Homogeneous charge compression ignition (HCCI) combustion of diesel fuel with external mixture formation

Energy 35 (2010) 148–157

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Homogeneous charge compression ignition (HCCI) combustion of diesel fuel with external mixture formation D. Ganesh*, G. Nagarajan Internal Combustion Engineering Division, Mechanical Engineering Department, College of Engineering, Anna University Chennai, Sardhar Patel Road, Chennai, Tamilnadu 600 025, India

a r t i c l e i n f o

a b s t r a c t

Article history: Received 10 September 2008 Received in revised form 4 September 2009 Accepted 7 September 2009 Available online 29 October 2009

HCCI (Homogeneous Charge Compression Ignition) has been touted for many years as the alternate technology of choice for future engines, preserving the inherent efficiency of CIDI (Compression Ignition Direct Injection) engines while significantly reducing emissions. The current direction for all published diesel HCCI research is mixture preparation using the direct injection – system, referred to as internal mixture formation. The benefit of internal mixture formation is that it utilizes an already available direct injection system. Direct injected diesel HCCI can be divided into two areas, early injection (early in the compression stroke) and late injection (usually after Top Dead Center (aTDC)). Early direct injection HCCI requires carefully designed fuel injector to minimize the fuel wall-wetting that can cause combustion inefficiency and oil dilution. Late direct injection HCCI requires a long ignition delay and rapid mixing rate to achieve the homogeneous mixture. The ignition delay is extended by retarding the injection timing and rapid mixing rate was achieved by combining high swirl with toroidal combustion-bowl geometry. There is a compromise between Direct Injection (DI) and HCCI combustion regimes. Even under ideal conditions, it can prove difficult to form a truly homogeneous charge, which leads to elevated emissions when compared to true homogenous charge combustion and also strongly contribute to the high sensitivity of the combustion phasing to external parameters. The alternative to the internal mixture formation is, predictably, external mixture formation. By introducing the fuel external to the combustion chamber one can use the turbulence intake process to create a homogeneous charge regardless of engine conditions. This eliminates the need for combustion system changes which were necessary for the internal mixture formation method. With this method, the combustion system remains fully optimized for direct injection and also capable of running in HCCI combustion mode with nearly ideal mixture preparation. The key to the external mixture formation with diesel fuel is proper mixture preparation. In the present investigation a fuel vapouriser was used to achieve excellent HCCI combustion in a single cylinder air-cooled direct injection diesel engine. No modifications were made to the combustion system. In this study a vaporized diesel fuel was mixed with air to form a homogeneous mixture and inducted into the cylinder during the intake stroke. To control the early ignition of diesel vapour–air mixture, cooled (30  C) Exhaust Gas Recirculation (EGR) technique was adopted. Experiments were conducted with diesel vapour induction without EGR and diesel vapour induction with 10%, 20% and 30% EGR and results are compared with conventional diesel fuel operation (DI @ 23 before Top Dead Center (bTDC) and 200 bar injection pressure). Ó 2009 Elsevier Ltd. All rights reserved.

Keywords: Internal combustion engines Homogeneous charge compression ignition Diesel HCCI Nitrogen oxides reduction Emission reduction External mixture formation

1. Introduction In recent years, a lot of attention has been focused on air pollution caused by automotive engines. Diesel engines have been particularly targeted for their production of oxides of nitrogen

* Corresponding author. Tel.: þ91 44 22203267, þ91 9444274548 (mobile); fax: þ91 44 22203255. E-mail address: [email protected] (D. Ganesh). 0360-5442/$ – see front matter Ó 2009 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2009.09.005

(NOx) and smoke emissions. NOx is formed at high rates when temperatures are high, whereas smoke is formed in fuel rich regions within the combustion chamber [1,2]. Hence, it is essential to keep the peak cylinder temperature low in order to minimize NOx emission and also to allow for better fuel–air mixing thereby, reducing the smoke emission. In the HCCI engine, the combustion process is modified so that combustion occurs under lean mixture conditions which lower the local combustion temperature. The absence of locally high temperatures and a rich fuel–air mixture

