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Diethyl ether as an ignition improver for biogas homogeneous charge compression ignition (HCCI) operation - An experimental investigation
Energy 35 (2010) 3614e3622
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Diethyl ether as an ignition improver for biogas homogeneous charge compression ignition (HCCI) operation - An experimental investigation K. Sudheesh, J.M. Mallikarjuna* Internal Combustion Engines Laboratory, Department of Mechanical Engineering, Indian Institute of Technology Madras, Chennai 600 036, India
a r t i c l e i n f o
a b s t r a c t
Article history: Received 24 October 2009 Received in revised form 28 April 2010 Accepted 29 April 2010 Available online 14 June 2010
This paper deals with experimental investigations of a homogeneous charge compression ignition (HCCI) engine using biogas as a primary fuel and diethyl ether (DEE) as an ignition improver. The biogas is inducted and DEE is injected into a single-cylinder engine. For each load condition, best brake thermal efficiency DEE flow rate is determined. The results obtained in this study are also compared with those of the available biogas-diesel dual-fuel and biogas spark ignition (SI) modes. From the results, it is found that biogas-DEE HCCI mode shows wider operating load range and higher brake thermal efficiency (BTE) at all loads as compared to those of biogas-diesel dual-fuel and biogas SI modes. In HCCI mode, at 4.52 bar BMEP, as compared to dual-fuel and SI modes, BTE shows an improvement of about 3.48 and 9.21% respectively. Also, nitric oxide (NO) and smoke emissions are extremely low, and carbon monoxide (CO) emission is below 0.4% by volume at best brake thermal efficiency points. Also, in general, in HCCI mode, hydrocarbon (HC) emissions are lower than that of biogas SI mode. Therefore, it is beneficial to use biogas-DEE HCCI mode while using biogas in internal combustion engines. Ó 2010 Elsevier Ltd. All rights reserved.
Keywords: Biogas Diethyl ether HCCI Ignition improver
1. Introduction Generally, spark ignition (SI) and compression ignition (CI) engines are most commonly used for power generation and in transportation vehicles. The SI engines have low smoke emissions and brake thermal efficiency with high HC, CO and NOx emissions. However, the CI engines show comparatively higher brake thermal efficiency, but they emit high amounts of smoke and NOx. In these engines, achieving low exhaust emissions with low fuel consumption is a challenge. Also, today’s emission legislations are forcing engine manufacturers to search for engines having low exhaust emissions and higher fuel economy. Today, HCCI is emerging as an effective alternative combustion process for CI mode. With certain fuels, it can provide higher brake thermal efficiency like CI engines with ultra-low NOx and particulate matter (PM) emissions. In these engines, the premixed nature of a charge effectively eliminates the PM, whereas combustion with lean mixture without definite flame propagation reduces cylinder gas temperatures leading to ultra low NOx emissions simultaneously. Biogas is an attractive source of energy for rural areas especially in countries like India. It is generated by anaerobic digestion of cow dung, other animal wastes and plant matters such as leaves and
* Corresponding author. Tel.: þ91 44 22574698; fax: þ91 44 22574652. E-mail address: [email protected] (J.M. Mallikarjuna). 0360-5442/$ e see front matter Ó 2010 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2010.04.052
water hyacinth. All the above mentioned sources are renewable in nature and abundantly available. Biogas contains approximately two-thirds (by volume) of methane (CH4) and the rest is mostly carbon dioxide (CO2) with traces of hydrogen sulphide (H2S). Biogas has a low energy density due to the presence of CO2. The properties of biogas are given in Table 1. Usage of biogas in SI engines enhances knock resistance due to the presence of CO2. However, it reduces brake thermal efficiency and increases HC emissions [1,2]. It is not possible to use biogas directly in CI mode due to its higher autoignition temperature. Dual-fuel mode of operation is a feasible way of using biogas in CI engines with diesel as a secondary fuel [3,4]. But, dual-fuel mode emits more HC and CO emissions with lower brake thermal efficiency. However, smoke and NOx emissions are lower as compared to the conventional CI mode due to the gaseous nature of biogas and the presence of CO2. Diethyl ether is considered as a renewable fuel because it can be produced from ethanol through the dehydration process [5]. Low autoignition and boiling temperature of DEE are reasons for selecting it as an ignition improver along with lethargic fuels like biogas. Also, DEE has a higher energy density than ethanol. The important properties of DEE are given in Table 2. Generally, DEE is used as a cold starting aid in CI engines. DEE is also blended with diesel for improving the brake thermal efficiency and reducing emissions in CI engines [6]. Onishi et al. (1979) introduced controlled autoignition combustion in a two-stroke engine in order to reduce instability at
K. Sudheesh, J.M. Mallikarjuna / Energy 35 (2010) 3614e3622 Table 2 Properties of DEE.
Nomenclature
lBG lDEE lT Mair MBTE NOP MBG MDEE AFBG AFDEE CAD
Calorific value Density Boiling point Stoichiometric air fuel ratio (mass basis) Autoignition temperature Cetane number
biogas excess air ratio DEE excess air ratio total excess air ratios mass flow rate of the air best brake thermal efficiency nozzle opening pressure mass flow rate of the biogas mass flow rate of the DEE biogas stoichiometric airefuel ratio DEE stoichiometric airefuel ratio crank angle degree
part-loads and they also achieved good reduction of emissions and fuel consumption [7]. Najt et al. (1983) extended the HCCI combustion into a four-stroke engine using primary reference fuels [8]. A four-stroke gasoline HCCI engine was tested by Thring et al. (1989) and the important parameters required for successful gasoline HCCI operation at part-loads were investigated [9]. Ryan et al. (1996) used a port fuel injection (PFI) injector to supply diesel into the intake air stream at various inlet air temperatures and compression ratios [10]. This resulted in early heat release during compression stroke itself, and they concluded that low compression ratio is most suitable for port injected diesel fuelled HCCI engine. Garcıa et al. (2009) investigated the effect of inlet charge temperature, cool EGR, injection timings and equivalence ratio on the performance of a diesel fuelled HCCI engine and achieved comparatively higher loads by using cool EGR [11]. Shi et al. (2006) studied the effect of internal and externally cooled EGR on the performance of a diesel fuelled HCCI engine. They concluded that internal EGR benefited the formation of homogeneous mixture and reduced smoke emission, whereas externally cooled EGR could help extend upper load limit of HCCI operation [12]. Due to the early heat release characteristics with low volatility of diesel, researchers tried to use various alternative fuels with high autoignition temperature, viz. natural gas, methanol, ethanol, liquefied petroleum gas (LPG) without many modifications to the original engine. Inlet charge heating, variable compression ratio (VCR) and use of secondary low octane fuels as an ignition improver were methods tried for achieving HCCI combustion with the above fuels. Christensen et al. (1997) investigated the HCCI combustion characteristics using natural gas with inlet air heating [13]. Swami Nathan et al. (2008) adopted acetylene as a fuel for HCCI engine because of its moderate autoignition temperature and high flammability limits. They used inlet charge temperature to control combustion phasing [14,15]. Chen et al. (2000) introduced dual-fuel HCCI combustion of natural gas with dimethyl either (DME). They
Table 1 Properties and composition of biogas. Calorific value Density (1 atm and 15 C) Flame speed Stoichiometric air fuel ratio (mass basis) Autoignition temperature Flammability limits with air (%) Research octane number Typical biogas composition in % volume Methane Carbon dioxide Carbon monoxide
3615
17 MJ/kg 1.2 kg/m3 0.25 m/s 5.7 650 C 7.5e14 130 57.37 42.1 0.08
33.9 MJ/kg 713 kg/m3 34.4 C 11.1 160 C >125
found that by optimizing a proportion of DME and natural gas, NOx emissions could be lowered to near zero levels. The dual-fuel operation gave higher brake thermal efficiency than that of CI mode [16]. Zheng et al. (2004) used DME as an ignition controller in a methanol fuelled HCCI engine [17]. Mack et al. (2009) experimentally proved the suitability of wet ethanol in HCCI combustion. They used inlet charge heating to control the combustion phasing [18]. Swami Nathan et al. (2008) investigated biogas HCCI combustion with manifold injection of diesel. They compared the results of biogas-diesel HCCI mode with that of biogas-diesel duelfuel mode of operation. Due to the low volatility and high boiling temperature of diesel, inlet heating was used for proper mixing of diesel with manifold inducted biogas. They concluded that biogasdiesel HCCI operation is superior to dual-fuel mode of operation in a BMEP range of 2.5 to 4 bar [19]. However, they couldn’t operate an engine in HCCI mode below 2.5 bar and above 4 bar BMEPs due to system limitations and difficulty in controlling inlet charge heating. The motivation for the present work is to operate a single-cylinder engine in biogas fuelled HCCI mode in a wide load range by using DEE as an ignition improver. These types of studies are limited in literature. Therefore, a detailed study would help not only to evaluate and compare the HCCI mode with conventional biogas SI and biogas diesel-dual-fuel modes, but also to find the feasibility of using biogas in HCCI mode. In this study, the effect of DEE flow rate on the performance, emissions and operating load range of a biogas fuelled HCCI mode are studied at a constant engine speed of 1500 rev/min., and cooling water outlet temperature of 50 C. Finally, DEE mass flow rate for best brake thermal efficiency point at each load condition has been found out. In addition, the performance and emission characteristics of biogas-DEE HCCI mode are compared with the available results of biogas SI [1,2] and biogas-diesel dual-fuel [3] modes. 2. Experimental setup A single-cylinder, water-cooled, direct injection CI engine is used for conducting experiments. The engine is coupled to an eddy current dynamometer for loading and measurement purposes. The engine specifications are shown in Table 3. The DEE is stored in an accumulator and is injected into the intake manifold using an injector at a line pressure of 2 bar. An in-house built electronic circuit is used for controlling DEE flow rate. The biogas is directly inducted through the intake manifold. The biogas is generated in a nearby plant, which uses cow dung and water to produce it. It is collected in a flexible bag at the plant and transported to the place of usage. It is actually sent through a floating drum in order to maintain the Table 3 Engine specifications. Bore stroke Connecting rod length Compression ratio Rated power output Displacement volume Injector NOP
