Performance and Emission Characteristics of Dual Injection in Compression Ignition (CI) Engine

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PERFORMANCE AND EMISSION CHARACTERISTICS OF DUAL

INJECTION IN COMPRESSION IGNITION (CI) ENGINE

PRAVIN KUMAR 1 & A. REHMAN 2 1Research Scholar, Department of Mechanical Engineering, Maulana Azad National Institute of Technology,

Bhopal, Madhya Pradesh, India 2Professor, Department of Mechanical Engineering, Maulana Azad National Institute of Technology,

Bhopal, Madhya Pradesh, India

ABSTRACT

The continuously stringent emission regulations and the fast depleting primary energy resources have forced the

researchers to explore the environment friendly and more efficient combustion concepts. One such combustion concept

called Homogeneous Charge Compression Ignition (HCCI) technology has shown promises to reduce oxides of nitrogen

(NOx) and particulate matter (PM) emissions simultaneously while maintaining the thermal efficiency comparable with

that of conventional compression ignition direct injection (CIDI) diesel engine combustion. But the main problems of

HCCI are the combustion phasing and the ignition timing control. Recently, a combustion concept called compound HCCI

or HCCI-DI has been presented, which is a compromise to full HCCI. In HCCI-DI concept, only a part of the total fuel

inducted is premixed through port injection and the remaining part is injected directly into the combustion chamber.

This mode can overcome the problems of the combustion phasing and the ignition timing control.

The objective of the present study is to investigate the effects of the port injection timings on the performance and

the emission characteristics of HCCI-DI combustion mode on a single cylinder CIDI engine. The fuel used for both port as

well as direct injection is diesel. HC and CO and the NOx emissions for HCCI-DI mode were found to be higher, whereas

the smoke emissions were found to be lower than those of conventional CIDI. For HCCI-DI, the indicated specific fuel

consumption (ISFC) was lower, whereas the brake specific fuel consumption (BSFC) was higher relative to those of CIDI.

The indicated thermal efficiency (ITE) and the brake thermal efficiency (BTE) were higher and lower respectively for

HCCI-DI than those of CIDI mode.

KEYWORDS: HCCI, Nox, PM, Emissions, CIDI, HCCI-DI

INTRODUCTION

Due to stringent emissions regulations and the strong demand to reduce fuel consumption, there is a strong

interest to develop new highly efficient and environmental friendly combustion mechanism. One such combustion

mechanism is Homogeneous Charge Compression Ignition (HCCI), which has been widely investigated in recent times

[1-3].

The characteristic feature of HCCI combustion is that the fuel and air are mixed prior to the start of combustion

and the mixture is auto-ignited due to increase in temperature at the end of compression stroke. Therefore in some regards,

International Journal of Automobile Engineering Research and Development (IJAuERD) ISSN(P): 2277-4785; ISSN(E): 2278–9413 Vol. 4, Issue 6, Dec 2014, 15-28 © TJPRC Pvt. Ltd.

16 Pravin Kumar & A. Rehman

Impact Factor (JCC): 5.1066 Index Copernicus Value (ICV): 3.0

HCCI combustion incorporates the feature of both the spark ignition (SI) and compression ignition (CI) engines. It is

similar to SI in the sense that the both combustion modes use premixed charge, whereas it is similar to CI as both rely on

auto-ignition to start combustion [1, 3-5]. In HCCI combustion mode, the combustion starts by spontaneous auto-ignition

of the mixture at multiple sites under high temperature and high pressure conditions [5].

The main drawbacks of the diesel engines are their higher production of oxides of nitrogen (NOx) and smoke

emissions. The NOx is generated at high rates in high temperature regions, whereas the smoke is formed at high rates in

fuel rich regions in the combustion chamber. Hence, it is necessary to reduce the peak cylinder temperature to minimize

the NOx emission and also to allow for better fuel-air mixing thereby, reducing the smoke emission [6].

