%, 32.3% and 54.4%; UHC increase, 21.4%, 71.4%, and 214% respectively. The premixed phase combustion was amplified and distinguishable with increase of shared volume of n-butanol in DF. The combustion cycles of the blends were more stable than the cycles of dual fuel mode. However, total fuel cost for used n-butanol/diesel fuel blends becomes higher than diesel fuel [24]. Şahin and Durgun [25] used the fumigation method, they injected gasoline into intake air, either by a carburetor, which main nozzle section was adjustable or by a simple injection system. They used four-cylinder, water-cooled indirect injection (IDI), Ford XLD 418 T automotive diesel engine. The effects of gasoline fumigation at (2, 4, 6, 8, 10, 12)% (by vol.) gasoline ratios on the combustion, NOx emission, fuel economy, and engine performance of automotive diesel engine was investigated. Gasoline fumigation test results showed that NOx emission is lower than that of neat diesel fuel. Vallinayagam et al. [26] investigated the emission reduction potential of pine oil, a plant based bio-fuel, when fumigated in a single cylinder diesel engine. From the experimental investigation, it was observed that pine oil can replace diesel up to 60% and 36%, at low and full load conditions, respectively. Significantly, smoke emission was drastically

reduced by 64.2% than normal diesel operation at full load condition, with a slight increase in NOx (oxides of nitrogen) emission. Moreover, CO and HC emissions have been found to be 67.5% and 47.8% lower than that of diesel at full load condition. Overall, investigations showed that biodiesel–alcohols reduce NO emissions while increasing CO and HC emissions, perhaps due to the cooling effects of alcohols [27]. In this paper are presented results of investigation of one-cylinder direct injection air cooled diesel engine in which is made the attempt to burn a mixture of diesel–ethanol. This type of engine, which is still manufactured in Poland, is equipped with low pressure supply systems of fuel pump. In case using diesel fuel this supply system work properly but when we try to power the engine by blends of diesel–alcohol it is noted a disturbing loss of power already at more than a dozen percent alcohol. The power drop was larger than is apparent from the analysis of the calorific value of the fuel dose. It turned out that the cause lies in the filling level of the high-pressure section of injection pump. During testing, it was found that the power of the fuel injection pump under a pressure of at least 1 bar significantly improves the engine performance. Further increasing the supply pressure of the pump no longer affects engine performance. The paper presents a comparative analysis of the engine parameters and the toxicity of exhaust gases to the engine powered by factory supply system and with overpressure 1 bar. 2. Experimental setup The engine used for this study was a naturally aspirated, 1cylinder, Andoria 1CA90 direct injection diesel engine. This engine is still manufactured in Poland. The experimental setup is shown in Fig. 1. Detailed engine specifications are presented in Table 1. Tests conducted at a constant angle of diesel fuel injection, full load and constant rotational speed equal to 1500 rpm. During the research recorded 100 consecutive engine cycles with resolution 1 deg of CA. For each investigated engine operating point the measurement repeated 3 times. It was recorded simultaneously: rotational speed of engine, air and fuel consumption, air temperature, fuel temperature, exhaust gas temperature, ambient temperature and pressure. The excess air coefficient was constantly monitored. Using the exhaust gas analyzers recorded changes in concentration of the components in the engine exhaust gases, such as NOx, HC, CO, CO2, and O2.

Fig. 2. Volumetric fuel fractions (a) and lower heating value of fuels mixture (b).

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Fig. 3. Energy dose (a) and heat of evaporation of mixture dose (b).

