Control of combustion process in an HCCI-DI combustion engine using dual injection strategy with EGR

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Control of combustion process in an HCCI-DI combustion engine using dual injection strategy with EGR

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Pranab Das ⇑, P.M.V. Subbarao, J.P. Subrahmanyam

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Department of Mechanical Engineering, Indian Institute of Technology, Delhi, India

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h i g h l i g h t s  HCCI-DI combustion phasing can be controlled by varying premixed ratio and EGR.

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 Peak pressure varies within a narrow range of 2 deg (5–7 deg BTDC).

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 Peak HRR varies within a narrow range of 3 deg (1–4 deg BTDC).  A reduction of 76% NOX and 40% smoke emissions are achieved.

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a r t i c l e 2 3 0 3 21 22 23 24 25 26 27 28 29 30 31 32

i n f o

Article history: Received 20 February 2015 Received in revised form 17 May 2015 Accepted 2 July 2015 Available online xxxx Keywords: Homogeneous charge compression ignition Dual injection Premixed equivalence ratio Premixed ratio Two phase heat release

a b s t r a c t The objective of this study was to achieve a controlled HCCI-DI combustion using in-cylinder dual injection technique and to study the effect of premixed ratio and EGR on combustion, performance and emission characteristics of the engine. Studies were carried out up to 67% (0–67%) of the full load and a constant speed of 1500 rev/min. The premixed ratio was varied from zero to a maximum of 80% of the total fuel by mass. At 50% load condition, occurrences of peak pressure, peak PRR and peak HRR varied within 5–7 deg BTDC, 1.5–5 deg BTDC and 1–4 deg BTDC respectively. It was observed that diesel homogeneous combustion showed a two phase heat release pattern at higher premixed ratios and all combustion parameters advanced dramatically with increasing premixed ratio. The results showed that a reduction of 76% in NOX and 40% in smoke opacity were achieved using 80% premixed ratio with 30% EGR. As the premixed ratio was increased, there was an improvement in IMEP, ISFC and smoke opacity with penalties in HC and CO emissions. Ó 2015 Elsevier Ltd. All rights reserved.

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1. Introduction

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Homogeneous Charge Compression Ignition (HCCI) combustion is a promising alternative to existing internal combustion engine operating modes and combines the benefits of conventional SI and CI engines [1]. Combustion occurs simultaneously throughout the whole combustion chamber in an HCCI combustion engine unlike conventional CI or SI engines. Different types of combustion systems have been proposed by several researchers; nevertheless all these systems demonstrate the auto-ignition characteristics of a

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Abbreviations: ATDC, after top dead centre; BDC, bottom dead centre; BTDC, before top dead centre; CI, compression ignition; EGR, exhaust gas recirculation; IMEP, indicated mean effective pressure; ISFC, indicated specific fuel consumption; ITE, brake thermal efficiency; PRR, pressure rise rate; RPM, revolutions per minute; SI, spark ignition; TDC, top dead centre. ⇑ Corresponding author. Tel.: +91 955970910/8800267347. E-mail address: [email protected] (P. Das).

homogeneous mixture. The HCCI combustion concept was first introduced by Onishi et al. [2] and Noguchi et al. [3] named as Active Thermo-Atmospheric Combustion (ATAC) and Toyota-Sokhen (TS) combustion respectively. Najt and Foster [4] investigated the Compression-Ignited Homogeneous Charge (CIHC) combustion in a 4-stroke engine and reported a very low cyclic variation and lower HC emissions at low load conditions. A lot of advancements have been achieved in gasoline HCCI combustion engine. To name a few Lund Institute of Technology [5,6]; Shell Global Solution [7–9]; GM R&D [4,10]; ERC, Wisconsin University [11,12]; and Sloan Automotive research lab, MIT [13] are the most significant. Now a day, several US research lab mostly funded by U.S. Department of Energy (U.S. DOE) are closely involved in HCCI combustion research [14]. For instance Argonne National Lab (ANL) [11,12], Sandia National Lab (SNL) [15–17], Lawrence Barkley National Lab [LBNL] and Oak Ridge National Lab (ORNL) are among the most significant. All of the

http://dx.doi.org/10.1016/j.fuel.2015.07.009 0016-2361/Ó 2015 Elsevier Ltd. All rights reserved.