80 110 mm 231 mm 16:1 3.7 kW @ 1500 rpm 553 cm3 220 bar
3616
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1. Air flow measurement
13. Cooling water inlet
2. DEE supply unit
14. Cooling water flow regulator
3. DEE pressure gauge
15. Cooling water out let temperature sensor
4. DEE pulse width controller
16. Cooling water outlet
5. DEE fuel injector
17. Engine
6. Biogas flow meter
18. Exhaust gas temperature sensor
7. Biogas flow control valve
19. Eddy current dynamometer
8. Biogas floating drum
20. Electromagnetic clutch
9. Pressure transducer
21. Electric motor
10. Crank angle encoder
22 to 25. Exhaust gas analyzers
11. Data acquisition system
26. Engine exhaust
12. Personal computer Fig. 1. Schematic of experimental setup.
required constant pressure as in the plant. The biogas composition is measured every time using a MRU make non-dispersive infrared (NDRI) analyzer meant for this purpose. A typical composition of the biogas is shown in Table 1. The flow rate of biogas is controlled by a fine-control valve and is measured with the help of a turbine type
gas flow meter before being admitted to intake manifold. The cooling water outlet temperature is measured by a resistance temperature detector (RTD) and the air flow rate is measured by using a dry turbine type flow meter. Exhaust gas analyzers working on the principles of FID for HC, NDIR for CO and CLD for NO, respectively, are used to measure the emission levels. A Bosch smoke meter is used to measure smoke emissions. Figs. 1 and 2 show schematic and photographic views respectively of the experimental setup developed and used in this study. 3. Experimental procedure First, the engine is motored at about a speed of 600 rev/min., and simultaneously DEE is injected into the intake manifold. Thus, the engine is started in HCCI mode with DEE as the only fuel and
Table 4 Mixture strengths used in biogas-DEE HCCI mode.
lDEE ¼ 6.1
Fig. 2. Photograph of experimental setup.
lDEE ¼ 7.6
lDEE ¼ 9.8
BMEP
lBG
lT
BMEP
lBG
lT
BMEP
lBG
lT
2.27 1.73 1.30 0.86 0.54
4.74 6.64 9.30 15.27 20.0
2.66 3.21 3.75 4.49 4.84
2.80 2.27 1.73 1.29 0.86
3.75 4.53 5.26 6.58 8.23
2.52 2.86 3.15 3.60 4.05
4.54 4.33 3.89 3.35 2.80
2.17 2.27 2.45 2.51 2.84
1.77 1.84 1.96 2.01 2.21
K. Sudheesh, J.M. Mallikarjuna / Energy 35 (2010) 3614e3622
80
80
Pressure (bar)
60
BMEP = 2.27 Bar BMEP = 1.73 Bar BMEP = 1.29 Bar BMEP = 0.86 Bar BMEP = 0.54 Bar
Pressure (bar)
70
3617
50 40 Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 6.1
30 20
BMEP = 4.54 Bar 70 BMEP = 4.33 Bar BMEP = 3.89 Bar 60 BMEP = 3.35 Bar BMEP = 2.80 Bar 50 40 30 Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 9.8
20
10
10
0 340
350
360 Crank angle (degree)
370
0 350
380
Fig. 3. Variation of cylinder gas pressure with load for lDEE of 6.1.
360
370 Crank angle (degree)
380
390
Fig. 5. Variation of cylinder gas pressure with load for lDEE of 9.8.
the engine speed is allowed to reach about 1000 rev/min., by adjusting DEE flow rate. Then, biogas is inducted through intake manifold and its flow rate is adjusted to reach rated engine speed of 1500 rev/min. At this condition, the engine is allowed to run until the coolant water temperature reached 50 C. Afterwards, DEE flow rate is gradually reduced, because after the warm-up period, the engine knocks due to a higher DEE flow rate. At a given DEE flow rate, the possible operating load range is found out by varying biogas flow rate. In fact, the operating load range is decided by the misfiring and knocking limits of the engine. The above procedure is repeated with various DEE flow rates. Every time, the possible operating load range is determined by varying biogas flow rates based on the misfiring and knocking limits. Here, the knocking limit is considered as the rate of pressure rise of more than 10 bar/CAD. Finally, the combination of DEE and biogas flow rates for best brake thermal efficiency at each load condition is determined. Uncertainty and error analysis of the measured and calculated parameters is done, and the values are shown in the Appendix [20].
lBG ¼ lT ¼
Mair ; MBG *AFBG
lDEE ¼
Mair MDEE *AFDEE
Mair ðMBG *AFBG þ MDEE *AFDEE Þ
In the following discussion, biogas-DEE HCCI, biogas SI and biogasdiesel dual-fuel modes are referred to as HCCI, SI and dual-fuel modes, respectively. Here, the present results of biogas-DEE HCCI mode are compared with those of biogas SI mode [2] and biogasdiesel dual-fuel mode [3] at 25, 50, 75 and 100% of the maximum possible load (4.52 bar BMEP) of biogas-DEE HCCI mode. Where the exact values of parameters are not available, interpolated values are used for comparison. However, such comparison may not be very accurate because the composition of biogas used in each case is different. The composition of the biogas in biogas-diesel dual-fuel mode is 19% CO2 and 73% CH4 by volume [3]. In biogas SI mode [2] and in the present study, the composition of biogas is 41% CO2 and 58% CH4 by volume.
4. Results and discussion
4.1. Cylinder gas pressure
Both biogas and DEE flow rates affect combustion characteristics of biogas-DEE HCCI mode. In order to present their effect on engine performance and emissions, in the following discussion, biogas excess air ratio (lBG), DEE excess air ratio (lDEE) and total excess air ratios (lT) are defined as follows and used.
Table 4 shows various excess air ratios used to operate the engine in biogas-DEE HCCI mode under different load conditions. The DEE excess air ratios shown in Table 4 correspond to the DEE flow rates, which have the maximum number of best brake thermal efficiency points in the possible load range.
80
80
BMEP = 2.80 Bar BMEP = 2.27 Bar BMEP = 1.73 Bar BMEP = 1.29 Bar BMEP = 0.86 Bar
Pressure (bar)
60 50
60
40 30 20
0 330
340
350 360 370 Crank angle (degree)
380
Fig. 4. Variation of cylinder gas pressure with load for lDEE of 7.6.
50 40 30 20
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 7.6
10
DEE = 5.2 DEE = 6 1 DEE =7.6 DEE =8.5
70
Pressure (bar)
70
Cooling water temperature = 50 Celsius Load = 1.8 bar BMEP Engine speed = 1500 rev/min
10
390
0 330
340
350 360 370 Crank angle (degree)
380
Fig. 6. Variation of cylinder gas pressure with DEE flow rate.
390
3618
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Temperature (kelvin)
1400 1200
1600
BMEP = 2.27 Bar BMEP = 1.73 Bar BMEP = 1.29 Bar BMEP = 0.86 Bar BMEP = 0.54 Bar
1400 Temperature (kelvin)
1600
1000 800 600 400 200 0 320
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 6.1 340
360 Crank angle (degree)
380
1200 1000
BMEP = 4.54 Bar BMEP = 4.33 Bar BMEP = 3.89 Bar BMEP = 3.35 Bar BMEP = 2.80 Bar
800 600 Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 9.8
400 200 0 320
400
330
340
350 360 370 Crank angle (degree)
380
390
400
Fig. 7. Variation of cylinder gas temperature with load for lDEE of 6.1.
Fig. 9. Variation of cylinder gas temperature with loads for lDEE of 9.8.
Figs. 3e5 show variations of cylinder pressures with loads at different DEE excess air ratios. From Figs. 3e5, it can be seen that the occurrence of peak pressure advances with respect to the top dead center (TDC) with an increase in load. Also, the occurrence of peak pressure retards with an increase in DEE excess air ratio. At lower DEE excess air ratios, energy liberated from DEE is higher and thus autoignition occurs at an early stage. In general, peak pressure varies from about 51 to 72 bar for the entire load range considered. Fig. 6 shows the variation of cylinder pressure for various DEE excess air ratios at 1.8 bar BMEP. From Fig. 6, it is seen that at lower DEE excess air ratio, the occurrence of peak pressure advances and peak cylinder pressure increases due to the excess amount of DEE energy supply. This leads to an increased rate of pressure rise and engine noise, whereas, at high DEE excess air ratio, the cylinder pressure reduces and the occurrence of peak pressure retards due to a low DEE flow rate. It is observed that during experiments if lDEE increases above 8.5, the engine misfires. In this condition, the energy liberated from DEE may not be sufficient to autoignite the biogas.
lower DEE excess air ratio due to advanced combustion phasing. From Figs. 7e9, it can be seen that after 320 CAD, there is a slight increase in the cylinder gas temperature. It may be due to the low temperature reactions or cool flames (see Figs. 10e12). However, the cylinder gas temperature continues to increase even after the cool flames due to the compression of cylinder gases by upward movement of the piston. Occurrence of the peak cylinder gas temperature also follows a similar trend as that of the cylinder gas pressure. Peak cylinder gas temperatures are obviously higher at higher loads. In general, with a reduction in DEE excess air ratio, peak cylinder gas temperature increase due to advanced combustion phasing.