In the first stage of the study investigated the engine powered by pure diesel fuel and the results are taken as a reference to further research. In main stage of the investigation the test engine is fueled with various fuels blends of diesel/ethanol in volumetric fractions: 0–45% with increment 5%. In the study used two supply systems of the engine injection pump. The first supply system ensures overpressure before the injection pump equal 0.1 bar (factory solution) and named it Low Pressure Supply System (LPSS). The second supply system is proposed by the authors, which allows obtain higher pressures before fuel pump and named it High Pressure Supply System (HPSS). During the tests work it turned out that the overpressure of 1 bar or higher significantly improves engine performance. For testing assumed the smallest overpressure equal to 1 bar. The researches carried out twice once with LPS and the second time with the HPSS. It turned out that the use of HPSS allows operate of the engine up to 45% of ethanol fuel in blend. It was 10% higher fraction of EF in comparison with LPSS. As the limit of the EF share was clearly noticeable deterioration in engine operation conditions and clearly visible cycles without burning. 2.1. Test stand Modification of fuel injection system was necessary as well to avoid operation of mechanical fuel injection regulator. The test stand equipped with complex measurement system and additional supply system with regulated pressure of fuel supplied mechanical fuel pump (Fig. 1). In Table 1 are presented main engine parameters. The study was conducted on the test bench, which included the following elements: − modernized engine 1CA90 adapted for a multi-fuels powering, − indication system,

− engine power system with a various fuel pressure, 0.1 bar (LPSS), 1.0 bar (HPSS), − Exhaust gas analyzer: THC, CO, CO2, O2—Bosch BEA 350:

THC CO CO2 O2 λ

Range 0–9999 ppm vol Range 0–10 vol.% Range 0–18 vol.% Range 0–22 vol.% Range 0.5–9.999

Accuracy: 12 ppm vol Accuracy: 0.06 vol.% Accuracy: 0.4 vol.% Accuracy: 0.1 vol.% Accuracy: 0.01

− Exhaust gas analyzer: NOx—Radiotechnika AI9600:

Accuracy

±32 ppm for range 0–1000 ppm,

Indication system: − Digital measurement system for acquisition and analysis of fast changing data, − Measuring system for fuel consumption, − Digital system for measuring the engine rotational speed based on the encoder position of the crankshaft with the resolution of 1 deg CA. − Measuring system for air flow.

Digital measurement system for data acquisition: − piezoelectric pressure transducer, Kistler 6061 SN 298131, sensitivity: ±0.5%, − charge amplifier, Kistler 5011B, linearity of FS b±0,05%, − data acquisition module, Measurement Computing USB-1608HS—16 bits resolution, sampling frequency 20 kHz, − computer PC, − crank angle encoder, resolution 360 pulses/rev, − software for digital recording and analysis of the frequency signals [32].

A system for measuring air flow: − expansion tank with a capacity of 150 dm3, − temperature sensor TP-204K-1b-100-1.5, range +200 °C, resolution 0.1 °C, − Roots flowmeter Common CGR-01, range 0.2–650 m3/h, accuracy b ± 1%.

Fig. 4. Mass fraction of oxygen in the blends.

At each engine mode, the engine was allowed to run for a few minutes until the exhaust gases temperature has reached steady-stated values and the data were measured subsequently. All the emissions were measured 3 times at the exhaust pipe of a test engine.

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Fig. 5. In cylinder pressure traces of the engine with: (a) LPSS, (b) HPSS.

2.2. Fuel blends characteristic In comparative studies used wet ethanol blended with diesel fuel in volumetric fraction up to 45% with increment of 5%. The properties of fuels used during tests are presented in Table 2. Alcohol fuel can be used in CI engines as pure, by blending with conventional diesel fuel or using dual fuel technology. Ethanol can be produced from any plants which contains of sugar or other components which can be converted into sugar, such as starch or cellulose in the fermentation, distillation and dehydration process. This is especially important taking into account the global trend to reduce CO2 emissions. CO2 is the main greenhouse gas and the CO2 released by biofuel combustion can be fixed by growing plants and therefore makes no net contribution to global warming [28]. Alcohol, due to oxygen content in structure (CaHbOc), requires less air to maintain the same air-fuel equivalence ratio if compared to hydrocarbon based fuels [15]. Additionally, LHV of alcohol is lower in comparison to diesel fuel hence to obtain the same engine performance the higher amounts of alcohol should be provided [15]. The most important disadvantage encountered in case of alcohol– diesel blend is the phase separation. Analyzing the usefulness of alcohol to co-combustion in IC engine it should be taken into account its properties such as: lower heating value and also stoichiometric air fuel ratio and heat of evaporation. Ethanol fuel (EF) belongs to the group of oxygenated fuels. Oxygen content of ethanol is equal to 34.8%. This high oxygen content has a big influence on the combustion phases of the diesel engine. In a diesel engine, there are two combustion phases, the premixed combustion phase and diffusion combustion phase. The oxygenated fuel causes the change in the ratio of these phases. With the increase in EF fraction it is observed the intensification of the premixed combustion phase and on the other hand it decreases diffusion phase. Ethanol is characterized by higher carbon content and it is therefore more energy