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Nomenclature rp Fp

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premixed ratio premixed equivalence ratio

above mentioned research lab are mainly focused on Gasoline fuelled HCCI combustion engine. ‘‘Lund Institute of Technology’’ is one of the pioneers in HCCI combustion technology where gasoline like fuels were used mainly because, highly volatile fuel helps in forming homogeneous mixtures, in addition, higher auto ignition temperature of gasoline like (high octane) fuel helps in extending the high load range of HCCI combustion engine [6]. This is the main advantage of gasoline like fuel over diesel like fuel when HCCI combustion is practised. At the same time achieving auto ignition with gasoline like fuel at low load is extremely difficult which results in misfire and thereby high cycle by cycle variation. Kalghatgi [9] proposed stratification controlled HCCI combustion which can achieve auto ignition even at very low load. He further suggested stratification can help in controlling auto ignition across various load range and hence combustion phasing can be controlled. Recent studies in Sandia National lab [15–17] and Delphi Automotive [18] have shown notable achievements in extending the load range of gasoline HCCI combustion engine using stratification technique. Stratification technique has become a key combustion control strategy for extending the load range of HCCI combustion [18,19]. Ciatti and Subramanian [20] in Argonne National Lab has shown that low octane fuel, stratification and EGR are extremely important to achieve controlled LTC combustion. In a parallel way, stratification controlled diesel fuelled HCCI combustion concept are practised widely among automotive industries and academia all across the globe. Stratification technique has evolved as the most promising technology which is expected to dominate in near future. Toyota and Nissan Motors have already commercialised diesel fuelled HCCI like combustion concept in Japanese market. Delphi Power train used this stratification technique for achieving HCCI combustion with diesel as fuel [21]. Studies [22–24] have reported that fuel consumptions worsened while running on HCCI mode using diesel fuel, which remained one of the major issues of such HCCI engines. Kimura et al. [25] demonstrated the Modulate Kinetics (MK) combustion concept and ultra low emissions of NOX and smoke were reported. One of the most successful HCCI combustion technologies [26] using dual injection strategy was successfully implemented by Toyota Motors where early pilot injection was used to form a homogeneous mixture followed by a main injection to trigger the combustion. A detailed review [27] on diesel fuelled HCCI combustion provided the insight and importance of various operating parameters in such combustion systems. Homogeneous charge preparation itself is a major issue in diesel fuelled HCCI combustion engines. Thring. [28] showed that a port injection system could be used to prepare the homogeneous mixture but intake air heating and high EGR would be required to achieve HCCI combustion. Another major concern of HCCI combustion is knocking. This limits the high load operation of HCCI combustion engines [29]. To overcome these difficulties switching to conventional combustion during high load operation is an accepted option for HCCI combustion engines [29]. An interesting observation suggested that transition from conventional to premixed charge compression ignition (PCCI) during high load operation was possible by varying the fuel injection timing [30]. Ganesh et al. [31,32] reported that using external mixture formation they were able to achieve HCCI combustion with ultra low