D. Ganesh, G. Nagarajan / Energy 35 (2010) 148–157

during combustion process makes simultaneous reduction of NOx and Particulate Matter (PM) emissions possible. Thring et al., [3] used a labeco Cooperative Lubricant Research (CLR) engine and studied four stroke, diesel-fueled, HCCI operations with control by varying the intake temperature and EGR fraction over a range of equivalence ratios. Thring, was the first to call this type of combustion system as homogeneous charge compression ignition. Yanagihara et al., [4] used a combination of early injection and late injection. Here, 50% of the fuel was injected at 13  CA (Crankangle) aTDC and the rest was injected during the early compression stroke. It resulted in gradual heat release beginning at 10  CA bTDC and extending upto 20  CA aTDC. The main combustion occurred after the second injection (20–25  CA aTDC), which significantly improved the combustion efficiency. HC emissions were from about 5000 ppm to below 2000 ppm and CO emissions were also high. NOx and smoke levels were reported as very low. Lei shi et al., [5] have utilized the thermal energy of trapped exhaust gases to vaporize the fuel. Diesel was directly injected into the cylinder near intake TDC and valve overlap was adjusted to obtain a high internal exhaust gas recirculation. The effect of the engine load, speed, inlet temperature, external and internal EGR on HCCI combustion and emissions was studied. Simescu et al., [6] conducted an experimental investigation of Premixed Charge Compression Ignition (PCCI)-DI combustion coupled with cooled and uncooled EGR in an Heavy Duty (HD) diesel engine. The study showed significant NOx reduction at light load conditions with up to 20% PFI (Port Fuel Injector). The study however showed that early PCCI combustion could adversely affect NOx emissions by increasing in-cylinder temperatures at the start of diffusion combustion. The PCCI-DI combustion also showed increased Brake Specific Fuel Consumption (BSFC) and HC, CO and PM emissions. Ryan et al., [7] used an electronic port fuel injector located approximately 15 diameters upstream of the intake valve. This injector was used for injecting fuel into the intake air for HCCI mode of operation. The heated intake air and EGR allowed intake temperatures up to 240  C for fuel vaporization. The results showed that through controlled homogeneous charge compression ignition operation near zero smoke emission is possible. It was observed that HCCI operation was possible with compression ratio ranging from 8 to14, EGR rates from 30 to 60%, and air–fuel ratio from 12 to 28. From the experiments it was found that the best SOC (Start of Combustion) for HCCI fall in the range of 20 bTDC to TDC. Shawn Midlam-Mohler et al., [8] has developed an atomizer for external mixture preparation and the authors have investigated the effect of uncooled EGR, Boost pressure, Air–fuel ratio, intake air temperature, swirl and engine speed on HCCI combustion. From the previous research works it is noticed that, early injection, late injection and port fuel injection systems like Air assisted port injection with DI system (PCCI-DI) were used. In addition to that high levels of EGR and reduced compression ratio were demonstrated for simultaneous and substantial reduction of NOx and smoke emissions. In the above detailed methods the mixture was partially homogeneous. Hence the present system was developed to prepare a homogeneous mixture. This paper reports the results of detailed examination of the basic characteristics of HCCI combustion in which a different method of preparing homogeneous mixture of fuel and air (usually by early injection, late injection and air assisted port fuel injection) was demonstrated by using a device called fuel vapouriser and its effect on engine emissions, performance and combustion are investigated with various EGR proportions and without modifying the compression ratio of an engine. The results show that through this approach simultaneous and substantial reduction of NOx and smoke emission can be achieved.

149

In the present investigation, a homogeneous mixture of fuel and air was prepared by using a diesel fuel vapouriser [9–15]. The diesel vapour provided by this device forms a very light and dispersed aerosol, where due to their sizes, the droplets lose their momentum a short distance downstream of the nozzle (no wall targeting), follow the air motion very well, have very fast evaporation due to very high surface to volume ratio, and disperse very uniformly in the surrounding air stream. All the properties of this ‘‘gas-like’’ aerosol make it ideally suitable for external diesel mixture preparation, with the creation of a highly disperse and homogeneous mixture, minimal wall-wetting and very fast evaporation during the compression stroke. Fig. 1 shows the picture of a vaporized fuel leaving the vapouriser into ambient conditions. 2. Experimetal set up The test engine used was a single cylinder, air-cooled research engine which was modified to operate in HCCI mode. The engine test bench was equipped with     

Fuel vapouriser. Electronic control unit (ECU) to control Port Fuel Injector (PFI). EGR system. Data acquisition system and crankangle encoder. Pressure transducer

The technical specifications are summarized in Table 1. 2.1. Fuel vapouriser mounting & fuel injection system The provision was made in the intake manifold to house the fuel vapouriser as shown in Fig. 2. A fuel vapouriser consists of a heating element, ceramic pipe and stainless steel pipe of dia 20 mm which was inserted into a ceramic pipe of 22 mm dia. The length of the fuel vapouriser was 150 mm. The heating element (nichrome) was wound over the ceramic pipe. The ceramic pipe was provided to conduct heat alone. The fuel vapouriser was insulated with glass wool which also adds cushioning to the brittle ceramic pipe to protect against engine vibration during engine operation. The details of the fuel vapouriser are given in Table 2. The Port Fuel Injector (up to 6 bar) was mounted on the top of the fuel vapouriser to supply the correct quantity of fuel to the vapouriser. The port fuel

Fig. 1. Fuel vapouriser plume in ambient air, diesel fuel.

150

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Table 1 Technical data of single cylinder engine. Cooling system Displacement Stroke Bore Compression ratio Injection pressure Injection timing Rated output Rated speed

Table 2 Specifications of the fuel vapouriser. Air 662 cm3 110 mm 87.5 mm 17.5:1 200 bar 23 bTDC 4.4 kW at 1500 rpm 1500 rpm

injector was controlled by an Electronic Control Unit (ECU). The ECU controls both the timing and quantity of the fuel. Tables 3 and 4 shows the key timings for ECU design and ECU timing range. The bench test was conducted with ECU by varying the injection pressure and injection duration to determine the fuel consumption. Table 5 shows the data obtained from the bench test and it was compared with the fuel consumption of an engine with conventional operation. From the results Fig. 3 is obtained by calculating the fuel consumption of the vapouriser at each load. The ECU controlled low pressure fuel injector that injects the pre-determined quantity of fuel into the diesel fuel vapouriser. The fuel vapouriser was heated through a power supply. A temperature controller was used to maintain the temperature (85  C) to provide diesel vapour at all the loads. A low pressure electrical pump was used to supply diesel fuel to the low pressure injector. Both the pump and the injector were controlled by ECU. 2.2. Experimental procedure Experimental procedure involves observing and taking safety precautions, observations and measurement of engine performance, emissions and combustion parameters using appropriate Instruments. It includes intake air measurement, fuel measurement, power measurement, cylinder pressure measurement and emission measurements. The schematic diagram of the experimental set up is shown in Fig. 4. An orifice meter attached with an anti-pulsating drums measures air consumption of an engine with the help of a U tube manometer. The anti-pulsating drum fixed on the inlet side of the engine maintains a constant suction pressure, to facilitate constant air flow through the orifice meter. Two separate fuel-metering systems were provided to meter both conventional fuel injector and a low pressure injector. The fuel supplied to the low pressure injector was preset. If the preset quantity was not sufficient to maintain the rated speed then the