fuel than for example methanol. Ethanol has low stoichiometric air–fuel (A/F) ratio, high oxygen content and high H/C ratio may be beneficial at improving the combustion process. In Table 3 are presented the composition of used fuels blends during the tests. Diesel fuel is used as a reference fuel to compare obtained results. To calculate properties of used blends used the data from Table 2. Blends of diesel fuel and ethanol fuel are named for example: DE20—D—diesel fuel, E—ethanol fuel, 20—20% of ethanol fuel volumetric fraction in blend. In addition, ethanol fuel (EF) means mixture of ethanol with water, wet ethanol. Ethanol fuel used during tests was wet ethanol which consisted of 89% of C2H5OH and 11% of water. With higher share of EF in the blend the water content growth as well. This water and ethanol content in blends is presented in Fig. 2a. The higher water content is noticed in the blend DE45, of course, and it is equal to 5.0%. In Fig. 2a are presented the volume fraction of three components of blend. In case of 30% ethanol fuel fraction the water fraction is equal to 2.2% but in case of 45% of ethanol fuel the fraction of water is near to 5%. In Fig. 2b are presented the lower heating value (LHV) of used blends. Increase of EF fraction in blend causes a decrease in the heating value of blend. It is caused by lower LHV of bioethanol in comparison with diesel fuel and in addition the water fraction reduces the energy value of fuels blend. In Fig. 3a are presented the energy doses of used fuel blends. With the increase of EF in blend the energy dose decreases which is directly related to LHV of mixture of both fuels (Fig. 2b). In case of blend with 45% of ethanol fuel (DE45) LHV of blend is 16% lower in comparison with diesel fuel. In Fig. 3b are presented the values of heat of evaporation of blends. With the increase in EF fraction in blends increases the value of heat of evaporation of blend. In case of 45% of ethanol fuel (DE45) the heat of evaporation is equal to 623 kJ/kg and it is 2.4-times higher than of diesel fuel. In case of blends the higher dose of energy is necessary to evaporate of this same dose of fuels blend.

Fig. 6. Pressure vs. cylinder volume (p − V), (a) LPSS, (b) HPSS.

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Fig. 7. Pressure rise dp/dφ traces of the engine with: (a) LPSS, (b) HPSS.

The oxygen fraction in used blends is presented in Fig. 4. In case of blend with 20% of ethanol fuel (DE20) the oxygen mass fraction is equal to 6.2% but in case of 45% of ethanol fuel (DE45) the oxygen mass fraction is equal to 13.9%. Actual air–fuel ratio decreases with increasing EF fraction. The stoichiometric air–fuel ratio of ethanol is 9.0 which it is significantly lower than diesel's 14.7.

COVIMEP ¼

3. Results and discussion Analysis of combustion process in the internal combustion engine usually is carried out with the rate of heat release. Heat release rate (dQ/dφ) is calculated on the basis of the measured in-cylinder pressure data and crank angle readings. The basis for determining the heat release rate is the first law of thermodynamics and the equation of state. After rearranging and simplifications, the heat release rate vs. crank angle is obtained in well-known form as follows:
 dQ 1 dV dp ¼ κp þV dφ κ−1 dφ dφ

the COVIMEP. The COVIMEP is calculated on the basis of IMEP data obtained from experimental results analysis. In this case this factor is determined on the basis of 100 cycles. The COVIMEP is a parameter that represents cyclic variability and it is defined as the standard deviation in IMEP divided by the mean IMEP:

ð1Þ

σ IMEP  100 ðIMEPÞmean

where: σIMEP—standard deviation of IMEP. The mean IMEP was determined as: i¼100 X