Mp Mm

mass of the pilot or premixed fuel mass of the main fuel

emissions. The Single early injected HCCI combustion [33] would have no control over the identifiable start of combustion and this is recognised as one of the major problems in achieving controlled HCCI combustion. HCCI combustion is governed by operating parameters like intake pressure, temperature, fuel air equivalence ratio, fuel type, engine speed and residual gas fraction. In other words, it is majorly controlled by chemical kinetics. Several studies [34–37] have been reported on dual and multi injection strategies as a means to achieve controlled homogeneous combustion. Double and multi-level injection strategies to achieve HCCI combustion in diesel engines are being investigated these days. An early pilot injection forms a homogeneous mixture and late injection near TDC controls the combustion phasing. It is realised that split injection timing and split injection ratio play significant roles in controlling combustion phasing of HCCI combustion. Several other control strategies have been proposed to control the combustion phasing of an HCCI combustion engine. A few strategies are EGR [37–39], controlling intake air temperature [40,41], reduced CR, retarded main injection timing and retarded single injection timing [42]. Some researchers [43,44] investigated a wide range of early single injection timings and studied their effect on HCCI combustion. They observed that as the mixing time was increased, combustion phasing was retarded thereby reducing NOX and PM simultaneously but advanced injection also worsened the fuel consumption. Ra et al. [44] proposed that low pressure injection system was better as higher injection pressure would tend to increase the spray tip penetration resulting in wall wetting. Thus a low pressure injection system has been recommended as this would reduce the cost of the HCCI engine. Effect of split ratio on combustion and emission characteristics of HCCI-DI combustion engine using a novel dual injection strategy was presented in an earlier work. These studies showed that a two phase heat release pattern appeared in HCCI-DI combustion when low octane fuel was used [45,46]. From the literature, it is seen that pilot and main injection strategy is one of the most promising HCCI combustion phasing control technique. It is well understood that a second injection or main injection would certainly initiate the combustion and at the same time misfire can be avoided for low load HCCI operation. For HCCI combustion, controlling combustion phasing is extremely difficult. In the recent past, stratification technique has become very popular due to its inherent merits. The prediction of occurrences of peak pressure or the control of combustion phasing is not yet developed as combustion phasing is highly sensitive to variation in operating conditions while running on HCCI combustion mode. The available literature is limited to in-cylinder stratified controlled HCCI combustion. The present study is focused on the roles of premixed ratio and EGR on combustion phasing, performance and emission behaviour of a diesel fuelled HCCI combustion engine using dual in-cylinder injection. Dual injection is used to achieve homogeneous combustion where an early (pilot) injection forms a premixed homogeneous charge and a second (main) injection near the TDC is intended to trigger the combustion and thus a controlled combustion phasing is obtained. Present study covers both extreme low load range and the limiting high load range. Strategies have been identified and discussed how this load range can be extended further. This study also presents the effect of

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cooled external EGR which is very important to extend the load range of HCCI-DI combustion.

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2. Experimental apparatus

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A four stroke single cylinder diesel engine was modified to run on HCCI mode by providing dual injection. The specifications of the engine are given in Table 1. The detailed experimentation technique has been described in [45,46]. To acquire the cylinder pressure data in order to study combustion parameters, a piezoelectric pressure transducer was mounted on the cylinder head. The charge output was converted to give a voltage signal using a charge amplifier and the signal was recorded and stored in a computer for post processing of the data with the help of a compatible software package, LabVIEWÒ. A crank angle rotary encoder was used to get the crank angle corresponding to the cylinder pressure data. Pressure signals of 100 consecutive cycles were averaged to study the combustion parameters of the engine. Table 2 gives the test conditions of the experiments studied. K-type thermocouples were used to measure exhaust, intake, outer cylinder wall and lubricating oil temperatures during the experiments. A multi gas analyser was used to measure engine CO, HC and NOX Emissions. A smoke meter was used to measure smoke opacity of the engine exhaust. Fig. 1 shows the schematic diagram of the experimental setup. Fig. 2 gives the photograph of the experimental test rig. Table 3 gives the properties of the fuel used during experiment. Table 4 gives the measurement technique and accuracy of different instruments used for the experiments. Uncertainty analysis of different parameters were carried out and these are given in Table 4.

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3. Experimental procedure

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The engine was first started in the conventional single injection mode until the temperatures of lubricating oil, cylinder wall and intake air were stabilised. Baseline performance data was obtained by running the engine using conventional single injection. To study the HCCI combustion, dual injection strategy was used where the first injection (pilot) was made early in the suction stroke followed by a second injection (main) near the TDC during the compression stroke. Pilot fuel forms a homogeneous mixture owing to sufficient mixing time whereas main fuel starts the combustion by providing a locally rich mixture near the TDC. Thus the main injection event plays a significant role in controlling combustion phasing in HCCI-DI engine. Because of dual injection strategy, this method of HCCI is also referred to as partial HCCI combustion. The partial HCCI strategy can be used to attain the benefits of homogeneous combustion and also the controlled combustion phasing by the main injection event. Experiments were conducted for (zero) 0–67% of the full load conditions and at a constant speed of 1500 rev/min. Premixed ratio was varied from zero (conventional) to a maximum permissible premixed ratio which is limited by a very early combustion phasing and knocking. Premixed ratio (rp) is defined as the ratio of mass of pilot fuel to the total fuel and is given in Eq. (1):

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Table 1 Test engine specifications. Engine type Piston type Bore  stroke (mm) Displacement (cc) Max. power (kW) Compression ratio

Single cylinder 4-stroke diesel Bowl type 116  102 950 7.4 @ 1500 rpm 17.5

Table 2 Test conditions. Fuel used Injection pressure (bar)

Commercial diesel 200

Injection timing Pilot inj. timing (BTDC) Main inj. timing (BTDC) Premixed ratio (%) EGR (%) Engine speed (RPM) Load range