Maximum power consumption Diameter of the vapouriser Current rating of the heating element Length of the vapouriser System warm-up time

240 W 20 mm 4A 150 mm 8 min

duration was adjusted to maintain the speed and the corresponding fuel consumption was calculated from Fig. 3. The electrical dynamometer was used for measuring power output of the engine. The electrical dynamometer works in similar manner to hydraulic dynamometer applying the same torque reaction principle. The electrical dynamometer used was a motor–generator type consisting of a cradle generator that is operated as a generator for loading the engine. As the armature of the generator is rotated by the engine, its reaction with magnetic field tends to pull the field coils and the casing along with it. This rotation is prevented in the same way as with the hydraulic dynamometer. The electrical power generated by the dynamometer is usually dissipated as heat through banks of electrical resistances. The load and speed can be increased or decreased on the dynamometer and thereby on the engine, by switching on or off the load resistances and by varying the field strength. The output of the generator must be measured by electrical instrument and corrected in magnitude for generator efficiency. Exhaust emission from the engine was measured with the help of QROTECH QRO-402 gas analyzer and smoke intensity was measured with the help of Bosch smoke meter. Bosch smoke meter usually consists of a piston type – sampling pump and a smoke level measuring unit. Two separate sampling probes were used to receive sample exhaust gases from the engine for measuring emission and smoke intensity, respectively. A 50 mm diameter filter paper was used to collect smoke samples from the engine. A K-type thermocouple and a temperature indicator were used to measure the exhaust gas temperature. The cylinder pressure was measured using a Kistler (601A) water cooled pressure transducer. A Kistler crankangle encoder on the crankshaft (7200 points per cycle) was used to clock pressure data acquisition. For each measured point, the pressure data of 137 cycles were recorded. The pressure data saved was fed to the Engine Combustion Pressure (ECP) analysis software (Yokogawa, Japan) to determine the Heat release rate, cumulative heat release rate, etc.

dQ =dq ¼ ðg=g  1ÞPðdv=dqÞ þ ð1=g  1Þvðdp=dqÞ

(1)

Initially the test engine was started at no load in direct injection mode for engine to warm up. Once ECU gives the signal to inject the fuel into the vapouriser which was maintained at a temperature of 85  C provides the diesel vapour into the intake manifold. In the manifold a diesel vapour mixed with air to form a homogeneous mixture. This homogeneous mixture was inducted inside the cylinder during the intake stroke. When the engine attained the rated speed through diesel vapour–air mixture induction, the engine governor cuts the fuel supply to the conventional fuel injector (200 bar pressure) thereon the engine operated completely

Table 3 Key timing for ECU design. Events

Duration

Suction Compression Expansion Exhaust

220 144.5 144.5 220



Fig. 2. Fuel vapouriser with PFI Injector.

CA

ms 24.44 16.05 16.05 24.44

D. Ganesh, G. Nagarajan / Energy 35 (2010) 148–157 Table 4 ECU timing.

151

ECU

Duration in (ms)

Delay in (ms)

Maximum Minimum

23 0.4

7 1.5

in homogeneous mixture of diesel vapour–air mixture. Thus ensures that engine switched to diesel HCCI operation. It is well known that EGR [16–19] is a useful way to vary the cylinder gas temperature and the ignition timing could be delayed. In this investigation the temperature of the exhaust gas recirculated was held at 30  C by an EGR cooling device. EGR cooling would increase the temperature differential term in equation (2). The heat absorbed from the combustion process is proportional to EGR rate, specific heat at constant pressure, and the differences between combustion and EGR temperatures. It follows then that the greater the difference in temperature, the greater the heat absorbed from the combustion process, which leads to a lower rate of NOx formation. In addition lesser EGR temperature means that less volume is occupied in the inlet system. A lower EGR volume displaces a smaller fraction of fresh filtered intake air, thus displacing less O2, which helps in maintaining the combustion efficiency. In this study EGR was varied (10%, 20% and 30%) and its effects on the engine performance, emission and combustion were studied.