4.2. Cylinder gas temperature Figs. 7e9 show the variations of cylinder gas temperatures with loads at various DEE excess air ratios. Here, the cylinder gas temperature is calculated from the measured cylinder pressure and geometric volume at a given instant. The cylinder gas temperature follows a similar trend as that of the cylinder gas pressure. For a given load condition, the cylinder gas temperature is higher for
4.3. Heat release rate Figs. 10e12 show heat release rate patterns with respect to loads and various DEE excess air ratios. From Figs. 10e12, it is observed that with an increase in the DEE excess air ratio, the start of low temperature reactions (cool flames) is delayed and peaks of them are reduced. This is because of increased biogas flow rate, which reduces the reaction rate. It may be due to the higher dilution effect of CO2 present in biogas, which can even delay the start of the low temperature reactions. At high load conditions, the peak of cool flame reduces due to a reduction in the DEE flow rate. Generally, a low DEE flow rate is required for high load conditions because of higher overall wall temperatures. From Figs. 10e12, it is also seen that the peak of the main combustion phase increases and the combustion duration reduces with an increase in load. It may be 70
Temperature (kelvin)
1200 1000
BMEP = 2.80 Bar BMEP = 2.27 Bar BMEP = 1.73 Bar BMEP = 1.29 Bar BMEP = 0.86 Bar
60 Heat release rate (J/ CA)
1400
800 600 400 200 0 320
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 7.6
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 6.1
50 BMEP = 2.27 Bar BMEP = 1.73 Bar BMEP = 1.29 Bar BMEP = 0.86 Bar BMEP = 0.54 Bar
40 30 20 10 0
340
360 Crank angle (degree)
380
Fig. 8. Variation of cylinder gas temperature with loads for lDEE of 7.6.
400
-10
320
330
340
350
360
370
Crank angle (degree) Fig. 10. Patterns of heat release rate at various load for lDEE of 6.1.
380
K. Sudheesh, J.M. Mallikarjuna / Energy 35 (2010) 3614e3622
70
50 40
BMEP = 2.80 Bar BMEP = 2.27 Bar BMEP = 1.73 Bar BMEP = 1.29 Bar BMEP = 0.86 Bar
30 20 10 0 320
-10
330
340
350
Cooling water temperature = 50 Celsius Load = 1.8 bar BMEP Engine speed = 1500 rev/min
60 Heat release rate (J/ CA)
60 Heat release rate (J/ CA)
70
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 7.6
360
370
DEE = 5.2 DEE = 6.1 DEE = 7.6 DEE = 8.5
50 40 30 20 10 0
380
Crank angle (degree)
3619
-10
320
340
360
380
400
Crank angle (degree)
Fig. 11. Patterns of heat release rate at various loads for lDEE of 7.6.
Fig. 13. Patterns of heat release rate at various DEE flow rate.
due to lower total excess air ratio and higher wall temperatures at higher load conditions. The occurrence of peak heat release rate and the start of main combustion advance as the DEE excess air ratio reduces. This is because of higher energy release by DEE. Overall, for the entire load range, the maximum heat release rate is about 80 J/CAD. Fig. 13 shows the effect of DEE flow rate on heat release rate. From Fig. 13, it is seen that at a low DEE excess air ratio, combustion phasing advances due to the excess amount of energy supplied from DEE. This reduces the indicated power and leads to a lower brake thermal efficiency, whereas a very low DEE flow rate retards combustion phasing and increases combustion duration with lower cylinder pressures, and it again reduces the brake thermal efficiency. At 1.8 bar BMEP, lDEE of 6.1 gives the best brake thermal efficiency operating condition.
Fig. 14 shows a variation of the maximum rate of pressure rise with DEE excess air ratio and load conditions. At all the load conditions, the maximum rate of pressure rise is below 9 bar/CAD. It may be due to the presence of CO2 in the biogas, which has a good dilution effect. Also, the higher heat capacity of CO2, which acts as a diluent, reduces the rate of reactions leading to lower combustion rates. At all the best brake thermal efficiency points, the maximum rate of pressure rise is below 7.5 bar/CAD. It helps smooth engine operations without knocking, and also extends the operating load range.
70
Fig. 16 shows the percentage of energy supplied by DEE and biogas at various load conditions. From Fig. 16, it can be seen that at low loads, the energy supplied by the biogas is low. At low loads, biogas cannot autoignite easily due to a low overall engine temperature; therefore, a larger amount of DEE is required for autoignition. At 0.54 bar BMEP, the energy supplied by biogas is about 23.5% of total energy (at the best BTE point). As load increases, the energy contribution of biogas increases. At 4.52 bar BMEP, the energy contribution from biogas is around 82%. In Fig. 16, the dotted line shows energy share of the biogas at the best brake thermal efficiency points.
9 Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 9.8
60 50
Rate of pressure rise (bar/ oCA)
Heat release rate (J/ CA)
80
Fig. 15 shows the possible operating load range for various DEE excess air ratios. From Fig. 15, it is observed that as the DEE excess air ratio increases, the load limits are extended. It means that the operating load range shifts towards the right-hand side. At low load conditions, the engine temperature is low and therefore higher DEE flow rate is required to autoignite the biogas. But, as the load increases, the engine temperature increases, and, along with energy supplied by the DEE, leads to engine knock. This restricts upper load limit for a given DEE flow rate. In Fig. 15, the dotted line shows the best brake thermal efficiency DEE excess air ratio for the entire load range. The biogas-DEE HCCI mode has a wide load range of zero to 4.52 bar BMEP. 4.6. Energy share of biogas and DEE
4.4. Maximum rate of pressure rise
90
4.5. Operating load range
BMEP = 4.54 Bar BMEP = 4.33 Bar BMEP = 3.89 Bar BMEP = 3.35 Bar BMEP = 2.80 Bar
40 30 20 10 0 -10 320
340
360 Crank angle (degree)
380
Fig. 12. Patterns of heat release rate at various loads for lDEE of 9.8.
400
8 7 6
DEE = 5.2 DEE = 6.1 DEE = 7.6 DEE = 8.5 DEE = 9.8
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
5 4 3 2 1 0 0.0
1.0 2.0 3.0 4.0 Brake mean effective pressure (bar)
5.0
Fig. 14. Variation of maximum rate of pressure rise with DEE excess air ratio and load.
3620
K. Sudheesh, J.M. Mallikarjuna / Energy 35 (2010) 3614e3622
30
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
Brake thermal efficiency (%)
11
DEE air excess ratio
10 9 8 7 6 5 0.0
0.5
1.0 1.5 2.0 2.5 3.0 3.5 4.0 Brake mean effective pressure (bar)
4.5
Fig. 17 shows a variation of brake thermal efficiency (BTE) with load and DEE excess air ratios in HCCI mode. In Fig. 17, the dotted line represents the best BTE points for different load conditions in HCCI mode. From Fig. 17, it is observed that in HCCI mode, the best BTE point for each load condition depends upon mass flow rate of DEE. It may be because the optimum mass flow rates of DEE and biogas provide a better combustion phasing (Fig. 13). Fig. 18 shows the comparison of BTE in HCCI, SI [2] and dual-fuel [3] modes of operation. Only the best BTE points in HCCI mode are considered for comparison. In general, from Fig. 18, it is observed that the HCCI mode shows better BTE at all loads as compared to the other two modes. It may be because of full-throttle operation and higher compression ratio used in HCCI mode as compared to SI mode. Also, it may be due to the reduced combustion duration in HCCI mode as compared to that of dual-fuel mode, which is similar to conventional CI engine combustion. From Fig. 18, it is observed that at 25, 50, 75 and 100% loads, with HCCI mode, an improvement in BTE of about 29.5, 15.7, 19.3 and 9.2% respectively is observed as compared to that in SI mode. Further, at 50, 75 and 100% loads, an improvement of about 27.3, 8.81 and 3.48% respectively is observed as compared to the dual-fuel mode. 4.8. NO and smoke emissions Fig. 19 shows a variation of NO emissions with DEE excess air ratios and loads in HCCI mode. Fig. 20 shows a comparison of NO emissions in HCCI (at the best BTE points) and SI modes at four
15 10 5
DEE = 5.2 DEE = 6.1 DEE =7.6 DEE =8.5 DEE =9.8 MBTE points
60 50 40 30
0.5
1.0 1.5 2.0 2.5 3.0 3.5 4.0 Brake mean effective pressure (bar)
4.5
5.0
4.9. CO emissions Carbon monoxide (CO) emission is generally an indication of incomplete oxidation of fuel. It is found that gaseous fuelled HCCI engine shows lower CO emissions compared to liquid fuels due to higher miscibility of gaseous fuel with air [11]. Fig. 21 shows a variation of CO emission with DEE excess air ratio and loads for HCCI mode. At low load conditions, CO emission is higher may be due to low combustion temperatures. At higher loads, higher 30
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
70
DEE = 5.2 DEE = 6.1 DEE = 7.6 DEE = 8.5 DEE = 9.8 MBTE points
percentages of loads considered. From Fig. 20, it is observed that at all load conditions considered, NO emissions are very much lower in HCCI mode as compared to those of SI mode. This may be because of instantaneous and low temperature combustion in HCCI mode. Due to instantaneous combustion, the compression effect on burned gas by burning mixture may be eliminated. Thereby, local high temperature regions in burned gas products are eliminated, which, in turn, reduce the formation of thermal NO [21,22]. As there is no flame propagation in case of HCCI combustion, prompt NO formation may be negligible. From Fig. 20, it is found that at 25, 50, 75 and 100% loads, there is a reduction in NO emissions by about 93.61, 99.6, 97.2 and 99.7% respectively in HCCI mode as compared to SI mode. During the experiments, under all the operating conditions, it was observed that smoke emissions are very much lower (less than 0.1 BSU) in HCCI mode. It may be mainly due to the absence of fuelrich zones during HCCI combustion because of pre-mixed lean gaseous phase HCCI mode, unlike the conventional CI mode of operation.
Brake thermal efficiency (%)
Bio gas energy (%)
20
Fig. 17. Variation of brake thermal efficiency with load and DEE flow rate.
4.7. Brake thermal efficiency
80
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
0 0.0
5.0
Fig. 15. Possible operating load range with DEE excess air ratio.
90
25
20 10
25
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
20 15 HCCI 10
Porpatham et.al (SI,25% throttle)
5
Porpatham et.al (SI,100% throttle) Duc et.al (Dual fuel)
0 0.0
0.5
1.0 1.5 2.0 2.5 3.0 3.5 4.0 Brake mean effective pressure (bar)
Fig. 16. Energy share of biogas and DEE at various loads.