ðIMEPÞmean ¼

IMEPi

i¼1

100

where: i—a sequence number engine cycle, The COVIMEP is directly related to the combustion stability. 3.1. Engine performance

where: κ—the ratio of specific heats, V—cylinder volume, p—in cylinder pressure.Instantaneous cylinder volume V is precisely described by engine geometry. Due to omitting as follows: heat transfer to walls, crevice volume, blow-by and the fuel injection effect, the resulted heat release rate is termed as the net heat release rate. The cumulative net heat released is obtained by integrating Eq. (1) over the crank angle φ [6,16]. Rate of pressure rise dp/dφ is determined as: Δp pi −pi−1 ¼ Δφ φi −φi−1 where: κ − p—in cylinder pressure, φ—crank angle, The cycle-by-cycle variations of combustion are usually investigated by experiments, and the measure of cyclic variability is represented by

During the study considered the impact of mechanical diesel fuel pomp supply system. The motivation of this research was the fact that older stationary engines are powered by fuel blends using the gravity supply systems of injection fuel pump of that engine. The example of such engine is 1CA90 used to power small equipments or pumps. Within the study investigated the effect of supply pressure of fuel pump on engine performance and emission. It was used two supply systems: low pressure supply system (LPSS) with pressure equal to 0.1 bar and high pressure supply system (HPSS) with pressure 1.0 bar. In Fig. 5 are presented the pressure traces of blends combustion using both supply systems. In case of LPSS the limit of ethanol fuel volumetric fraction was up to 35%. This limit, in case of high pressure supply system (1.0 bar) HPSS, was achieved at 45% of ethanol fuel

Fig. 8. Heat release rate in analyzed cases, (a) LPSS, (b) HPSS.

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Fig. 9. Mass fraction burned in analyzed cases, (a) LPSS, (b) HPSS.

fraction (DE45) in blend. It should be mentioned that the blends of fuels were made using an electromagnetic mill. To created blends do not add any chemicals to prevent phase separation. Using the LPSS observed a small increase in the maximum value of pressure in the combustion chamber. In case of LPSS the increase in peak pressure was 1.5 bar and is achieved with DE25 but using HPSS the increase was near to 6 bar at DE30. DE30 blend which was the limit of LPSS could be successfully burned by using the HPSS. In Fig. 6 the pressure traces vs. cylinder volume where is clear visible differences in pressure courses using both supply systems. In zoom area are presented visible negative loops. Such negative area limits operating parameters of the engine. In case of LPSS already with 20% of ethanol fuel (DE20) the negative area on p-V chart is noticed. In case of HPSS this is noticed at 35% of ethanol fuel fraction (DE45). With HPSS noticed the higher pick pressure. It was because high pressure volume of fuel pump is better filled by fuel. This will have a positive impact on better performance of the engine. In Fig. 7 are presented the pressure rises of analyzed cases. In case with lower pressure supply system (LPSS) with the increase ethanol fuel (EF) fraction in blend the peak pressure rise decreases and it is achieved for larger angles after TDC. In case of HPSS with the increase EF fraction the peak pressure rise increases up to 20% of ethanol and than it started to decreases. With 20 and 25% EF fraction the pressure rise dp/dφ is higher than 10 bar/deg and it is too high value to diesel engine. This causes so called hard operate of the engine. In Fig. 8 are presented the heat release rate in both analyzed cases. In first case it is noticed small increase in peak value of heat release rate up to DE30. After this limit the combustion process clearly began to deteriorate. This is probably due to high value of heat evaporation of ethanol which takes a lot of heat from the compressed charge and it difficult the autoignition process of fuel. On the basis of these traces it can be stated that the cooling effect of ethanol in blend is visible on each cases. With HPSS this impact is lower up to DE35. In the second case with the increase of EF fraction in blend up to 20% the maximum value of heat release rate grows. In case of DE30 with LPSS the peak of dQ/dφ is reached