Table 5 ECU controlled low pressure injection bench test results. Parameters

Total fuel consumption (kg/h)

2.5

2 bar

3 bar

No of rev

Duration in ms

TFC in kg/h

TFC in kg/h

TFC in kg/h

1 1.25 1.5 1.75 2 2.25 2.5 2.75 3 3.25 3.5 3.75 4 4.25 4.5 4.75 5 5.25 5.5 5.75 6 6.25 6.5 6.75 7 7.25 7.5 7.75 8 8.25 8.5 8.75 9 9.25 9.5 9.75 10

2.4 3 3.5 4.1 4.7 5.3 5.8 6.4 7 7.5 8 8.7 9.5 10 10.5 11 11.5 12 12.5 13.4 14 14.5 15 15.5 16 17 17.5 18 19 19.5 20 20.5 21 21.5 22 22.5 23

0.118950719 0.140531277 0.162976867 0.180142796 0.204786681 0.215553122 0.226224232 0.230733156 0.241334205 0.249513989 0.258380744 0.274272972 0.29381905 0.307788552 0.318274933 0.335035751 0.351261304 0.380412371 0.40076025 0.414781509 0.423347196 0.452830189 0.468274112 0.482510624 0.512233212 0.554574488 0.602326056 0.625026466 0.667118644 0.685873606 0.709786006 0.745266347 0.793121977 0.817275748 0.864168618 0.872340426 0.917054986

0.24433 0.309013 0.352857 0.385731 0.421534 0.428198 0.445182 0.454014 0.457674 0.485926 0.52536 0.563681 0.592295 0.619128 0.64426 0.676599 0.719298 0.726736 0.754408 0.761806 0.796761 0.818636 0.845119 0.881457 0.904412 0.933291 0.981709 1.003058 1.053157 1.073845 1.126288 1.150429 1.233083 1.326742 1.346715 1.47012 1.481184

0.247734 0.316365 0.358601 0.392188 0.426898 0.468126 0.515273 0.564005 0.607033 0.642717 0.683175 0.701688 0.738185 0.760629 0.801738 0.841745 0.847302 0.9022 0.980405 0.998647 1.025712 1.067631 1.137572 1.216818 1.230513 1.319624 1.40038 1.43301 1.499238 1.552867 1.618421 1.665914 1.718277 1.751929 1.908209 1.978552 2.132948

1.5 1 0.5 0 0

5

10

15

20

25

Duration (ms) Pressure = 1 bar

Pressure = 2 bar

Pressure = 3 bar

Fig. 3. Variation of total fuel consumption with duration.

Q ¼ mCp DT

(2)

where m ¼ EGR flow rate, Cp ¼ specific heat at constant pressure, DT ¼ temperature differential between combustion temperature and that of EGR. 2.3. Uncertainty analysis All measurements of physical quantities are subject to uncertainties. Uncertainty analysis is needed to prove the accuracy of the experiments. To get the realistic error limits for the computed result, the principle of root-mean square method was used to get the magnitude of error [20].

Injection pressure 1 bar

2

DR ¼

"

vR Dx vx1 1

2  2 2 #1=2  vR vR Dx2 þ/ þ Dxn þ vx2 vxn

(3)

Using equation (3) the uncertainty in the computed values such as brake power, brake thermal efficiency and fuel flow measurements were estimated. The uncertainties in the measured parameters, voltage (DV) and current (DI), estimated by the Gaussian method, are 3 V and 0.14 A respectively. For fuel time (Dtr) and fuel volume (Dt), the uncertainties are taken as 0.2 s and 0.1 cc/s respectively. 1 Temperature measurement Uncertainty in temperature is:  1% (T > 150  C),  2% (150  C < T < 250  C),  3% (T > 250  C) 2 Percentage of uncertainty for the measurement of speed, mass flow rate, NOx, hydrocarbon, smoke and pressure are given below: i) Speed: 1.1 ii) Mass flow rate of air: 1.3 iii) Mass flow rate of diesel: 1.0 iv) NOx: 1.1 v) Hydrocarbon: 0.01 vi) CO: 0.8 vii) CO2: 1.2 vi) Smoke: 2.0 vii) Pressure: 1.1 3. Results and discussion The following sections of the paper describe the results of diesel HCCI combustion (external mixture formation) with various EGR proportions. The engine performance, emission, and combustion characteristics were investigated.

3.1. Engine performance and emissions The test engine was operated with diesel vapour induction without and with EGR, and its performance and emission characteristics were compared with those of conventional diesel combustion. The readings were taken only up to 75% load operation due to difficulty observed in controlling the combustion at full load. At full load operation the engine was shifted to direct injection mode of operation. Fig. 5 shows the variation of brake thermal efficiency with load for diesel vapour induction with 0%, 10%, 20%, and 30% EGR. From the figure it is observed that the brake thermal efficiency decreases with increase in EGR percentage for diesel vapour induction compared to direct injection mode of operation. The decrease in brake thermal efficiency is about 1.9%, 5.8%, 9%, and 12.6% respectively for diesel vapour induction with 0%, 10%, 20%, and 30% EGR at 75% load condition compared with that of direct injection mode of operation. As the mixture is formed externally, combustion timing can only be influenced by diluting the cylinder charge with exhaust

Brake thermal efficiency (%)

Fig. 4. Experimental set up.

30 25 20 15 10 5 0

0

25

50

75

Load (%) Diesel-direct injection mode Diesel vapour induction without EGR Diesel vapour induction with 10% EGR Diesel vapour with 20% EGR Diesel vapour induction with 30% EGR

Fig. 5. Variation of brake thermal efficiency with load.

100

Hydrocarbon (g/kWh)

Oxides of nitrogen (g/kWh)

D. Ganesh, G. Nagarajan / Energy 35 (2010) 148–157

12 10 8 6 4 2

3.5 3 2.5 2 1.5 1 0.5 0

0 0

25

50

75

153

0

100

25

Load (%)

75

Diesel vapour induction without EGR Diesel vapour induction with10% EGR Diesel vapour induction with 20% EGR Diesel vapour induction with 30% EGR

Fig. 8. Variation of hydrocarbons with load.