4.5
5.0
0 0.0
0.5
1.0
1.5 2.0 2.5 3.0 3.5 4.0 4.5 Brake mean effective pressure (bar)
5.0
5.5
Fig. 18. Comparison of brake thermal efficiency in HCCI, SI and dual-fuel modes.
K. Sudheesh, J.M. Mallikarjuna / Energy 35 (2010) 3614e3622
35
HCCI
Carbon monxides (% vol)
Nitric oxides (ppm)
2.5
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
30
DEE = 5.2 DEE = 6.1 DEE = 7.6 DEE = 8.5 DEE = 9.8 MBTE points
25 20 15
3621
10
Porpatham et. al ( SI, 25% throttle) Porpatham et. al ( SI,100% throttle)
2 1.5
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
1 0.5
5 0 0.0
0.5
1.0
1.5
2.0
2.5
3.0
3.5
4.0
4.5
0 0.0
5.0
0.5
Brake mean effective pressure (bar)
Nitric oxides (ppm)
800 600 400 200 0 25 50 75 100 Percentage of maximum possible HCCI load Fig. 20. Comparison of NO emission in HCCI and SI modes.
combustion temperatures may reduce the CO emissions comparatively. Fig. 22 shows a comparison of CO emission in HCCI and SI modes. From Fig. 22, it is seen that except at 2.7 bar BMEP, CO emission in HCCI mode is higher than that of SI mode. It may be due to low combustion temperature in HCCI mode, which may lead to flame quenching. At 25, 50, 75 and 100% loads, CO emissions are 4.2, 1.85, 1.73 and 0.64 times higher, respectively, than those of SI mode. At 2.7 bar BMEP, CO emission in SI mode is about 20 times
Fig. 23 shows a variation of hydrocarbon (HC) emissions with DEE excess air ratio and load. From Fig. 23, it can be seen that at low load conditions, HC emissions are comparatively lower because of the higher DEE flow rate. Due to this, the start of combustion advances (Fig. 10) and leads to comparatively higher cylinder gas temperatures (about 1100 to 1300 K with respect to loads as shown in Fig. 7), which improve DEE oxidization. Also, during the expansion stroke, the left out DEE can easily oxidize due to the high temperature of hot cylinder gas because of its low autoignition temperature. At intermediate loads, the HC emission increases, which may be due to an increase in the biogas flow rate with load. At higher biogas flow rates, more biogas could occupy crevices and would come out into the cylinder during expansion stroke, as the cylinder gas temperature is lower than the autoignition temperature of methane. This may inhibit oxidization of HC during expansion stroke, leading to higher HC emissions, whereas, at higher loads, a higher cylinder gas temperature helps oxidization of biogas and thereby the HC emissions are comparatively lower at higher loads. Fig. 24 shows a variation of HC emissions in HCCI and SI modes at four percentages of loads considered. In HCCI mode, at 25, 50 and 75% loads, the HC emissions are lower by about 83.9, 29.6 and
4500 Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 5.2 DEE = 6.1 DEE = 7.6 DEE = 8.5 DEE =9.8 MBTE points
0.40 0.30 0.20
Cooling water temperature = 50 Celsius
4000 Engine speed = 1500 rev/min Hydro carbons (ppm)
Carbon mooxides (% vol)
0.60 0.50
5.5
4.10. Hydrocarbon emissions
HCCI mode SI mode
1000
5.0
higher than that of HCCI mode. It may be due to the very rich mixture used at that load condition at 25% throttle in SI mode.
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
1200
1.5 2.0 2.5 3.0 3.5 4.0 4.5 Brake mean effective pressure (bar)
Fig. 22. Comparison of CO emission in HCCI and SI modes.
Fig. 19. Variation of NO emissions with DEE excess air ratio and load.
1400
1.0
0.10
3500 3000 2500
DEE = 5.2 DEE = 6.1 DEE = 7.6 DEE = 8.5 DEE =9.8 MBTE points
2000 1500 1000 500
0.00 0.0
0.5
1.0 1.5 2.0 2.5 3.0 3.5 4.0 Brake mean effective pressure (bar)
Fig. 21. Variation of CO with DEE excess air ratio and load.
4.5
5.0
0 0.0
0.5
1.0 1.5 2.0 2.5 3.0 3.5 4.0 Brake mean effective pressure (bar)
4.5
Fig. 23. Variation of HC emissions with DEE excess air ratio and load.
5.0
3622
K. Sudheesh, J.M. Mallikarjuna / Energy 35 (2010) 3614e3622
Appendix
9000 Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
Hydro carbons (ppm)
8000 7000
HCCI 6000
SI
5000 4000 3000 2000
Uncertainty and error analysis. Parameter
Uncertainty
Error of instrument
Speed Air flow rate
0.06% 0.90%
Fuel flow rate
0.40%
Intake charge temperature NO emission UBHC (FID) BMEP
0.06% 3.10% 2.10% 0.10%
Brake power
0.40%
Brake thermal efficiency
0.60%
1 rpm Volume ¼ 0.001 m3 Time ¼ 0.1 s Volume ¼ 0.0001 m3 Time ¼ 0.1 s Temperature ¼ 0.1 C 1 ppm 1 ppm Speed ¼ 1 rpm Torque ¼ 0.0485 nm Torque ¼ 0.0485 nm Speed ¼ 1 rpm Torque ¼ 0.0485 nm Speed ¼ 1 rpm Volume (fuel) ¼ 0.0001 m3
1000 0
25 50 75 100 Percentage of maximum possible HCCI load Fig. 24. Comparison of HC emissions in HCCI and SI modes.
45.57% respectively, as compared to SI mode. This may be due to lean mixtures in HCCI mode, unlike near stoichiometric mixtures in the SI mode. However, at 100% load, the HC emissions at both modes (HCCI and SI) are almost at equal levels. 5. Conclusions From the experimental investigations in biogas-DEE HCCI mode, the following conclusions are drawn: Biogas can be used more effectively in a HCCI mode when compared to other modes of operation. The engine can be easily started in HCCI mode by using DEE as an ignition improver unlike other fuels and methods used for HCCI operation. In this work, biogas-DEE HCCI mode of operation is possible in a load range of zero to 4.52 bar BMEP. The biogas-DEE HCCI mode shows higher BTE at all loads as compared to biogas-diesel dual-fuel and biogas SI modes. In addition, it is found that HC, NO and smoke emissions are lower in biogas-DEE HCCI mode as compared to biogas SI mode. However, CO emissions are higher at all loads. At 25, 50, 75 and 100% loads in biogas-DEE HCCI mode, BTE shows an improvement of 29.5, 15.7, 19.3 and 9.2% respectively as compared to SI mode. In addition, compared to dual-fuel mode, at 50, 75 and 100% loads, there is an improvement in BTE by about 27.3, 8.81 and 3.48% respectively. In biogas-DEE HCCI mode, at 25, 50 and 75% loads, HC emissions are reduced by about 83.9, 29.6, and 45.57% respectively as compared to the SI mode. At 100% load, HC emissions are comparable in both modes. In biogas-DEE HCCI mode, at 25, 50, 75 and 100% loads, CO emissions are 4.2, 1.85, 1.73 and 0.64 times respectively more than those of SI mode. The percentage reduction of NO emission in biogas-DEE HCCI mode at 25, 50, 75 and 100% loads are 93.61, 99.6, 97.2 and 99.7% respectively as compared to those in SI mode. At all load conditions, it is observed that, smoke emissions are very much lower (less than 0.1 BSU) in biogas-DEE HCCI mode. Therefore, it is feasible to operate biogas fuelled engine in HCCI mode for better performance and emission characteristics with DEE as an ignition improver compared to other modes of operation. Acknowledgments The authors wish to thank Suresh Kumar N., Cyril Mathew, Nagarajan K., Michael John Bose M., Subramanian M.K., and Babu S. for their help in carrying out the experiments. Their contribution is greatly appreciated.