180° 20° 0, 20, 40, 60 and 80 15 and 30 1500 0–67% of full load

Premixed ratio ðrp Þ ¼ Mp =ðMp þ Mm Þ

ð1Þ

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Here the maximum premixed ratio is kept at 80% of the total fuel. This 80% premixed ratio corresponds to a premixed equivalence ratio of (Fp) 0.41. It is observed that the maximum threshold premixed ratio is different for different loads and for very low load, knocking does not appear even for 100% premixed ratio but it would cause rough running of the engine which is due to very lean mixture and no specific control over the start of combustion. Nevertheless, even for low load conditions combustion phasing advances with the increasing premixed ratio similar to the other operating loads. Premixed ratio is increased slowly starting from 0% premixed ratio which corresponds to the conventional single injection and subsequently 20%, 40%, 60% and 80% premixed ratios were studied. The main objective of this study was to achieve the maximum permissible premixed ratio (80%) while maintaining a safe running of the engine. The effect of EGR was studied only for the 80% premixed case, where 15% and 30% EGR rates were used. When the pilot fuel quantity was increased, combustion parameters such as SOC, PRR and HRR showed an advancing trend. The retarded injection timing of 20 deg BTDC was set for main injection [47] and 180 deg BTDC was set for pilot injection. As the pilot injection quantity was continuously increased beyond a threshold value, the engine started knocking due to a very early combustion phasing and thus all the experiments were carried out at a retarded main injection timing of 20 deg BTDC instead of conventional injection timing.

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4. Results and discussion

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In this section, analysis of combustion parameters, performance and emission behaviour at various premixed ratio is discussed for 50% load condition. Low load (0–33%) and high load (67%) are discussed in Appendix.

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4.1. Combustion analysis at baseline in diesel mode

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The combustion characteristics of the conventional engine were first obtained. Fig. 3 shows the combustion parameters of the engine at its original injection timing of 26 deg. BTDC. The engine was run at 50% load and at 1500 rev/min. The angle of occurrences of different combustion parameters were obtained in order to compare with the premixed combustion mode.

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4.2. Combustion analysis of premixed fuel mode

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Fig. 4 shows the effect of increasing premixed ratio on in-cylinder pressure and pressure rise rate. Here, premixed ratio was varied keeping load and speed constant. It is seen that increased premixed fuel results in higher peak pressure and advanced combustion by reducing ignition delay due to higher premixed combustion of the fuel. It is observed that with the higher premixed ratio, peak pressure increased and advanced with respect to crank angle. It is also observed that rate of pressure rise (PRR)

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Fig. 1. Schematic diagram of the experimental test rig.

Table 4 Accuracy of instruments and uncertainty of measured parameters.

Fig. 2. Photograph of the experimental test rig.

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advances with respect to crank angle as premixed ratio is increased. This is because, at higher premixed ratios, a higher proportion of the fuel burns in the homogeneous combustion mode which results in higher cylinder pressure and higher PRR. It is also seen in the figure that as the premixed ratio is increased, peak pressure increases and more importantly the occurrences of peak pressures are advanced. This is because the pilot fuel undergoes pre-flame chemical reactions and thereby causes the peak pressure to advance. At 80% premixed ratio, the peak cylinder pressure is much higher than that under baseline conditions thereby increasing NOX emissions and limits the high load range of HCCI combustion. To reduce the peak pressure, EGR was introduced for 80% premixed combustion. Fig. 5 shows the effect of EGR on cylinder gas pressure and PRR for HCCI combustion. As seen in the figure, use

Table 3 Fuel properties. Density (kg/m3) Viscosity Flash point (°C) Pour point (°C) Cloud point (°C) Calorific value (MJ/kg) Cetane number

838 2.5 @ 40 °C (cst) 76 35 5 42.4 MJ/kg 45–55

Sl. no.