Fig. 6. Variation of oxides of nitrogen with load.

Carbon monoxide (g/kWh)

through the dissociation of CO2 (chemical effect) and results in a reduction in combustion pressure and temperature. The combustion products in EGR such as CO2 and H2O have a higher specific heat per unit mass than air; hence it reduces the combustion temperature and chemical kinetic reaction rate. Hence the EGR role is very crucial in controlling combustion phase as well as the rapid rise in cylinder pressure in HCCI combustion. The aim of this study is to reduce exhaust raw emission of NOx and smoke by homogenisation of the combustion. This is achieved using an appropriate diesel fuel premixing technology i.e. external mixing of diesel fuel and air. Fig. 6 shows the variation of oxides of nitrogen with load for diesel vapour induction with 0%, 10%, 20%, and 30% EGR. The figure clearly shows that the NOx emission reduces by 45%, 80%, 86% and 95% respectively for diesel vapour induction with 0%, 10%, 20%, and 30% EGR at 75% load condition compared with that of the conventional diesel combustion. Low emission of smoke is another attraction of HCCI combustion. However, diesel HCCI engine smoke reduction is dependent on efficient control of EGR. The mechanism for reduction in smoke emission is not clearly documented [21], but it is expected that the absence of diffusion combustion and localized fuel rich mixture discourage the formation of smoke emission. Fig. 7 shows the smoke emission behaviour of the engine operated with diesel vapour–air mixture compared to conventional diesel operation.

14 12 10 8 6 4 2 0 0

25

50

75

Load (%) Diesel-direct injection mode Diesel vapour induction without EGR Diesel vapour induction with 10% EGR Diesel vapour induction with 20% EGR Diesel vapour induction with 30% EGR Fig. 7. Variation of smoke with load.

100

Diesel-direct injection mode Diesel vapour induction without EGR Diesel vapour induction with 10% EGR Diesel vapour induction with 20% EGR Diesel vapour induction with 30% EGR

Diesel-direct injection mode

gas. Homogeneous combustion with its steeply rising pressure produces high combustion noise to avoid this; the charge must be highly diluted using EGR. As the proportion of cooled EGR increases, decrease in brake thermal efficiency is observed due to energy loss of the vapouriser and the losses due to the unburned fuel. Other reasons may be due to increase in HC/CO emission. In the CI engine, the NO is formed in very hot zones closer to stoichiometric conditions and soot is formed in the fuel rich regions. The in-cylinder average air/fuel ratio is globally always lean but locally the combustion process is not. This means that there is a large potential to reduce emissions of NOx and PM by simply mixing fuel and air before combustion. In the present study, a homogeneous mixture of fuel and air was formed by using a device called fuel vapouriser in the intake manifold (external mixture formation) to achieve HCCI combustion. The HCCI combustion can reduce NOx emissions up to 90–98% as compared to conventional direct injection diesel combustion. When cooled EGR was inducted with premixed diesel vapour–air mixture NOx emissions further reduces due to lower combustion temperature and pressure. The EGR acts as a heat absorption sink, primarily due to the heat absorbing capacity of CO2 (thermal effects) as well as

50

Load (%)

Fig. 9. Variation of carbon monoxide with load.

100

Exhaust gas temperature (degree celsius)

154

D. Ganesh, G. Nagarajan / Energy 35 (2010) 148–157 400 350 300 250 200 150 100 50 0 0

25

50

75

100

Load (%) Diesel-direct injection mode Diesel vapour induction without EGR Diesel vapour induction with 10% EGR Diesel vapour induction with 20% EGR Diesel vapour induction with 30% EGR

Fig. 10. Variation of exhaust gas temperature with load.

From the figure it can be observed that the engine operated with diesel vapour–air mixture exhibits a significant reduction in smoke at all loads. Smoke and NOx emission rise significantly as charge heterogeneity increases. However mixture is available inside the cylinder as a homogeneous mixture. Hence there is an absolutely free from liquid fraction of fuel pockets, unlike heterogeneous combustion that takes place in a conventional direct injection diesel engine. The reduction in smoke emission is observed to be about 39%, 56%, 78% and 83%, respectively at 75% load for diesel vapour induction with 0%, 10%, 20%, and 30% EGR compared to that of conventional diesel operation. Higher HC and CO emissions are one of the major setbacks of HCCI engine. One major factor, which contributes to higher HC and CO emissions, is low temperature combustion due to lean mixture and higher EGR levels which are necessary for HCCI operation. For HCCI combustion, the whole cylinder volume is full of homogeneous mixture of fuel and air, the combustion temperature is low, and so more HC can be generated. Also the low exhaust temperature inhibiting from oxidation of HC in exhaust process also makes the measured HC higher. EGR has two primary effects on HC emission; one is that the intake of some unburnt HC with exhausted gas into the next cycle leads to a decrease in HC