In HCCI, the combustion process is so modified that the combustion takes place under lean mixture conditions,

which lower the local combustion temperature. The absence of locally high temperatures and a rich fuel-air mixture during

combustion process makes the simultaneous reduction of NOx and Particulate Matter (PM) emissions possible

[6, 7].However, in spite of the existing benefits of HCCI combustion, there are still many challenges hindering its

commercialization. The main challenges are the control of the ignition timing and the combustion over a wide range of

speeds and loads duration. Besides, there are problems in controlling the high emissions of unburned Hydrocarbon (HC)

and Carbon mono-oxide (CO) [1, 4, 5, 7, 8]. Unlike the conventional combustion, HCCI lacks in the control of ignition

timing because there is no external control mechanism such as the fuel injection or spark timing, which are used in CI or SI

engines respectively [1, 3, 4, 7]. Achieving the required level of control during transient engine operation becomes even

more challenging, because the charge temperatures have to be correctly matched to the operating conditions during rapid

transients with a high repeatability since the speed and load are changing [3]. In HCCI, the ignition timing is completely

controlled by the chemical kinetics, and is hence influenced by the fuel composition, equivalence ratio, thermodynamic

state and the temperature-time history of the mixture. Therefore, the fuel physical and chemical properties, mixture

conditions, environmental conditions including pressure and temperature, residual rate and possibly reactivity of the

residual mixture, homogeneity, compression ratio, engine operating conditions such as engine speed and load, heat transfer

to the engine and other engine dependent parameters are the key factors affecting ignition timing and the combustion

duration of the HCCI engine. Hence, the occurrences of misfiring at low loads and knocking at high loads are generally

observed, which leads to the limited operation range of HCCI engine [1, 3, 4, 5].

The HCCI family can be classified into four types on the basis of the strategy used for introduction of fuel into the

combustion chamber. They are-1.Port injection [3, 6] 2.Early in-cylinder injection [9] 3.Late in-cylinder injection [10] and

4.Port and in-cylinder injection (HCCI-DI) [1]. Early in-cylinder direct injection requires properly designed injector to

minimize the wall wetting problem that may result in combustion inefficiency, lower thermal efficiency and oil dilution.

When the injection timing is sufficiently advanced, then the ignition delay is elongated. Late in-cylinder direct injection

requires a long ignition delay and fast mixing rate to obtain the homogeneous mixture. The ignition delay is extended by

retarding the injection timing, whereas the fast mixing rate is obtained by combining high swirl with toroidal combustion

ball geometry. In both, early and late direct injections, even under ideal conditions, it could prove difficult to prepare a

truly homogeneous mixture. So, higher emissions as compared to that from true homogeneous charge and high sensitivity

of the combustion phasing on external factors are inevitable [6, 7].

Christensen et al. [11] investigated port injection diesel fuelled HCCI combustion as part of their investigation of

variable compression ratios to control HCCI with different fuels. They reported significant smoke emissions for some

Performance and Emission Characteristics of Dual Injection in Compression Ignition (CI) Engine 17


conditions and very low NOx emissions, but not lower than those with gasoline. This behaviour was thought to be due to

poor vaporization of diesel fuel forming an inhomogeneous mixture. In addition, in order to avoid knocking, the

compression ratios need to be reduced, which results in poor thermal efficiency. D. S. Kim et al. [12, 13] presented a HCCI

diesel engine concept, in which only a part of the fuel was premixed by an injector incorporated at the premixing chamber

and the combustion was regulated by directly injecting the fuel by a Direct Injection (DI) injector. The premixing by fuel

injection through port gave sufficient time for the premixed fuel to evaporate and mix with the air before ignition, which

leads to the reduction of the NOx and soot emissions.