Measurements

Instrument type

Accuracy

Uncertainty (%)

1 2 3 4 5 6 7 8

Fuel consumption Time Temperature Pressure CO emission NOX emission HC emission Smoke opacity

Glass burette Stop watch Thermocouple Piezoelectric NDIR Electrochemical NDIR Light absorption

±0.1 cc ±0.1 s ±1 °C ±0.5% ±0.02% ±25 ppm ±4 ppm ±1%

±0.7 ±0.5 ±0.5 ±0.2 ±1.3 ±2.4 ±1.7 ±1.8

of EGR can significantly reduce the cylinder pressure and retard the combustion phasing of HCCI combustion. Thus, a very low peak pressure and delayed combustion phasing with 80% premixed ratio was achieved. This is because as the EGR is introduced at 40 °C, oxygen availability of the combustion chamber reduced which in turn slows down the reaction rate so that all combustion parameters keep moving away from TDC and the peak cylinder pressure sharply falls. Fig. 6 shows the peak pressure values at different premixed ratios. As the premixed ratio increases, peak pressure increases. For the 80% premixed ratio, peak pressure is nearly 3 bar higher than the base line cylinder pressure. It is observed that for 80% premixed condition, 15% EGR brought the peak pressure down to near the level corresponding to baseline condition while 30% EGR brought the peak pressure to nearly 5 bar less than the baseline value. A further increase in EGR generated unacceptable levels of smoke and hence maximum EGR was limited to 30% in the present work. Fig. 7 shows the angle of occurrences of peak pressure at different premixed ratios. As the premixed ratio is increased, the occurrence of peak pressure starts advancing. At baseline, occurrence of peak pressure is about 5.5 deg after TDC whereas by retarding the baseline injection timing to 20 deg BTDC, peak pressure occurred at 6.5 deg ATDC. It is seen that as the premixed ratio is increased, peak pressure starts shifting towards TDC. At 80% premixed ratio the peak pressure occurs nearly 5 deg ATDC. When EGR is introduced, the occurrences of peak pressure starts shifting away from TDC and for 30% EGR, peak pressure occurs at 7 deg ATDC. Fig. 8 shows the variation of peak PRR at different premixed ratios. It is seen that peak PRR increases with increasing premixed ratio and for 80% premixed ratio, peak PRR is higher than that at the baseline value. At 15% EGR the peak PRR reduces by about 8%, whereas 30% EGR reduces the peak PRR by 20.7%.

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Fig. 3. Variation of baseline cylinder pressure, heat release rate and pressure rise rate with respect to crank angle at 50% load and 1500 rev/min.

Fig. 4. Effect of premixed ratio on pressure rise rate with respect to crank angle.

Fig. 5. Effect of EGR on cylinder pressure and pressure rise rate for HCCI combustion with respect to crank angle with 80% premixed ratio.

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Fig. 9 shows the angle of occurrence of peak PRR at different premixed ratios. It is seen that the peak PRR occurs at about 1.5 deg BTDC for the retarded single injection combustion (0% pre-mixed) which is due to retarded combustion phasing and all combustion parameters move away from TDC with increasing premixed ratio. As the premixed ratio was increased the occurrences of peak PRR advanced and for a premixed ratio of 80%, peak PRR comes close to the baseline angle of peak PRR as seen in the figure.