References [1] Porpatham E. Studies on improving a biogas fuelled SI engine through charge and combustion chamber modification (PhD Thesis). Indian Institute of Technology Madras, Chennai, 2008. [2] Porpatham E, Ramesh A, Nagalingam B. Investigation on the effect of concentration of methane in biogas when used as a fuel for a spark ignition engine. Fuel 2008;87:1651e9. [3] Duc PM, Wattanavichien K. Study on biogas premixed charge diesel dual fuelled engine. Int J Energy Conversion and Management 2007;48:2286e308. [4] Bari S. Effect of carbon dioxide on the performance of biogas/diesel dual fuel engine. Int J Renewable Energy 1996;9(3):1007e110. [5] Baily B, Eberhardt J, Goguen S, Erwin J. Dietyl ether (DEE) as a renewable diesel fuel. SAE 1997; Paper No.972978. [6] Iranmanesh M, Subrahmanyam JP, Babu MKG. Potential of Diethyl Ether as supplementary fuel to improve combustion and emission characteristics of diesel engines. SAE 2008; Paper No. 2008-28-0044. [7] Onishi S, Jo SH, Shoda K, Jo PD, Kato S. Active thermo-atmospheric combustion (ATAC) e a new combustion process for internal combustion engines. SAE 1979; Paper No. 790501. [8] Najit PM, Foster DE. Compression-ignited homogeneous charge combustion. SAE 1983; Paper No. 830264. [9] Thring RH. Homogeneous charge compression ignition (HCCI) engines. SAE 1989; Paper No. 892068. [10] Ryan TW, Callahan TJ. Homogeneous charge compression ignition (HCCI) of diesel fuel. SAE 1996; Paper No. 961160. [11] Garcıa MT, Aguilar FJE, Lencero TS. Experimental study of the performances of a modified diesel engine operating in homogeneous charge compression ignition (HCCI) combustion mode versus the original diesel combustion mode. Energy 2009;34:159e71. [12] Shi L, Cui Y, Deng K, Peng H, Chen Y. Study of low emission homogeneous charge compression ignition (HCCI) engine using combined internal and external exhaust gas recirculation (EGR). Energy 2006;31:2665e76. [13] Christensen M, Einewall P, Johansson B. Homogeneous charge compression ignition (HCCI) using iso-octane, ethanol and natural gas e a comparison with spark ignition operation. SAE 1997; Paper No. 972874. [14] Swami Nathan S, Mallikarjuna JM, Ramesh A. HCCI engine operation with acetylene the fuel. SAE 2008; Paper No. 2008-28-0032. [15] Swami Nathan S, Mallikarjuna JM, Ramesh A. Effects of charge temperature and exhaust gas re-circulation on combustion and emission characteristics of an acetylene fuelled HCCI engine. Fuel 2010;89:515e21. [16] Chen Z, Konno M. Experimental study of natural e gas/DME homogeneous charge engine. SAE 2000; Paper No. 2000-01-0329. [17] Zheng Z, Yao M, Chen Z, Zhang B. Experimental study on HCCI combustion of dimethyl ether (DME)/methanol dual fuel. SAE 2004; Paper No. 2004-012993. [18] Mack JH, Aceves SM, Dibble RW. Demonstrating direct use of wet ethanol in a homogeneous charge compression ignition (HCCI) engine. Energy 2009;34: 782e7. [19] Swami Nathan S, Mallikarjuna JM, Ramesh A. Homogeneous charge compression ignition versus dual fuelling for utilizing biogas in compression ignition engines. Proceedings of the Institution of Mechanical Engineers, Part D: Journal of Automobile Engineering 2008;223(3):413e22. [20] Marangoni RD, Beckwith TG, Lienhard JH. Mechanical measurements. New York: Addison-Wesley; 1993. [21] Zhao H. HCCI and CAI engines for the automotive industry. Cambridge: Wood Head Publishing Limited; 2007. [22] Merker GP, Schwarz C, Stiesch G, Otto F. Simulating combustion. Heidelberg, Berlin: Springer; 2006.
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Diethyl ether as an ignition improver for biogas homogeneous charge compression ignition (HCCI) operation - An experimental investigation K. Sudheesh, J.M. Mallikarjuna* Internal Combustion Engines Laboratory, Department of Mechanical Engineering, Indian Institute of Technology Madras, Chennai 600 036, India
a r t i c l e i n f o
a b s t r a c t
Article history: Received 24 October 2009 Received in revised form 28 April 2010 Accepted 29 April 2010 Available online 14 June 2010
This paper deals with experimental investigations of a homogeneous charge compression ignition (HCCI) engine using biogas as a primary fuel and diethyl ether (DEE) as an ignition improver. The biogas is inducted and DEE is injected into a single-cylinder engine. For each load condition, best brake thermal efficiency DEE flow rate is determined. The results obtained in this study are also compared with those of the available biogas-diesel dual-fuel and biogas spark ignition (SI) modes. From the results, it is found that biogas-DEE HCCI mode shows wider operating load range and higher brake thermal efficiency (BTE) at all loads as compared to those of biogas-diesel dual-fuel and biogas SI modes. In HCCI mode, at 4.52 bar BMEP, as compared to dual-fuel and SI modes, BTE shows an improvement of about 3.48 and 9.21% respectively. Also, nitric oxide (NO) and smoke emissions are extremely low, and carbon monoxide (CO) emission is below 0.4% by volume at best brake thermal efficiency points. Also, in general, in HCCI mode, hydrocarbon (HC) emissions are lower than that of biogas SI mode. Therefore, it is beneficial to use biogas-DEE HCCI mode while using biogas in internal combustion engines. Ó 2010 Elsevier Ltd. All rights reserved.
Keywords: Biogas Diethyl ether HCCI Ignition improver
1. Introduction Generally, spark ignition (SI) and compression ignition (CI) engines are most commonly used for power generation and in transportation vehicles. The SI engines have low smoke emissions and brake thermal efficiency with high HC, CO and NOx emissions. However, the CI engines show comparatively higher brake thermal efficiency, but they emit high amounts of smoke and NOx. In these engines, achieving low exhaust emissions with low fuel consumption is a challenge. Also, today’s emission legislations are forcing engine manufacturers to search for engines having low exhaust emissions and higher fuel economy. Today, HCCI is emerging as an effective alternative combustion process for CI mode. With certain fuels, it can provide higher brake thermal efficiency like CI engines with ultra-low NOx and particulate matter (PM) emissions. In these engines, the premixed nature of a charge effectively eliminates the PM, whereas combustion with lean mixture without definite flame propagation reduces cylinder gas temperatures leading to ultra low NOx emissions simultaneously. Biogas is an attractive source of energy for rural areas especially in countries like India. It is generated by anaerobic digestion of cow dung, other animal wastes and plant matters such as leaves and
* Corresponding author. Tel.: þ91 44 22574698; fax: þ91 44 22574652. E-mail address: [email protected] (J.M. Mallikarjuna). 0360-5442/$ e see front matter Ó 2010 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2010.04.052
water hyacinth. All the above mentioned sources are renewable in nature and abundantly available. Biogas contains approximately two-thirds (by volume) of methane (CH4) and the rest is mostly carbon dioxide (CO2) with traces of hydrogen sulphide (H2S). Biogas has a low energy density due to the presence of CO2. The properties of biogas are given in Table 1. Usage of biogas in SI engines enhances knock resistance due to the presence of CO2. However, it reduces brake thermal efficiency and increases HC emissions [1,2]. It is not possible to use biogas directly in CI mode due to its higher autoignition temperature. Dual-fuel mode of operation is a feasible way of using biogas in CI engines with diesel as a secondary fuel [3,4]. But, dual-fuel mode emits more HC and CO emissions with lower brake thermal efficiency. However, smoke and NOx emissions are lower as compared to the conventional CI mode due to the gaseous nature of biogas and the presence of CO2. Diethyl ether is considered as a renewable fuel because it can be produced from ethanol through the dehydration process [5]. Low autoignition and boiling temperature of DEE are reasons for selecting it as an ignition improver along with lethargic fuels like biogas. Also, DEE has a higher energy density than ethanol. The important properties of DEE are given in Table 2. Generally, DEE is used as a cold starting aid in CI engines. DEE is also blended with diesel for improving the brake thermal efficiency and reducing emissions in CI engines [6]. Onishi et al. (1979) introduced controlled autoignition combustion in a two-stroke engine in order to reduce instability at
K. Sudheesh, J.M. Mallikarjuna / Energy 35 (2010) 3614e3622 Table 2 Properties of DEE.
Nomenclature
lBG lDEE lT Mair MBTE NOP MBG MDEE AFBG AFDEE CAD
Calorific value Density Boiling point Stoichiometric air fuel ratio (mass basis) Autoignition temperature Cetane number
biogas excess air ratio DEE excess air ratio total excess air ratios mass flow rate of the air best brake thermal efficiency nozzle opening pressure mass flow rate of the biogas mass flow rate of the DEE biogas stoichiometric airefuel ratio DEE stoichiometric airefuel ratio crank angle degree
part-loads and they also achieved good reduction of emissions and fuel consumption [7]. Najt et al. (1983) extended the HCCI combustion into a four-stroke engine using primary reference fuels [8]. A four-stroke gasoline HCCI engine was tested by Thring et al. (1989) and the important parameters required for successful gasoline HCCI operation at part-loads were investigated [9]. Ryan et al. (1996) used a port fuel injection (PFI) injector to supply diesel into the intake air stream at various inlet air temperatures and compression ratios [10]. This resulted in early heat release during compression stroke itself, and they concluded that low compression ratio is most suitable for port injected diesel fuelled HCCI engine. Garcıa et al. (2009) investigated the effect of inlet charge temperature, cool EGR, injection timings and equivalence ratio on the performance of a diesel fuelled HCCI engine and achieved comparatively higher loads by using cool EGR [11]. Shi et al. (2006) studied the effect of internal and externally cooled EGR on the performance of a diesel fuelled HCCI engine. They concluded that internal EGR benefited the formation of homogeneous mixture and reduced smoke emission, whereas externally cooled EGR could help extend upper load limit of HCCI operation [12]. Due to the early heat release characteristics with low volatility of diesel, researchers tried to use various alternative fuels with high autoignition temperature, viz. natural gas, methanol, ethanol, liquefied petroleum gas (LPG) without many modifications to the original engine. Inlet charge heating, variable compression ratio (VCR) and use of secondary low octane fuels as an ignition improver were methods tried for achieving HCCI combustion with the above fuels. Christensen et al. (1997) investigated the HCCI combustion characteristics using natural gas with inlet air heating [13]. Swami Nathan et al. (2008) adopted acetylene as a fuel for HCCI engine because of its moderate autoignition temperature and high flammability limits. They used inlet charge temperature to control combustion phasing [14,15]. Chen et al. (2000) introduced dual-fuel HCCI combustion of natural gas with dimethyl either (DME). They