In recent years, a lot of research has been conducted to investigate the potential control methods of HCCI

combustion. Remarkable progresses have been achieved, but still the satisfactory results are to be obtained in thoroughly

expanding the HCCI combustion operation range particularly at high loads [1]. Some of the potential control methods

investigated and proposed to control the HCCI combustion mode without changing the cetane number of diesel fuel are

intake charge heating system [6], exhaust gas recirculation (EGR) [14-16], variable compression ratio (VCR) and the

variable valve timing (VVT) to change the effective compression ratio and/or the amount of the hot gases retained in the

cylinder [5]. Besides, in-cyliner direct dual injection [7, 17], in-cylinder direct injection [5, 18], fuel blending [2, 19-21]

have also been employed as a control methods in HCCI combustion mode.

D Ganesh et al. [6] study shown that the simultaneous reduction of NOx and PM with high HC and CO emissions

could be achieved using port injection with heating and with and without EGR. R. K. Mourya et al. [3] investigated the

combustion and emission characteristics of a port injected ethanol fuelled modified two cylinder engines with varying

intake temperatures and various air-fuel ratios. The results showed extremely low NOx but higher HC and CO emissions

for all the stable operating conditions. Lu Xingcai et al. [21] conducted experiments on HCCI engine using port injection

of neat n-heptane and 10-50% ethanol/n-heptane fuels. The results showed maximum Indicated Mean Effective Pressure

(IMEP), up to 50% Indicated Thermal Efficiency (ITE), longer combustion duration at light loads for blends. Whereas

constant energy input, cycle to cycle variation of injection timing, maximum gas pressure and the maximum gas pressure

crank angle deteriorated with increase in ethanol percentage in blend. The HC emission of n-heptane and 10-30% blend

were very low, which increased significantly when ethanol percentage was more than 40%. CO emission was the

maximum for IMEP in the range of 1.5-2.5 bars and decreased besides this range. Lei Shi et al. [14] achieved low NOx and

smoke emissions in a diesel fuelled HCCI engine injecting fuel before the Top Dead Centre (TDC) of the exhaust stroke

and employing negative valve overlap. Internal and external EGR were combined to control the combustion. Internal EGR

helped to reduce smoke emission, but reduced the high load limits of HCCI, whereas the external cooled EGR helped to

avoid the knock combustion of HCCI at high load. Tiegang Fang et al. [17] investigated the effects of fuel blends

(low sulphur diesel with biodiesel) and injection timings on NOx emission and natural luminosity on two stage in-cylinder

direct injection (DI) diesel engine at varying injection timings of dual injection.

The results showed the lower natural luminosity for bio-diesel than that of diesel for all the cases, whereas the

higher NOx emission for bio-diesel than that of diesel for the conventional combustion cases. Simultaneous reduction of

NOx and natural luminosity achieved for advanced low temperature combustion mode. Furthermore, it was inferred that

the multiple injection strategy along with fuel effects can be used to fine-tune the combustion performance. Myung Yoon

Kim et al. [7] examined the effect of a narrow spray angle injector and dual injection on the exhaust emissions of a small

DI diesel engine. The results showed that the dual injection strategy was highly effective in reducing NOx emissions, while



maintaining a high thermal efficiency. Also, a dual injection strategy with early timing for the first injection for HCCI and

the late timing for the second injection has the potential to reduce CO emissions and to suppress the deterioration of the

combustion efficiency. Hyung Jun Kim et al. [18] investigated the emission and performance behaviour of a narrow spray

angle and advanced injection timing ranging from BTDC 80˚ to BTDC 10˚ with two fuel masses for HCCI combustion in a

dimethyl ether (DME) fuelled diesel engine. The results showed the significant reduction in NOx before BTDC 30˚ and

high levels of HC and CO at BTDC 70˚. Furthermore, the investigation revealed the decreasing and increasing trend of

IMEP and Indicated Specific Fuel Consumption (ISFC) respectively with advancing injection timing. A. Megaritis et al.