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It is observed that use of 15% and 30% EGR retard the occurrence of peak PRR by 2 deg and 3.5 deg respectively. Fig. 10 shows the effect of increasing premixed ratio on heat release rate. It is observed that with the higher premixed ratio, combustion phasing and the heat release rate advances. The higher premixed fuel undergoes two phase combustion characteristics where the first HRR peak shows the low temperature combustion followed by a high temperature heat release which is one of the main features of HCCI combustion. A small hump in the heat release rate is observed at nearly 25 deg BTDC. This is often termed as ‘‘cool flame heat release’’ which is a strong function of premixed fuel air equivalence ratio and therefore as the premixed equivalence ratio reduces this cool flame heat release reduces and for a very lean premixed equivalence ratio; cool flame diminishes and the HRR curve resembles more like conventional diesel combustion. As the premixed ratio is increased, cool flame heat release starts increasing as seen in the figure. Another noteworthy observation is that for higher premixed ratio the amount of cool flame heat release is higher which results in advancement of the occurrences of combustion parameters. The energy release during this cool flame phase actually defines the occurrences of start of combustion (SOC). The more the energy released during the cool flame period, the more the advancement of SOC. The figure clearly shows that as the premixed ratio is increased, the amount of cool flame heat release increases resulting in an advancement of SOC and subsequently all combustion parameters start advancing. This is because a higher amount of heat releases during cool flame phase which forms a higher number of active radicals that accelerate the combustion so that all combustion parameters start advancing. Fig. 11 shows the effect of EGR on HCCI combustion. It is observed that as the EGR rate is increased, cool flame heat release reduces and ignition delay increases which in turn retards the combustion phasing. Thus EGR plays a very important role in retarding the combustion phasing of HCCI combustion. A detailed analysis on the occurrences of peak HRR and SOC is presented in Figs. 12 and 13. These plots also give a comparison with the baseline combustion characteristics. Fig. 12 shows the angle at which peak HRR occurs at different premixed ratios. As seen in the figure, the peak HRR significantly advances with increased premixed ratio. The same trend of reduced ignition delay is observed for both the conditions causing all other combustion parameters to occur in advance which is an issue that needs to be studied further. For base line combustion, peak HRR occurs at 4 deg BTDC whereas by retarding the baseline injection timing to 20 deg BTDC, peak HRR occurs at 1 deg BTDC and as the premixed ratio is increased the occurrence of peak HRR starts to advance. When 80% premixed ratio is employed, peak HRR occurs at 4 deg BTDC which is the same as that at baseline It is also observed that as the EGR is introduced, the occurrence of peak HRR starts retarding and for 30% EGR, the peak HRR retards to 3 deg BTDC. Without EGR the peak HRR occurs at 1 deg BTDC. Fig. 13 shows the angle of occurrence of SOC at different premixed ratios. It is observed that baseline SOC occurs at 10 deg BTDC and as the main injection timing is retarded, SOC occurs at 7 deg BTDC. A clear trend of advancement of SOC with respect to crank angle is seen as shown in the figure. SOC advances with the increasing premixed ratio. For 80% premixed ratio, SOC occurs at 12 deg BTDC. This is due to the same reason as discussed for other combustion parameters. As the EGR is introduced with HCCI combustion, SOC starts retarding with increasing EGR. Fig. 14 gives the effect of premixed ratio on average gas temperature and cumulative heat release. As seen in the figure, the gas temperature increases as the premixed ratio is increased. All the numerical values of combustion parameters obtained from the above study are given in Table 5.

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Fig. 6. Effect of premixed ratio on peak cylinder pressure. Fig. 9. Effect of premixed ratio on the occurrences of peak PRR.

Fig. 7. Effect of premixed ratio on the occurrences of peak pressure.

Fig. 10. Effect of premixed ratio on cylinder pressure and heat release rate with respect to crank angle.

Fig. 8. Effect of premixed ratio on peak pressure rise rate (PRR).

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4.3. Performance analysis

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The parameters considered for performance analysis are IMEP, ISFC and exhaust gas temperature. Fig. 15a gives the effect of premixed ratio on IMEP. It is observed that the IMEP at 0% premixed ratio (injection timing 20 deg BTDC) is much lower than baseline (injection timing 26 deg BTDC). As premixed ratio is increased, IMEP increases consistently and at 80% premixed ratio the IMEP reaches to the baseline value. This indicates that higher premixed ratio will improve the performance of the engine by increasing indicated thermal efficiency. Considering the best performance of the engine without

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overloading, it is important to reduce peak cylinder pressure in order to reduce emission. Thus 15% and 30% EGR was introduced with 80% premixed ratio condition. When 15% EGR is applied, there is a slight fall (1.4%) in IMEP is observed where as a sharp 8.5% fall in IMEP is observed when 30% EGR is applied. Fig. 15b shows the variation of COV of IMEP with premixed ratio. It is observed that at all test points, COV of IMEP is below 3.5%. This signifies cycle by cycle variation at 50% load is very low or in other words combustion is extremely stable. Fig. 16 shows the variation in ISFC with respect to premixed ratio. Figure shows a decreasing trend of ISFC as premixed ratio is increased. This is probably because more complete combustion takes place at higher premixed ratio resulting in higher IMEP as seen in Fig. 15 and thus lower ISFC is observed. It is seen that, when injection timing was retarded to 20 deg BTDC from baseline injection timing of 26 deg BTDC, there is a rise in ISFC for 0% premixed combustion. This is because, retarded single injection results in retardation of all combustion parameters and therefore more fuel is needed to maintain the same power output of the engine which results in higher ISFC. It is also seen in Fig. 16 that EGR results in increase in ISFC. Fig. 17 shows the variation of exhaust temperature at different premixed ratios. Exhaust temperature shows a reducing trend as the premixed ratio is increased. This is because of the advanced combustion phasing and reduced diffusion combustion that occurs