at 9 deg after TDC but in case of HPSS the peak of dQ/dφ is reached at 7 deg after TDC. Using HPSS it was possible to burn of larger fraction of EF in the blend. In Fig. 9 are presented the mass fraction burned (MFB) traces obtained for both analyzed cases. On the basis of these traces determined the phases of combustion process. For combustion process in the IC engine two phases are the most important, first it is ignition delay (ID) and second the combustion duration (CD). The ignition delay is defined as the time between the start of diesel fuel injection and the crank angle of 10% of mass fraction burned (CA 0–10% MFB) [29]. This delay period consist of physical delay and chemical delay which occur simultaneously. In the physical delay takes place atomization, vaporization and mixing of air fuel, and in the chemical delay attributed pre-combustion reactions [30]. Burn duration is also calculated by reading the time between CA 10–90% the crank of mass fraction burned (CA 10–90% MFB). The results of phase determination are presented in Fig. 10. In Fig. 10 are presented the ignition delay (ID) and combustion duration (CD) in both analyzed cases. In case of ID it can be stated that with increase of pressure in the supply system of this factor decreases. It is advantage for engine. In both cases with the increase of EF fraction ignition delay increased as well. In case of CD in both realizations with the increase of EF fraction this factor decreased. It is associated with alcohol fuel properties which are analyzed in Section 2.2. At low pressure in supply system the combustion duration decreased almost linearly with all ethanol percentages. At high pressure supply system the combustion duration decreasing until 15% of ethanol fuel fraction and after it has remained near to constant up to 35% of EF. In Fig. 11 are presented the indicated thermal efficiency (ITE) and indicated mean effective pressure (IMEP). It could be stated that in case of HPSS obtained the better parameters of engine performance. This is probably due to the better filling by fuel blend the chamber of the injection pump. The biggest difference in value of indicated thermal efficiency noticed with blend DE35 (with 35% ethanol fuel) which was equal to 6.5%. In case of high pressure system IMEP to 40% of ethanol fuel fraction was kept almost constant with value near to 6 bar. In

Fig. 10. Combustion phases, (a) ignition delay, (b) combustion duration.

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Fig. 11. Indicated thermal efficiency and indicated mean effective pressure in the test engine, (a) LPSS, (b) HPSS.

Fig. 11 are presented the uncertainty errors calculated on the basis of 150 cycles. It is also calculated the measurement error of IMEP which consists of the error of in-cylinder pressure measurement and the error of measurement of the instantaneous displacement. The error of piezoelectric pressure transducer is δpt = 0.5% and the amplifier δa = 2.0%. The measurement error of IMEP is δIMEP =2.1% and error of ITE is δITE =3.2%. In Fig. 12 the specific fuel consumption (SFC) and cycle variation expressed as COVIMEP. On the basis of results can be stated that with high pressure supply system (HPSS) the SFC was kept almost constant up to 35% ethanol fuel fraction. In case of low pressure supply system (LPSS) with the increase EF fraction in blend the SFC increased simultaneously. For 35% EF fraction the difference is equal to 64 g/kWh that means the difference in SFC is near to 20%. For IC engine operation it would be ideally if each successive cycle is the same. This would ensure a very smooth engine running. In case of HPSS especially for small ethanol fuel fraction the unrepeatability of cycles is significantly lower then in LPSS case. For 15% EF fraction in HPSS case the COVIMEP is 2.5-times lower. Overall, specific fuel consumption of alcohol blends is higher than for diesel fuel. The measurement error of SFC is δSFC = 4%. The uncertainty errors of SFC and COVIMEP are presented on Fig. 12. 3.2. Emission characteristic The compression ignition engine emission is generally characterized by nitrogen oxides (NOx) and soot emission. The nitrogen oxides formation rate has a strong relationship to temperature. High temperature in the combustion chamber of the engine during relatively short phase of premixed combustion is characterized by high NOx formation rate. The NOx formation can be reduced by decreasing the premixed burning rate using fuel injection strategy. It is one way to decrease the high peak temperature in the combustion chamber which cause lower rate of NOx formation. On the other hand the decrease temperature in the combustion chamber causes lower thermal efficiency of the engine. As is known, the engine reaches its highest efficiency close to the knock