[19] explored the effects of water blending in the fuel in order to reduce the pressure rise rates in bioethanol fuelled HCCI

combustion with forced induction and the residual gas trapping. They reported that the fixed rate water-ethanol blending

was effective for the reduction of the pressure rise rates at the higher loads. Also, increasing the amount of water in ethanol

resulted in the effective load range and the increased emissions. Junjun Ma et al. [1] investigated HCCI-DI compound

combustion mode, which could be considered as a compromise between premixed HCCI and conventional CIDI.

By regulating the quantities of port injected fuel and direct injected fuel, different premixed ratios (rp=0 to rp=1) was

obtained. rp=0 was equivalent to fully HCCI, whereas rp=1 meant the conventional compression ignition direct injection

(CIDI).

The combustion and the emission characteristics of HCCI-DI combustion mode and the effect of premixed ratio

and the direct injection timing were investigated over full load range and constant speed of 1800 rpm on a single cylinder

diesel engine. Furthermore, the comparison among conventional CIDI, HCCI-DI and the fully HCCI were done. Normal

heptane was chosen as the premixed fuel due to its low boiling point and excellent ignition ability and injected at 340˚ CA

BTDC (before top dead centre), while diesel fuel was directly injected at 7˚ CA BTDC. The results showed that the NOx

emissions decreased firstly at low premixed ratios and increased at higher premixed ratios. No significant effect of

premixed ratio on soot emission was observed except at a certain higher premixed ratio related to the equivalence ratio,

at which it was maximum. Unburned HC increased almost linearly with the premixed ratio due to incomplete oxidation in

the boundary layer and crevices. Xingcai Lu et al. [22] investigated the performance and emission characteristics of

compound HCCI combustion fuelled with gasoline and diesel blends.

The objective of the present study was to investigate the effects of port injection timings on the combustion and

emission characteristics of HCCI-DI combustion concept on a single cylinder CIDI engine. The diesel fuel has been used

for both the port as well as direct injection. The engine experiments were conducted over the full load range at the constant

speed of 1500 rpm. The combustion and the emissions characteristics were compared with conventional CIDI mode.

For all the tests the quantities of the fuel supplied through port as well as direct injection have been kept constant.

EXPERIMENTAL APPARATUS AND PROCEDURE

The research engine is a water cooled, vertical, 4-stroke cycle totally, enclosed, direct injection cold starting,

naturally aspirated diesel engine. The main engine specifications are given in Table 1. The actual photograph of the test

set-up has been shown in Figure 1, whereas the line diagram of the test bench has been shown in Figure 2. The test

conditions and the properties of the diesel fuel have been given in Table 2 and Table 3 respectively. The pump of the port

injection system was driven by the modified crank arm through cam follower arrangement, where as the direct injection

system was driven by in-built mechanical injection pump. The modified cranking arm had 36 numbers of grooves over its

surface, over which a cam having same number of internal grooves was mounted. The port injection pump, mounted just



over the cam, was driven by the cam. The injection timing was set by changing the position of the cam lobe by taking out

the cam from the modified cranking arm. The engine was coupled to an eddy current dynamometer to load the engine.

The engine torque was measured by a calibrated load cell.

Figure 1: Experimental Set Up

The cylinder pressure was measured by a piezoelectric pressure transmitter (6613CA, Kistler make), fitted flush

with the wall of the combustion chamber. It contains a piezoelectric sensor and an integrated charge amplifier Amplifier

provides uniform output and connected directly to a data acquisition unit. An optical shaft encoder was coupled with the

crankshaft to indicate the crank shaft position.

Table 1: Main Engine Specifications

The pressure data was recorded using a high speed memory for each measuring point. The pressure data of 200

consecutive cycles was sampled and recorded. The pressure trace for a specific condition was obtained by averaging the

sampled pressure data of 200 cycles, to calculate the IMEP, rate of heat release and other combustion related parameters.