emissions, the other one is that the decrease of combustion temperature in the cylinder leads to an increase in HC emissions. Fig. 8 shows the variation of hydrocarbon emission with load. In HCCI mode, the HC and CO emissions are typically around 30 times higher than the standard diesel operation [22–27]. The hydrocarbon emissions are 0.51 g/kWh, 0.59 g/kWh, 0.69 g/kWh and 0.81 g/kWh at 75% load condition for diesel vapour induction with 0%, 10%, 20%, and 30% EGR whereas in conventional diesel operation it is about 0.25 g/kWh. With increase in EGR rate, the combustion reaction rate is reduced, the mean temperature in the cylinder is decreased and the combustion reaction becomes more incomplete, the reason is that more and more mid product CO cannot be oxidized completely into CO2 because of the decrease in temperature. Fig. 9 shows the variation of carbon monoxide emission with load. The carbon monoxide emissions are 4.32 g/kWh, 5.06 g/kWh, 6.3 g/kWh and 7.5 g/kWh at 75% load for diesel vapour induction with 0%, 10%, 20%, and 30% EGR whereas in conventional diesel operation it is about 1.28 g/kWh. The increase in HC and CO emission is due to the low temperature combustion of lean mixture and use of EGR which is necessary for HCCI operation. The variation of exhaust gas temperature for diesel vapour induction with 0%, 10%, 20%, and 30% EGR is shown in Fig. 10. It can be observed that in the external mixture formation method the exhaust gas temperature is found to be lower by 22%, 36%, 44% and 58% for diesel vapour induction with 0%, 10%, 20% and 30% EGR. It clearly shows that there is a substantial reduction in exhaust gas temperature particularly with high EGR percentage. From 10% to 30% EGR, at 75% load condition the exhaust gas temperature decreases from 162  C to 105  C. In general increase in EGR quantity results in a reduction in peak combustion temperature and hence a reduction in exhaust gas temperature.

3.2. Combustion characteristics The point of ignition is determined by the conditions in the cylinder during the compression stroke, so in order to influence the combustion phasing, the conditions of the compressed gas must be altered. Diesel engines have a high compression ratio, which has to be lowered to reduce the pressure and temperature during compression. Another possibility to retard the start of ignition and lower the rate of heat release is to use large amounts

Fig. 11. Variation of cylinder pressure with crankangle.

D. Ganesh, G. Nagarajan / Energy 35 (2010) 148–157

155

Fig. 12. Variation of heat release rate with crankangle.

both the maximum pressure rise and mean temperature were reduced and the observed trend was similar to the research finding [29] and is shown in Figs. 11 and 12. This is mainly due to the increased heat absorption capacity with increased CO2. Comparison of peak pressure and heat release rate at 75% load condition for diesel vapour induction with 0%, 10%, 20% and 30% EGR with conventional diesel operation is shown in Figs. 11 and 12. It is observed that the peak pressure decreases and start of combustion (SOC) also delayed when EGR rate increases from 0 to 30% EGR. The peak cylinder pressure occurs before and close to TDC. Many experimental and numerical results show that HCCI exhibits a twostage combustion consisting of cool flame region and the high temperature HCCI combustion region [30–33]. This kind of feature can be observed clearly in Fig. 12. It is worth noting the low temperature reaction and high temperature reaction heat release rates as the EGR percentage increases. Fig. 12 clearly shows that both Low Temperature Reaction (LTR) and High Temperature Reaction (HTR) peak is suppressed as the EGR percentage increases. This is due to the effect of EGR on peak pressure rise and cylinder mean temperature. As the EGR ratio increases the peak pressure shift towards TDC.

4 3 Load:75%, Speed : 1500 rpm

2

360

1

355

0 320 -1

330

340

350

360

370

380

390

400

410

420

-2

Crankangle (degree) Diesel-direct injection mode Diesel vapour induction without EGR Diesel vapour induction with 10% EGR Diesel vapour induction with 20 % EGR Diesel vapour induction with 30% EGR

Fig. 13. The effect of cooled EGR on HCCI combustion process.

430

CAD/degCA (TDC)

Rate of pressure rise (bar/deg CA)

of exhaust gas recirculation (EGR). EGR can be considered an inert gas that absorbs heat during the combustion, thus reducing the combustion rate. In this investigation EGR was used to reduce the combustion temperature and pressure. As the mixture is formed externally, combustion timing can only be influenced by diluting the cylinder charge with exhaust gas or by altering the temperature of the mixture. The start of combustion, therefore, can only be varied through gas mixture temperature. The temperature of the mixture after the vapouriser outlet has an average temperature of 315 K. The diesel vapour–air mixture flows out of the vapouriser with a temperature of 363 K. It clearly shows that the charge temperature is not very high. The theoretical temperature rise of the charge could be approximately 990 K at the end of compression. HCCI is governed by three temperatures. This required reaching the autoignition temperature to get things started; the combustion should then increase the temperature to at least 1400 K to have good combustion efficiency but it should not be increased to more than that 1800 K to prevent NO formation [28]. With increased EGR rate

350 345 340 335 330 325 0

5

10

15

20

25

EGR (%) Start of LTR

Start of HTR

Fig. 14. The effect of EGR on start of LTR and HTR.

30

35

D. Ganesh, G. Nagarajan / Energy 35 (2010) 148–157

16

18

14 17

CA50 (deg BTDC)

Interval CAD between LTR and HTR / degCA

156

16

15

13.85 13.1

12

11.2

10 8.1

8 6 4

14

2 13

0

5

10

15

20

25

30

35

EGR (%)

0 30

35

Load :75%, Speed:1500 rpm

Cumulative Heat Release Rate (J)

220 190 160 130 100 70 CA 50

10

-50

330

340

350

360

370

380

390

400

410

Crankangle (degree) Diesel-direct injection mode Diesel vapour induction without EGR Diesel vapour induction with 10% EGR Diesel vapour induction with 20% EGR Diesel vapour induction with 30% EGR

Fig. 16. Cumulative heat release rate.