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homogeneous charge combustion results in lower smoke formation. It is also observed that at low load condition smoke opacity remained more or less constant at all premixed ratios. This is because at low load, fuel air equivalence ratio is very low resulting in lower smoke emissions for the entire range of premixed ratio. At medium load, higher premixed ratio plays a significant role in abating smoke emissions. Fig. 19 shows the variation in NOX emissions with premixed ratio. At intermediate load it is observed that NOX emissions slightly increase with the increased premixed ratio. The peak temperature increases with increase in premixed ratio because of early heat release of premixed fuel and it becomes significant at higher load as equivalence ratio of premixed fuel increases. It is seen that when 60% premixed ratio is employed, NOX emissions start increasing when compared to retarded single injection combustion. At 80% premixed ratio NOX emissions were close to baseline emissions. As a result, EGR was introduced to reduce NOX emissions. When 15% and 30% EGR was introduced at 80% premixed ratio, a sharp drop of 43% and 76% NOX emissions respectively were observed. Beyond 30% EGR, smoke opacity increased significantly although NOX emissions reduced. It is also observed that at low load NOX emissions remained almost constant even at 80% premixed ratio. This is because the corresponding premixed equivalence ratio is very lean resulting in lower peak cylinder temperature. Fig. 20 shows the variation in CO emissions with premixed ratio. It is observed that CO emissions increase with increase in premixed ratio. The homogeneous charge near to the cylinder wall burns at a lower temperature which reduces the oxidation process which results in increase in CO emission with increase in premixed ratio. Another reason could be partly because there would be less time available during the combustion process. This characteristic has been observed by several researchers and identified as a

Fig. 11. Effect of EGR on cylinder pressure and heat release rate for HCCI (80% premixed) combustion with respect to crank angle.

Fig. 12. Effect of premixed ratio on the occurrences of peak heat release rate.

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in higher premixed combustion. As the main fuel quantity reduces with increasing premixed ratio, diffusion combustion of the main fuel reduces and therefore, exhaust gas temperature reduces. When EGR is introduced, exhaust gas temperature increases because EGR retards the combustion phasing.

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4.4. Emission analysis

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The main aim of HCCI combustion is to achieve low NOX and smoke. It is essential to select the parameters which generate lower emissions without much loss in efficiency. Fig. 18 shows the variation in smoke opacity with premixed ratio. It is seen that increased premixed ratio results in a drastic reduction in smoke opacity. This is to be expected because

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Fig. 13. Effect of premixed ratio on Start of Combustion (SOC).

Table 5 Calculated combustion parameters. Premixed ratio

Peak pressure (bar)

Peak pr. angle (ATDC)

Peak PRR (bar/ CA)

Peak PRR angle (BTDC)

Peak HRR (J/ CA)

Peak HRR angle (BTDC)

SOC (BTDC)

IMEP (bar)

0% 20% 40% 60% 80% (HCCI) HCCI with 15% EGR HCCI with 30% EGR Baseline

55.02 55.71 57.17 61.68 64.82 61.87

6.5 7.0 7.0 6.5 5.0 6.5

3.23 3.37 3.39 3.84 4.14 3.81

1.5 1.5 2.0 2.5 5.0 3

53.55 56.12 55.32 61.78 59.40 59.91

1.0 1.0 1.0 1.5 4.0 2

7.0 8.0 9.5 10.5 12.0 11

4.73 5.21 5.26 5.52 5.72 5.64

56.54

7.0

3.28

1.5

54.59

1.0

9.0

5.23

62.43

5.5

3.90

5.0

55.02

4.0

10.0

5.71

Please cite this article in press as: Das P et al. Control of combustion process in an HCCI-DI combustion engine using dual injection strategy with EGR. Fuel (2015), http://dx.doi.org/10.1016/j.fuel.2015.07.009

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Increasing premixed rao

Fig. 16. Effect of premixed ratio on indicated specific fuel consumption (ISFC). Fig. 14. Effect of premixed ratio on cumulative HRR and average gas temperature.

Fig. 15a. Effect of premixed ratio on net indicated mean effective pressure (IMEP).

Fig. 17. Effect of premixed ratio on exhaust gas temperature.

Fig. 15b. Effect of premixed ratio on COV of IMEP.