Table 1 Properties and composition of biogas. Calorific value Density (1 atm and 15 C) Flame speed Stoichiometric air fuel ratio (mass basis) Autoignition temperature Flammability limits with air (%) Research octane number Typical biogas composition in % volume Methane Carbon dioxide Carbon monoxide
3615
17 MJ/kg 1.2 kg/m3 0.25 m/s 5.7 650 C 7.5e14 130 57.37 42.1 0.08
33.9 MJ/kg 713 kg/m3 34.4 C 11.1 160 C >125
found that by optimizing a proportion of DME and natural gas, NOx emissions could be lowered to near zero levels. The dual-fuel operation gave higher brake thermal efficiency than that of CI mode [16]. Zheng et al. (2004) used DME as an ignition controller in a methanol fuelled HCCI engine [17]. Mack et al. (2009) experimentally proved the suitability of wet ethanol in HCCI combustion. They used inlet charge heating to control the combustion phasing [18]. Swami Nathan et al. (2008) investigated biogas HCCI combustion with manifold injection of diesel. They compared the results of biogas-diesel HCCI mode with that of biogas-diesel duelfuel mode of operation. Due to the low volatility and high boiling temperature of diesel, inlet heating was used for proper mixing of diesel with manifold inducted biogas. They concluded that biogasdiesel HCCI operation is superior to dual-fuel mode of operation in a BMEP range of 2.5 to 4 bar [19]. However, they couldn’t operate an engine in HCCI mode below 2.5 bar and above 4 bar BMEPs due to system limitations and difficulty in controlling inlet charge heating. The motivation for the present work is to operate a single-cylinder engine in biogas fuelled HCCI mode in a wide load range by using DEE as an ignition improver. These types of studies are limited in literature. Therefore, a detailed study would help not only to evaluate and compare the HCCI mode with conventional biogas SI and biogas diesel-dual-fuel modes, but also to find the feasibility of using biogas in HCCI mode. In this study, the effect of DEE flow rate on the performance, emissions and operating load range of a biogas fuelled HCCI mode are studied at a constant engine speed of 1500 rev/min., and cooling water outlet temperature of 50 C. Finally, DEE mass flow rate for best brake thermal efficiency point at each load condition has been found out. In addition, the performance and emission characteristics of biogas-DEE HCCI mode are compared with the available results of biogas SI [1,2] and biogas-diesel dual-fuel [3] modes. 2. Experimental setup A single-cylinder, water-cooled, direct injection CI engine is used for conducting experiments. The engine is coupled to an eddy current dynamometer for loading and measurement purposes. The engine specifications are shown in Table 3. The DEE is stored in an accumulator and is injected into the intake manifold using an injector at a line pressure of 2 bar. An in-house built electronic circuit is used for controlling DEE flow rate. The biogas is directly inducted through the intake manifold. The biogas is generated in a nearby plant, which uses cow dung and water to produce it. It is collected in a flexible bag at the plant and transported to the place of usage. It is actually sent through a floating drum in order to maintain the Table 3 Engine specifications. Bore stroke Connecting rod length Compression ratio Rated power output Displacement volume Injector NOP
80 110 mm 231 mm 16:1 3.7 kW @ 1500 rpm 553 cm3 220 bar
3616
K. Sudheesh, J.M. Mallikarjuna / Energy 35 (2010) 3614e3622
1. Air flow measurement
13. Cooling water inlet
2. DEE supply unit
14. Cooling water flow regulator
3. DEE pressure gauge
15. Cooling water out let temperature sensor
4. DEE pulse width controller
16. Cooling water outlet
5. DEE fuel injector
17. Engine
6. Biogas flow meter
18. Exhaust gas temperature sensor
7. Biogas flow control valve
19. Eddy current dynamometer
8. Biogas floating drum
20. Electromagnetic clutch
9. Pressure transducer
21. Electric motor
10. Crank angle encoder
22 to 25. Exhaust gas analyzers
11. Data acquisition system
26. Engine exhaust
12. Personal computer Fig. 1. Schematic of experimental setup.
required constant pressure as in the plant. The biogas composition is measured every time using a MRU make non-dispersive infrared (NDRI) analyzer meant for this purpose. A typical composition of the biogas is shown in Table 1. The flow rate of biogas is controlled by a fine-control valve and is measured with the help of a turbine type
gas flow meter before being admitted to intake manifold. The cooling water outlet temperature is measured by a resistance temperature detector (RTD) and the air flow rate is measured by using a dry turbine type flow meter. Exhaust gas analyzers working on the principles of FID for HC, NDIR for CO and CLD for NO, respectively, are used to measure the emission levels. A Bosch smoke meter is used to measure smoke emissions. Figs. 1 and 2 show schematic and photographic views respectively of the experimental setup developed and used in this study. 3. Experimental procedure First, the engine is motored at about a speed of 600 rev/min., and simultaneously DEE is injected into the intake manifold. Thus, the engine is started in HCCI mode with DEE as the only fuel and
Table 4 Mixture strengths used in biogas-DEE HCCI mode.
lDEE ¼ 6.1
Fig. 2. Photograph of experimental setup.
lDEE ¼ 7.6
lDEE ¼ 9.8
BMEP
lBG
lT
BMEP
lBG
lT
BMEP
lBG
lT
2.27 1.73 1.30 0.86 0.54
4.74 6.64 9.30 15.27 20.0
2.66 3.21 3.75 4.49 4.84
2.80 2.27 1.73 1.29 0.86
3.75 4.53 5.26 6.58 8.23
2.52 2.86 3.15 3.60 4.05
4.54 4.33 3.89 3.35 2.80
2.17 2.27 2.45 2.51 2.84
1.77 1.84 1.96 2.01 2.21
K. Sudheesh, J.M. Mallikarjuna / Energy 35 (2010) 3614e3622
80
80
Pressure (bar)
60
BMEP = 2.27 Bar BMEP = 1.73 Bar BMEP = 1.29 Bar BMEP = 0.86 Bar BMEP = 0.54 Bar
Pressure (bar)
70
3617
50 40 Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 6.1
30 20
BMEP = 4.54 Bar 70 BMEP = 4.33 Bar BMEP = 3.89 Bar 60 BMEP = 3.35 Bar BMEP = 2.80 Bar 50 40 30 Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 9.8
20
10
10
0 340
350
360 Crank angle (degree)
370
0 350
380
Fig. 3. Variation of cylinder gas pressure with load for lDEE of 6.1.
360
370 Crank angle (degree)
380
390
Fig. 5. Variation of cylinder gas pressure with load for lDEE of 9.8.
the engine speed is allowed to reach about 1000 rev/min., by adjusting DEE flow rate. Then, biogas is inducted through intake manifold and its flow rate is adjusted to reach rated engine speed of 1500 rev/min. At this condition, the engine is allowed to run until the coolant water temperature reached 50 C. Afterwards, DEE flow rate is gradually reduced, because after the warm-up period, the engine knocks due to a higher DEE flow rate. At a given DEE flow rate, the possible operating load range is found out by varying biogas flow rate. In fact, the operating load range is decided by the misfiring and knocking limits of the engine. The above procedure is repeated with various DEE flow rates. Every time, the possible operating load range is determined by varying biogas flow rates based on the misfiring and knocking limits. Here, the knocking limit is considered as the rate of pressure rise of more than 10 bar/CAD. Finally, the combination of DEE and biogas flow rates for best brake thermal efficiency at each load condition is determined. Uncertainty and error analysis of the measured and calculated parameters is done, and the values are shown in the Appendix [20].
lBG ¼ lT ¼
Mair ; MBG *AFBG
lDEE ¼
Mair MDEE *AFDEE
Mair ðMBG *AFBG þ MDEE *AFDEE Þ
In the following discussion, biogas-DEE HCCI, biogas SI and biogasdiesel dual-fuel modes are referred to as HCCI, SI and dual-fuel modes, respectively. Here, the present results of biogas-DEE HCCI mode are compared with those of biogas SI mode [2] and biogasdiesel dual-fuel mode [3] at 25, 50, 75 and 100% of the maximum possible load (4.52 bar BMEP) of biogas-DEE HCCI mode. Where the exact values of parameters are not available, interpolated values are used for comparison. However, such comparison may not be very accurate because the composition of biogas used in each case is different. The composition of the biogas in biogas-diesel dual-fuel mode is 19% CO2 and 73% CH4 by volume [3]. In biogas SI mode [2] and in the present study, the composition of biogas is 41% CO2 and 58% CH4 by volume.
4. Results and discussion
4.1. Cylinder gas pressure
Both biogas and DEE flow rates affect combustion characteristics of biogas-DEE HCCI mode. In order to present their effect on engine performance and emissions, in the following discussion, biogas excess air ratio (lBG), DEE excess air ratio (lDEE) and total excess air ratios (lT) are defined as follows and used.
Table 4 shows various excess air ratios used to operate the engine in biogas-DEE HCCI mode under different load conditions. The DEE excess air ratios shown in Table 4 correspond to the DEE flow rates, which have the maximum number of best brake thermal efficiency points in the possible load range.
80
80
BMEP = 2.80 Bar BMEP = 2.27 Bar BMEP = 1.73 Bar BMEP = 1.29 Bar BMEP = 0.86 Bar
Pressure (bar)
60 50
60
40 30 20
0 330
340
350 360 370 Crank angle (degree)
380
Fig. 4. Variation of cylinder gas pressure with load for lDEE of 7.6.
50 40 30 20
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 7.6
10
DEE = 5.2 DEE = 6 1 DEE =7.6 DEE =8.5
70
Pressure (bar)
70
Cooling water temperature = 50 Celsius Load = 1.8 bar BMEP Engine speed = 1500 rev/min
10
390
0 330
340
350 360 370 Crank angle (degree)
380
Fig. 6. Variation of cylinder gas pressure with DEE flow rate.
390
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Temperature (kelvin)
1400 1200
1600
BMEP = 2.27 Bar BMEP = 1.73 Bar BMEP = 1.29 Bar BMEP = 0.86 Bar BMEP = 0.54 Bar
1400 Temperature (kelvin)
1600
1000 800 600 400 200 0 320
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 6.1 340
360 Crank angle (degree)
380
1200 1000
BMEP = 4.54 Bar BMEP = 4.33 Bar BMEP = 3.89 Bar BMEP = 3.35 Bar BMEP = 2.80 Bar
800 600 Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 9.8
400 200 0 320
400
330
340
350 360 370 Crank angle (degree)
380
390
400
Fig. 7. Variation of cylinder gas temperature with load for lDEE of 6.1.
Fig. 9. Variation of cylinder gas temperature with loads for lDEE of 9.8.