60

65

Fig. 17. The effect of cooled EGR on combustion phasing.

From Fig. 13 it is observed that the maximum pressure rise rate is reduced with the increase of EGR rate. Increasing EGR rate can delay both LTR and HTR. Fig. 14 shows the influence of EGR rate on the starts of LTR and HTR. With increase of EGR rate, both the starts of LTR and HTR are retarded. The higher specific heat of the exhaust gas, which absorbs more of energy, reduces the in-cylinder temperature. At the same time, exhaust gas dilutes oxygen concentration in the inlet gas and thus decreases the speed of reaction. So, the increased EGR rate may delay both starts of LTR and HTR. Fig. 15 shows the effects of EGR on the interval between HTR and LTR. As the EGR rate increases the interval increases. It can be concluded that EGR can retard HTR longer than LTR and prolong the duration of the negative temperature coefficient area in HCCI combustion because more exhaust gas absorbs more energy released by the LTR. With the combustion before top dead center, the temperature will be increased both by the chemical reactions and the compression due to piston motion. Thus for a given autoignition temperature, combustion onset before TDC will result in faster reactions. With the conditions changed to give combustion onset close to TDC, the temperature will not be increased by piston motion; the only temperature driver would be the chemical reactions. This gives a more sensitive system and the later the combustion phasing the more sensitive the system is. This is the underlying problem with HCCI combustion control. It is desired to have a late combustion phasing to reduce the burn rate and hence pressure rise rate and peak pressure. The most usable parameter for

-20 320

45 50 55 50% Heat Release Rate (J) CA50 for 30%,20%,10%,0% EGR

Fig. 15. The effect of EGR on interval of LTR and HTR.

40

40

420

430

combustion phasing is the crankangle of 50% of the heat released (CA50). Fig. 16 shows the procedure to extract this 50% heat released point denoted by CA50. The best representation of combustion phasing was found by extracting the crankangle at which 50% of the maximum heat release was detected. Fig. 17 shows the CA50 for various EGR proportions such as 0%, 10%, 20% and 30%. It is noticed that the CA50 retards as EGR percentage increases. It shows that the combustion takes place before and close to TDC. Combustion phasing also shifts towards TDC, as mentioned earlier the late combustion phasing is the desired one for HCCI, this trend was observed from Fig. 17 as the EGR proportion increases. 4. Conclusions The combustion performance and emission performance of diesel HCCI with external mixture formation were studied using cooled EGR. The conclusions were obtained as follows. The investigation of diesel HCCI showed the potential of this type of combustion. HCCI is a promising concept for achieving low emissions at part load operations. It offers a solution for what is considered as a major drawback of diesel engines. This technique can be successfully applied to traditional direct injection diesel engines with low extra costs and no modification to the DI system by performing the mixture formation in the intake manifold. However, high HC and CO emissions are a major disadvantage. Even though it may be possible to reduce these emissions in the exhaust system with an oxidation catalyst, they also reduce the fuel efficiency considerably. The investigation has shown that a diesel engine can run on a homogeneous fuel/air mixture that is generated externally in a fuel vapouriser. The engine combustion with a homogenized mixture via fuel vapouriser is demonstrated. The aim of low NOx and smoke emission was achieved with fuel consumption about 12% higher compared to the conventional diesel operation when operating engine with 30% EGR. The main reasons for the fuel consumption penalty are the unburned fuel and the vapouriser loss. In order to delay and control the combustion, EGR was used in this investigation up to 30%. With EGR, combustion was controlled and delayed. The combustion occurs close to TDC rather than well before TDC due to EGR on combustion pressure and temperature. A further increase in the EGR for retarding the combustion would lead to higher HC/CO emission and to combustion instability. Thus the simultaneous reduction of NOx and Smoke emission with fuel vaporizer system was achieved in the limited engine operation area.