Fig. 18. Effect of premixed ratio on smoke opacity. 491 492 493 494 495 496 497 498

penalty of homogeneous combustion. It is also observed that use of EGR further increases CO emissions as expected. This is because higher EGR results in reducing the peak temperature and also reduced the oxygen availability inside the combustion chamber. As a result CO emission increases with increasing EGR. Fig. 21 shows the variation in HC emission with premixed ratio. It is seen that higher premixed ratio results in higher unburnt hydrocarbon emissions because a higher fraction of fuel enters

the crevice regions and remains unburnt. Thus HC emissions will increase as premixed ratio increases. It is also observed that as the EGR is introduced, HC emissions further increases. This is due to the same reason as discussed in Fig. 20, higher EGR reduces peak temperature and therefore oxidation process slows down resulting in higher HC formation.

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2. Two phase heat release pattern is observed at higher premixed ratios. At low premixed ratios, fuel air premixed equivalence ratio is very lean. Thus premixed fuel does not undergo any noticeable low temperature heat release or two phase heat release. 3. EGR plays a significant role in reducing the peak cylinder pressure and retarding the combustion phasing of HCCI combustion. For higher premixed ratio, use of EGR is necessary to reduce the peak pressure which would reduce NOX Emissions. 4. At higher premixed ratio, low load range is limited by higher cycle by cycle variation whereas high load range is limited by knocking combustion.

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Fig. 19. Effect of premixed ratio on indicated specific NOX emissions.

Fig. 20. Effect of premixed ratio on indicated specific CO emissions.

The following conclusions may be drawn from the present study:

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1. Peak pressure rises by 17.8% and advances by 1.5 deg when premixed ratio is increased from zero to 80%. When 15% and 30% EGR is introduced, peak pressure reduces by 4.5% and 12.7% respectively when compared to without EGR condition. The corresponding angles of occurrence moved away by 1.5 deg and 2 deg. 2. The peak PRR increases by 28% and the angle of peak PRR advances by 3.5 deg. The corresponding values reduce by 8% and 2 deg retardation with 15% EGR and by 20% and 3.5 deg retardation with 30% EGR. 3. The occurrence of peak pressure moves towards TDC by about 2 deg when the premixed ratio increases from 0% to 80%. The corresponding occurrence of peak PRR, peak HRR and SOC moves away from TDC by 3.5 deg, 3 deg and 5 deg respectively. When EGR is introduced, the peak pressure moves away from TDC while the other parameters move towards the TDC. 4. At 80% HCCI combustion condition, Smoke opacity reduces by 58% and NOX increased by 10% when compared with Baseline combustion. At 15% and 30% EGR conditions, Smoke opacity reductions were 55% and 40% respectively whereas NOX emissions were reduced by 43% and 76% respectively when compared with Baseline combustion. HC and CO increases with increasing premixed ratio. There is a corresponding increase in HC and CO emissions with increase in EGR. 5. Based on results discussed in Appendix, it was observed that up to 17% load, premixed ratio of 60% or more results in high (COV of IMEP > 10%) cycle to cycle variation. In those situations, stratification should be increased by reducing premixed ratio. When load was 67%, Engine started knocking while running at 80% premixed ratio and the corresponding ringing intensity (RI) was 3.38 MW/m2. It is finally suggested that, either premixed ratio should be reduced or EGR must be supplied to suppress RI which in turn would help in extending high load range of such HCCI-DI combustion engine.

532

5. Conclusions

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The present study involves variation of premixed ratio from 0% to 80%. The main injection timing was retarded to 20 deg BTDC from the baseline value of 26 deg BTDC. At 80% premixed ratio 15% and 30% EGR were introduced. Based on this study the following observations may be noted:

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1. Both combustion phasing and SOC advance with increased premixed ratio. At higher premixed ratios, the SOC advances significantly. This could result in knocking combustion and thereby limiting the HCCI operating range. At higher load, premixed ratio needs to be reduced to avoid knock.

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Fig. 21. Effect of premixed ratio on indicated specific HC emissions. 505

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6. Uncited references [48–51].

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Acknowledgements

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This study is supported by the Department of Mechanical Engineering, IIT Delhi, India. The authors also express their sincere thanks to the technical staff for their immense support during the test rig development and experiments in the Internal Combustion Engines lab of Indian Institute of Technology-Delhi.

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Appendix A. Supplementary material

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Supplementary data associated with this article can be found, in the online version, at http://dx.doi.org/10.1016/j.fuel.2015.07.009.

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References

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