Figs. 3e5 show variations of cylinder pressures with loads at different DEE excess air ratios. From Figs. 3e5, it can be seen that the occurrence of peak pressure advances with respect to the top dead center (TDC) with an increase in load. Also, the occurrence of peak pressure retards with an increase in DEE excess air ratio. At lower DEE excess air ratios, energy liberated from DEE is higher and thus autoignition occurs at an early stage. In general, peak pressure varies from about 51 to 72 bar for the entire load range considered. Fig. 6 shows the variation of cylinder pressure for various DEE excess air ratios at 1.8 bar BMEP. From Fig. 6, it is seen that at lower DEE excess air ratio, the occurrence of peak pressure advances and peak cylinder pressure increases due to the excess amount of DEE energy supply. This leads to an increased rate of pressure rise and engine noise, whereas, at high DEE excess air ratio, the cylinder pressure reduces and the occurrence of peak pressure retards due to a low DEE flow rate. It is observed that during experiments if lDEE increases above 8.5, the engine misfires. In this condition, the energy liberated from DEE may not be sufficient to autoignite the biogas.
lower DEE excess air ratio due to advanced combustion phasing. From Figs. 7e9, it can be seen that after 320 CAD, there is a slight increase in the cylinder gas temperature. It may be due to the low temperature reactions or cool flames (see Figs. 10e12). However, the cylinder gas temperature continues to increase even after the cool flames due to the compression of cylinder gases by upward movement of the piston. Occurrence of the peak cylinder gas temperature also follows a similar trend as that of the cylinder gas pressure. Peak cylinder gas temperatures are obviously higher at higher loads. In general, with a reduction in DEE excess air ratio, peak cylinder gas temperature increase due to advanced combustion phasing.
4.2. Cylinder gas temperature Figs. 7e9 show the variations of cylinder gas temperatures with loads at various DEE excess air ratios. Here, the cylinder gas temperature is calculated from the measured cylinder pressure and geometric volume at a given instant. The cylinder gas temperature follows a similar trend as that of the cylinder gas pressure. For a given load condition, the cylinder gas temperature is higher for
4.3. Heat release rate Figs. 10e12 show heat release rate patterns with respect to loads and various DEE excess air ratios. From Figs. 10e12, it is observed that with an increase in the DEE excess air ratio, the start of low temperature reactions (cool flames) is delayed and peaks of them are reduced. This is because of increased biogas flow rate, which reduces the reaction rate. It may be due to the higher dilution effect of CO2 present in biogas, which can even delay the start of the low temperature reactions. At high load conditions, the peak of cool flame reduces due to a reduction in the DEE flow rate. Generally, a low DEE flow rate is required for high load conditions because of higher overall wall temperatures. From Figs. 10e12, it is also seen that the peak of the main combustion phase increases and the combustion duration reduces with an increase in load. It may be 70
Temperature (kelvin)
1200 1000
BMEP = 2.80 Bar BMEP = 2.27 Bar BMEP = 1.73 Bar BMEP = 1.29 Bar BMEP = 0.86 Bar
60 Heat release rate (J/ CA)
1400
800 600 400 200 0 320
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 7.6
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 6.1
50 BMEP = 2.27 Bar BMEP = 1.73 Bar BMEP = 1.29 Bar BMEP = 0.86 Bar BMEP = 0.54 Bar
40 30 20 10 0
340
360 Crank angle (degree)
380
Fig. 8. Variation of cylinder gas temperature with loads for lDEE of 7.6.
400
-10
320
330
340
350
360
370
Crank angle (degree) Fig. 10. Patterns of heat release rate at various load for lDEE of 6.1.
380
K. Sudheesh, J.M. Mallikarjuna / Energy 35 (2010) 3614e3622
70
50 40
BMEP = 2.80 Bar BMEP = 2.27 Bar BMEP = 1.73 Bar BMEP = 1.29 Bar BMEP = 0.86 Bar
30 20 10 0 320
-10
330
340
350
Cooling water temperature = 50 Celsius Load = 1.8 bar BMEP Engine speed = 1500 rev/min
60 Heat release rate (J/ CA)
60 Heat release rate (J/ CA)
70
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 7.6
360
370
DEE = 5.2 DEE = 6.1 DEE = 7.6 DEE = 8.5
50 40 30 20 10 0
380
Crank angle (degree)
3619
-10
320
340
360
380
400
Crank angle (degree)
Fig. 11. Patterns of heat release rate at various loads for lDEE of 7.6.
Fig. 13. Patterns of heat release rate at various DEE flow rate.
due to lower total excess air ratio and higher wall temperatures at higher load conditions. The occurrence of peak heat release rate and the start of main combustion advance as the DEE excess air ratio reduces. This is because of higher energy release by DEE. Overall, for the entire load range, the maximum heat release rate is about 80 J/CAD. Fig. 13 shows the effect of DEE flow rate on heat release rate. From Fig. 13, it is seen that at a low DEE excess air ratio, combustion phasing advances due to the excess amount of energy supplied from DEE. This reduces the indicated power and leads to a lower brake thermal efficiency, whereas a very low DEE flow rate retards combustion phasing and increases combustion duration with lower cylinder pressures, and it again reduces the brake thermal efficiency. At 1.8 bar BMEP, lDEE of 6.1 gives the best brake thermal efficiency operating condition.
Fig. 14 shows a variation of the maximum rate of pressure rise with DEE excess air ratio and load conditions. At all the load conditions, the maximum rate of pressure rise is below 9 bar/CAD. It may be due to the presence of CO2 in the biogas, which has a good dilution effect. Also, the higher heat capacity of CO2, which acts as a diluent, reduces the rate of reactions leading to lower combustion rates. At all the best brake thermal efficiency points, the maximum rate of pressure rise is below 7.5 bar/CAD. It helps smooth engine operations without knocking, and also extends the operating load range.
70
Fig. 16 shows the percentage of energy supplied by DEE and biogas at various load conditions. From Fig. 16, it can be seen that at low loads, the energy supplied by the biogas is low. At low loads, biogas cannot autoignite easily due to a low overall engine temperature; therefore, a larger amount of DEE is required for autoignition. At 0.54 bar BMEP, the energy supplied by biogas is about 23.5% of total energy (at the best BTE point). As load increases, the energy contribution of biogas increases. At 4.52 bar BMEP, the energy contribution from biogas is around 82%. In Fig. 16, the dotted line shows energy share of the biogas at the best brake thermal efficiency points.
9 Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 9.8
60 50
Rate of pressure rise (bar/ oCA)
Heat release rate (J/ CA)
80
Fig. 15 shows the possible operating load range for various DEE excess air ratios. From Fig. 15, it is observed that as the DEE excess air ratio increases, the load limits are extended. It means that the operating load range shifts towards the right-hand side. At low load conditions, the engine temperature is low and therefore higher DEE flow rate is required to autoignite the biogas. But, as the load increases, the engine temperature increases, and, along with energy supplied by the DEE, leads to engine knock. This restricts upper load limit for a given DEE flow rate. In Fig. 15, the dotted line shows the best brake thermal efficiency DEE excess air ratio for the entire load range. The biogas-DEE HCCI mode has a wide load range of zero to 4.52 bar BMEP. 4.6. Energy share of biogas and DEE
4.4. Maximum rate of pressure rise
90
4.5. Operating load range
BMEP = 4.54 Bar BMEP = 4.33 Bar BMEP = 3.89 Bar BMEP = 3.35 Bar BMEP = 2.80 Bar
40 30 20 10 0 -10 320
340
360 Crank angle (degree)
380
Fig. 12. Patterns of heat release rate at various loads for lDEE of 9.8.
400
8 7 6
DEE = 5.2 DEE = 6.1 DEE = 7.6 DEE = 8.5 DEE = 9.8
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
5 4 3 2 1 0 0.0
1.0 2.0 3.0 4.0 Brake mean effective pressure (bar)
5.0
Fig. 14. Variation of maximum rate of pressure rise with DEE excess air ratio and load.
3620
K. Sudheesh, J.M. Mallikarjuna / Energy 35 (2010) 3614e3622
30
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
Brake thermal efficiency (%)
11
DEE air excess ratio
10 9 8 7 6 5 0.0
0.5
1.0 1.5 2.0 2.5 3.0 3.5 4.0 Brake mean effective pressure (bar)
4.5
Fig. 17 shows a variation of brake thermal efficiency (BTE) with load and DEE excess air ratios in HCCI mode. In Fig. 17, the dotted line represents the best BTE points for different load conditions in HCCI mode. From Fig. 17, it is observed that in HCCI mode, the best BTE point for each load condition depends upon mass flow rate of DEE. It may be because the optimum mass flow rates of DEE and biogas provide a better combustion phasing (Fig. 13). Fig. 18 shows the comparison of BTE in HCCI, SI [2] and dual-fuel [3] modes of operation. Only the best BTE points in HCCI mode are considered for comparison. In general, from Fig. 18, it is observed that the HCCI mode shows better BTE at all loads as compared to the other two modes. It may be because of full-throttle operation and higher compression ratio used in HCCI mode as compared to SI mode. Also, it may be due to the reduced combustion duration in HCCI mode as compared to that of dual-fuel mode, which is similar to conventional CI engine combustion. From Fig. 18, it is observed that at 25, 50, 75 and 100% loads, with HCCI mode, an improvement in BTE of about 29.5, 15.7, 19.3 and 9.2% respectively is observed as compared to that in SI mode. Further, at 50, 75 and 100% loads, an improvement of about 27.3, 8.81 and 3.48% respectively is observed as compared to the dual-fuel mode. 4.8. NO and smoke emissions Fig. 19 shows a variation of NO emissions with DEE excess air ratios and loads in HCCI mode. Fig. 20 shows a comparison of NO emissions in HCCI (at the best BTE points) and SI modes at four
15 10 5
DEE = 5.2 DEE = 6.1 DEE =7.6 DEE =8.5 DEE =9.8 MBTE points
60 50 40 30
0.5
1.0 1.5 2.0 2.5 3.0 3.5 4.0 Brake mean effective pressure (bar)
4.5
5.0
4.9. CO emissions Carbon monoxide (CO) emission is generally an indication of incomplete oxidation of fuel. It is found that gaseous fuelled HCCI engine shows lower CO emissions compared to liquid fuels due to higher miscibility of gaseous fuel with air [11]. Fig. 21 shows a variation of CO emission with DEE excess air ratio and loads for HCCI mode. At low load conditions, CO emission is higher may be due to low combustion temperatures. At higher loads, higher 30
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
70
DEE = 5.2 DEE = 6.1 DEE = 7.6 DEE = 8.5 DEE = 9.8 MBTE points
percentages of loads considered. From Fig. 20, it is observed that at all load conditions considered, NO emissions are very much lower in HCCI mode as compared to those of SI mode. This may be because of instantaneous and low temperature combustion in HCCI mode. Due to instantaneous combustion, the compression effect on burned gas by burning mixture may be eliminated. Thereby, local high temperature regions in burned gas products are eliminated, which, in turn, reduce the formation of thermal NO [21,22]. As there is no flame propagation in case of HCCI combustion, prompt NO formation may be negligible. From Fig. 20, it is found that at 25, 50, 75 and 100% loads, there is a reduction in NO emissions by about 93.61, 99.6, 97.2 and 99.7% respectively in HCCI mode as compared to SI mode. During the experiments, under all the operating conditions, it was observed that smoke emissions are very much lower (less than 0.1 BSU) in HCCI mode. It may be mainly due to the absence of fuelrich zones during HCCI combustion because of pre-mixed lean gaseous phase HCCI mode, unlike the conventional CI mode of operation.