D. Ganesh, G. Nagarajan / Energy 35 (2010) 148–157

References [1] Dec JE. Advanced compression–ignition engines – understanding the incylinder processes. In: Proc. Combust. Inst, vol. 32; 2009. p. 2727–42. [2] Akihama K, Takatori Y, Inagaki K, Sasaki S, Dean AM. Mechanism of the smokeless rich diesel combustion by reducing temperature, SAE paper no 2001-01-0655; 2001. [3] Thring RH. Homogeneous-charge compression ignition (HCCI) engines. SAE Paper No 892068. Warrendale, PA: Society of Automotive Engineers Inc; 1989. [4] Yanagihara H. Ignition timing control at Toyota ‘‘unibus’’ combustion system. A new generation of engine combustion processes for the future. In: Proceedings of the IFP international congress; 2001. p. 34–42. [5] Shi Lei, Cui Yi, Peng Kangyao, chen Yuanyuan. Study of low emission homogeneous charge compression ignition (HCCI) engine using combined internal and external exhaust gas recirculation (EGR). Energy 2006;31:2665–76. [6] Simescu Stefan, Fiveland Scott B, Dodge Lee G. An experimental investigation of PCCI-DI combustion and emissions in a heavy-duty diesel engine, SAE paper no 2003-01-0345. [7] Ryan III TW, Callahan TJ. Homogeneous charge compression ignition of diesel fuel, SAE paper no 961160; 1996. [8] Midlam-Mohler Shawn, Guezennec Yann, Rizzoni Giorgio. Mixed-mode diesel HCCI with external mixture formation. DEER 2003 Newport; August 26th 2003. [9] Nakagome K, Shimazaki N, Niimura K, Kobayashi S. Combustion and emission characteristics of premixed lean diesel combustion engine. In: International congress & exposition, SAE paper no 970898; 24–27 February 1997. [10] Gray Allen III W, Ryan Thomas III W. Homogeneous charge compression ignition (HCCI) of diesel fuel, SAE paper no 971676; 1997. [11] Stanglmaier Rudolf H, Roberts Charles E. Homogeneous charge compression ignition (HCCI): benefits, compromises, and future engine applications, SAE Paper No 1999-01-3682; 1999. [12] Christensen Magnus, Johansson Bengt, Hultqvist Anders. The effect of combustion chamber geometry on HCCI operation, SAE paper no 2002-01-0425; 2002. [13] Hasegawa Roy, Yanagihara Hiromichi. HCCI combustion in a DI diesel engine, SAE paper no 2003-01-0745; 2003. [14] Buchwald R, Brauer M, Blechstein A, Sommer A, Kahrstedt J. Adaption injection system parameters to homogeneous diesel combustion, SAE paper no 2004-01-0936; 2004. [15] Pinchon P, Walter B, Re´veille´ B, Miche M. New concepts for diesel combustion. In: Thiesel 2004, Valencia, Spain; 7–10 September 2004. [16] Duret P, Gatellier B,Monteiro L, Zima P, Maroteaux D, Blundell D, et al. Progress in diesel HCCI combustion within the european space light project, SAE paper no 2004-01-1904; 2004. [17] Stein J, Du¨rnholz M, Wirbeleit F, Kopp C, Benz C. Homogeneous dieselmotorische verbrennung zur darstellung niedrigster emissionen. [Homogeneous diesel engine combustion for achieving minimum emissions]. In: 13th Aachen colloquium on vehicle and engine technology; 2004 www.iav.com.

157

[18] Puschmann Heike, Buchwald Ralf, Pannwitz Marcel, Sommer Ansgar. Homogeneous diesel combustion with external mixture formation by a cool flame vaporizer, SAE paper no 2006-01-3323; 2006. [19 ] Maiboom Alain, Tauzia Xavier, He´tet Jean-François. Experimental study of various effects of exhaust gas recirculation (EGR) on combustion and emissions of an automotive direct injection diesel engine. Energy 2008;33:22–34. [20] Holman JP. Experimental methods for engineers. New York: McGraw-Hill; 1994. [21] Tsolakis A, Megaritis A, Yap D. Application of exhaust gas fuel reforming in diesel and homogeneous charge compression ignition (HCCI) engines fuelled with biofuels. Energy 2008;33:462–70. [22 ] Megaritis A, Yap D, Wyszynski ML. Effect of water blending on bioethanol HCCI combustion with forced induction and residual gas trapping. Energy 2007;32:2396–400. [23] Juttu S, Thipse S, Marathe NV, Gajendra Babu MK. Homogeneous charge compression ignition (HCCI): a new concept for near zero NOx and particulate matter (PM) from diesel engine combustion, SAE paper no 2007-26-020; 2007. [24] Odaka Matsuo, Suzuki Hisakazu, Koike Noriyuki, Ishii Hajime. Search for optimizing control method of homogeneous charge diesel combustion, SAE paper no 1999-01-0184; 1999. [25] Jacobs Timothy J, Assanis Dennis N. The attainment of premixed compression ignition low-temperature combustion in a compression ignition direct injection engine. In: Proc. of the combustion institute, vol. 31; 2007. p. 2913–20. [26] Miller Jothi NK, Nagarajan G, Renganarayanan S. LPG fueled diesel engine using diethyl ether with exhaust gas recirculation. International Journal of Thermal Sciences 2008;47:450–7. [27] Lu¨ Xing-cai, Chen Wei, Hou Yu-chun, Huang Zhen. Study on ignition, combustion and emissions of HCCI combustion engines fueled with primary reference fuels, SAE paper no 2005-01-0155; 2005. [28] Kong Song-Charng. A study of natural gas/DME combustion in HCCI engines using CFD with detailed chemical kinetics. Fuel 2007;86:1483–9. [29] Morsy Mohamed H. Ignition control of methane fueled homogeneous charge compression ignition engines using additives. Fuel 2007;86:533–40. [30 ] Johansson Bengt. Homogeneous charge compression ignition in Lund, keynote paper. In: Symposium on international automotive technology, SIAT 2007, Pune, India; 17–20 January 2007. [31] Canova Marcello, Garcin Renaud, Midlam-Mohler Shawn, Guezennec Yann, Rizzoni Giorgio. A control-oriented model of combustion process in a HCCI diesel engine. In: 2005 American control conference, Portland, OR, USA; 8–10; June 2005. [32] Mack Hunter J, Aceves Salvador M, Dibble Robert W. Demonstrating direct use of wet ethanol in a homogeneous charge compression ignition (HCCI) engine. Energy 2009;34:782–7. [33] TorresGarcı Miguel, Jose´ Francisco, Aguilar Jime´nez-Espadafor, Lencero Toma´s Sa´nchez. 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;34:159–71.