limit, when there is also the high temperature. Further, it is difficult to reconcile the reduction of NOx emissions with a reduction in soot emissions. Technology applied in the diesel engines which decreased NOx emission simultaneously increased soot emission [31]. Emission of NOx is combined with premixed phase of combustion process but the soot emission is combined with diffusion phase. A proper control of these two phases of combustion can leads to low emission of NOx and soot. For emission are presented the uncertainty errors calculated on the basis of analyzers data presented in Section 2.1. In Fig. 13 are presented emissions of total hydrocarbons (THC) and nitrogen oxides (NOx). In case of THC emission it can be stated that in both supply systems it is increased or kept near a constant level. LPSS is characterized by higher THC emission with higher EF fraction due to the deterioration of the combustion process and significant unrepeatability of the engine cycles. There are noticed cycles with very low pressure which indicates the absence or disappearance of combustion processes. In both cases is similar trend in nitrogen oxides emission. With the increase of EF fraction in blend the NOx increased as far as combustion process started to deteriorate. With high EF fraction the value of heat of evaporation of blend become too high for the proper combustion process occurrence. In this case, too much heat is received from the compressed air and the ignition process becomes difficult. In Fig. 14 are presented the CO and CO2 emissions. On the basis of these components it can be concluded of combustion process efficiency. In case of LPSS emission of CO is lower then in HPSS because despite the same volume of fuel delivery the mass of fuel was lower and almost whole dose of fuel blend had time to burn. In both cases, with maximal EF fraction, reached the decrease in CO emission respectively of 8-times (LPSS) and 5-times (HPSS) in comparison with diesel engine. In case of CO2 emission is higher at HPSS because there is given a bigger dose of fuel into the cylinder which generates higher level of CO2 emission. It is clearly seen in Fig. 14 that CO emission decreases with increase ethanol concentration. LPSS is characterized by higher decrease in CO2 emission but it is connected with lower engine performances.

Fig. 12. The specific fuel consumption (SFC) and unrepeatability of IMEP.

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Fig. 13. Emission of THC and NOx.

Fig. 14. Emission of CO and CO2.

In Fig. 15 are presented the oxygen content and excess air ratio factor. Ethanol is included in the group of oxygenated fuels (C2H5OH) and as mentioned in Section 2.2, ethanol in mass contains near to 35% of oxygen. This is reflected in stoichiometric air fuel ratio which is more then 60% less then for diesel fuel. The control of the diesel engine generally is realized by changing the fuel delivery. With EF fraction increase, as presented in Fig. 4a, the oxygen content in exhaust increases simultaneously. In connection with this fact to burn subsequent doses fuel less and less oxygen was required. The not used oxygen by combustion process in the air is reflected in the exhaust. In case of HPSS noticed the increase in excess air ratio of 28% and in case of LPSS this increase was equal to 47%. 4. Conclusions On the basis of investigations can be stated that diesel–ethanol blends could be utilized in the compression ignited engines with quite large fraction. Using ethanol fuel in blend with diesel fuel it is supplied additional oxygen for the combustion which can cause improvements in overall combustion process. The combustion process of used blends

in IC diesel engine takes place in a shorter time in comparison with the diesel engine fuelled by pure diesel fuel. The main conclusions: − HPSS allows to burn blends with higher ethanol fraction, up to 45%, − With 20 and 25% EF fraction the pressure rise was higher than 10 bar/deg and it is too high value to diesel engine, with lower and higher ethanol fuel fractions maximal value of this parameter is acceptable and ensures smooth work of engine. − In case of high pressure system IMEP up to 40% of EF fraction was kept almost constant with value near to 6 bar. − With HPSS the SFC kept almost constant up to 35% EF fraction. In case LPSS with the increase EF fraction in blend the SFC increased simultaneously. − In case of HPSS especially for small EF fraction the unrepeatability of cycles was significantly lower then in LPSS case. For 15% ethanol fuel fraction in HPSS case the COVIMEP was 2.5-times lower. − In both cases with the increase of EF fraction ignition delay increased as well. In case of combustion duration in both realizations with the increase of EF fraction this factor decreased. With

Fig. 15. Oxygen content in exhaust (a) and excess air ratio λ (b).

W. Tutak et al. / Fuel Processing Technology 149 (2016) 86–95

HPSS the combustion duration was decreasing until the value of EF fraction 15%, after it has remained near to constant. − In both supply systems emission of THC increased or was at near a constant level. LPSS characterized by higher THC emission with higher EF fraction. − With the increase of EF fraction in blend the NOx increased as far as combustion process started to deteriorate.

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