Brake thermal efficiency (%)
Bio gas energy (%)
20
Fig. 17. Variation of brake thermal efficiency with load and DEE flow rate.
4.7. Brake thermal efficiency
80
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
0 0.0
5.0
Fig. 15. Possible operating load range with DEE excess air ratio.
90
25
20 10
25
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
20 15 HCCI 10
Porpatham et.al (SI,25% throttle)
5
Porpatham et.al (SI,100% throttle) Duc et.al (Dual fuel)
0 0.0
0.5
1.0 1.5 2.0 2.5 3.0 3.5 4.0 Brake mean effective pressure (bar)
Fig. 16. Energy share of biogas and DEE at various loads.
4.5
5.0
0 0.0
0.5
1.0
1.5 2.0 2.5 3.0 3.5 4.0 4.5 Brake mean effective pressure (bar)
5.0
5.5
Fig. 18. Comparison of brake thermal efficiency in HCCI, SI and dual-fuel modes.
K. Sudheesh, J.M. Mallikarjuna / Energy 35 (2010) 3614e3622
35
HCCI
Carbon monxides (% vol)
Nitric oxides (ppm)
2.5
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
30
DEE = 5.2 DEE = 6.1 DEE = 7.6 DEE = 8.5 DEE = 9.8 MBTE points
25 20 15
3621
10
Porpatham et. al ( SI, 25% throttle) Porpatham et. al ( SI,100% throttle)
2 1.5
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
1 0.5
5 0 0.0
0.5
1.0
1.5
2.0
2.5
3.0
3.5
4.0
4.5
0 0.0
5.0
0.5
Brake mean effective pressure (bar)
Nitric oxides (ppm)
800 600 400 200 0 25 50 75 100 Percentage of maximum possible HCCI load Fig. 20. Comparison of NO emission in HCCI and SI modes.
combustion temperatures may reduce the CO emissions comparatively. Fig. 22 shows a comparison of CO emission in HCCI and SI modes. From Fig. 22, it is seen that except at 2.7 bar BMEP, CO emission in HCCI mode is higher than that of SI mode. It may be due to low combustion temperature in HCCI mode, which may lead to flame quenching. At 25, 50, 75 and 100% loads, CO emissions are 4.2, 1.85, 1.73 and 0.64 times higher, respectively, than those of SI mode. At 2.7 bar BMEP, CO emission in SI mode is about 20 times
Fig. 23 shows a variation of hydrocarbon (HC) emissions with DEE excess air ratio and load. From Fig. 23, it can be seen that at low load conditions, HC emissions are comparatively lower because of the higher DEE flow rate. Due to this, the start of combustion advances (Fig. 10) and leads to comparatively higher cylinder gas temperatures (about 1100 to 1300 K with respect to loads as shown in Fig. 7), which improve DEE oxidization. Also, during the expansion stroke, the left out DEE can easily oxidize due to the high temperature of hot cylinder gas because of its low autoignition temperature. At intermediate loads, the HC emission increases, which may be due to an increase in the biogas flow rate with load. At higher biogas flow rates, more biogas could occupy crevices and would come out into the cylinder during expansion stroke, as the cylinder gas temperature is lower than the autoignition temperature of methane. This may inhibit oxidization of HC during expansion stroke, leading to higher HC emissions, whereas, at higher loads, a higher cylinder gas temperature helps oxidization of biogas and thereby the HC emissions are comparatively lower at higher loads. Fig. 24 shows a variation of HC emissions in HCCI and SI modes at four percentages of loads considered. In HCCI mode, at 25, 50 and 75% loads, the HC emissions are lower by about 83.9, 29.6 and
4500 Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min DEE = 5.2 DEE = 6.1 DEE = 7.6 DEE = 8.5 DEE =9.8 MBTE points
0.40 0.30 0.20
Cooling water temperature = 50 Celsius
4000 Engine speed = 1500 rev/min Hydro carbons (ppm)
Carbon mooxides (% vol)
0.60 0.50
5.5
4.10. Hydrocarbon emissions
HCCI mode SI mode
1000
5.0
higher than that of HCCI mode. It may be due to the very rich mixture used at that load condition at 25% throttle in SI mode.
Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
1200
1.5 2.0 2.5 3.0 3.5 4.0 4.5 Brake mean effective pressure (bar)
Fig. 22. Comparison of CO emission in HCCI and SI modes.
Fig. 19. Variation of NO emissions with DEE excess air ratio and load.
1400
1.0
0.10
3500 3000 2500
DEE = 5.2 DEE = 6.1 DEE = 7.6 DEE = 8.5 DEE =9.8 MBTE points
2000 1500 1000 500
0.00 0.0
0.5
1.0 1.5 2.0 2.5 3.0 3.5 4.0 Brake mean effective pressure (bar)
Fig. 21. Variation of CO with DEE excess air ratio and load.
4.5
5.0
0 0.0
0.5
1.0 1.5 2.0 2.5 3.0 3.5 4.0 Brake mean effective pressure (bar)
4.5
Fig. 23. Variation of HC emissions with DEE excess air ratio and load.
5.0
3622
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Appendix
9000 Cooling water temperature = 50 Celsius Engine speed = 1500 rev/min
Hydro carbons (ppm)
8000 7000
HCCI 6000
SI
5000 4000 3000 2000
Uncertainty and error analysis. Parameter
Uncertainty
Error of instrument
Speed Air flow rate
0.06% 0.90%
Fuel flow rate
0.40%
Intake charge temperature NO emission UBHC (FID) BMEP
0.06% 3.10% 2.10% 0.10%
Brake power
0.40%
Brake thermal efficiency
0.60%
1 rpm Volume ¼ 0.001 m3 Time ¼ 0.1 s Volume ¼ 0.0001 m3 Time ¼ 0.1 s Temperature ¼ 0.1 C 1 ppm 1 ppm Speed ¼ 1 rpm Torque ¼ 0.0485 nm Torque ¼ 0.0485 nm Speed ¼ 1 rpm Torque ¼ 0.0485 nm Speed ¼ 1 rpm Volume (fuel) ¼ 0.0001 m3
1000 0
25 50 75 100 Percentage of maximum possible HCCI load Fig. 24. Comparison of HC emissions in HCCI and SI modes.
45.57% respectively, as compared to SI mode. This may be due to lean mixtures in HCCI mode, unlike near stoichiometric mixtures in the SI mode. However, at 100% load, the HC emissions at both modes (HCCI and SI) are almost at equal levels. 5. Conclusions From the experimental investigations in biogas-DEE HCCI mode, the following conclusions are drawn: Biogas can be used more effectively in a HCCI mode when compared to other modes of operation. The engine can be easily started in HCCI mode by using DEE as an ignition improver unlike other fuels and methods used for HCCI operation. In this work, biogas-DEE HCCI mode of operation is possible in a load range of zero to 4.52 bar BMEP. The biogas-DEE HCCI mode shows higher BTE at all loads as compared to biogas-diesel dual-fuel and biogas SI modes. In addition, it is found that HC, NO and smoke emissions are lower in biogas-DEE HCCI mode as compared to biogas SI mode. However, CO emissions are higher at all loads. At 25, 50, 75 and 100% loads in biogas-DEE HCCI mode, BTE shows an improvement of 29.5, 15.7, 19.3 and 9.2% respectively as compared to SI mode. In addition, compared to dual-fuel mode, at 50, 75 and 100% loads, there is an improvement in BTE by about 27.3, 8.81 and 3.48% respectively. In biogas-DEE HCCI mode, at 25, 50 and 75% loads, HC emissions are reduced by about 83.9, 29.6, and 45.57% respectively as compared to the SI mode. At 100% load, HC emissions are comparable in both modes. In biogas-DEE HCCI mode, at 25, 50, 75 and 100% loads, CO emissions are 4.2, 1.85, 1.73 and 0.64 times respectively more than those of SI mode. The percentage reduction of NO emission in biogas-DEE HCCI mode at 25, 50, 75 and 100% loads are 93.61, 99.6, 97.2 and 99.7% respectively as compared to those in SI mode. At all load conditions, it is observed that, smoke emissions are very much lower (less than 0.1 BSU) in biogas-DEE HCCI mode. Therefore, it is feasible to operate biogas fuelled engine in HCCI mode for better performance and emission characteristics with DEE as an ignition improver compared to other modes of operation. Acknowledgments The authors wish to thank Suresh Kumar N., Cyril Mathew, Nagarajan K., Michael John Bose M., Subramanian M.K., and Babu S. for their help in carrying out the experiments. Their contribution is greatly appreciated.
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