Software was used to record the in-cylinder pressure versus crank angle for 200 consecutive cycles and to analyze the

resulting data. The exhaust gas compositions of CO, UBHC, and NOx



Figure 2: The Schematic Diagram of Experimental Set Up

Table 2: Test Conditions

Items Details Engine speed 1500 rpm

Port injection timings 80˚, 110˚, 140˚, 170˚ & 190˚ CA ATDC in suction stroke

Port injection quantity (fixed) 1.5 mg/cycle Number of injector holes 1 Type of fuel injection pump-line-nozzle injection system

Table 3: Properties of the Diesel Fuel

Density at 15˚ C 852 kg/m3 Viscosity at 40˚ C 3.38 cst Calorific value 43963 KJ/kg Cetane index 44.3 Flash point 67˚ C Fire point 73˚ C Carbon residue 0.28 % weight Ash 0.004 % weight Sulphur content 0.030 % weight

were measured by AVL di-gas 444 gas analyzer; whereas smoke opacity was measured by AVL 437 smoke

meter. An orifice-meter attached with an anti-pulsating drum was used to measure the air consumption of the engine with

the help of a U-tube manometer. The anti-pulsating drum fixed on inlet side maintains constant air flow through

orifice-meter and eliminates cyclic fluctuations. Two separate fuel-metering systems were provided to measure both the

port injected fuel and direct injected fuel. For all data presented, 0CA has been defined as the top dead centre (TDC) at

suction stroke. In this paper, CIDI indicates conventional compression ignition direct injection (CIDI) combustion mode,

whereas ATDC 80 indicates the port injection timing of 80˚ after top dead centre (ATDC) in suction stroke and so on.



The port injected fuel quantity is fixed, and is equal to 1.5 mg/cycle for all the tests.

In this paper the premixed ratio (Rp) was defined as: Rp= mp / (mp+ md), where mp indicates the port injected fuel

quantity, whereas md indicates the direct injected fuel quantity.

RESULTS AND DISCUSSIONS

Figure 3: Effect of Injection Timings on Brake Specific Fuel Consumption of HCCI-DI

The brake specific fuel consumption (BSFC) versus indicated mean effective pressure (IMEP) graph shown in

Figure 3 is in good agreement with the widely accepted trend. The BSFC decreases with increasing load on account of

higher power output at higher loads. Furthermore, it is evident from Figure 3 that the BSFC for HCCI-DI mode is greater

than that of conventional compression ignition direct injection (CIDI) mode. It may be due to the fact that slightly more

fuel is consumed in HCCI-DI as compared to CIDI and some of the power output is used to drive the pump for the port

injector. The difference in BSFC between HCCI-DI and CIDI is more pronounced at lower loads, when the fuel supply

through direct injection is lesser and the power output is lower.

Figure 4: Effect of Injection Timings on Indicated Specific Fuel Consumption of HCCI-DI

It is evident from the Figure 4 that the indicated specific fuel consumption (ISFC) decreases with increasing load,

which is generally agreed. Contrary to BSFC, ISFC for HCCI-DI mode is lower than that of CIDI mode. The indicated

power as well as fuel consumption for HCCI-DI is more than those for CIDI as calculated. But the increase in indicated

power (IP) is predominant, which reduces the ISFC for HCCI-DI. The increase in IP might be due to the two reasons.

The minor reason may be due to the increase in fuel consumption and the major reason may be due to the turbulence



created in air because of the pressurized port injection in air stream during suction process, which leads to better mixing

and therefore combustion resulting in higher IP. It can also be observed from Figure 4 that there is marginal difference in

ISFC at different injection timings.

Figure 5: Effect of Injection Timings on Brake Thermal Efficiency (BTE) of HCCI-DI

It can be seen from Figure 5 that the brake thermal efficiency (BTE) increases with the load, which is widely

accepted. Furthermore, the BTE for HCCI-DI is lower than that of CIDI, which was expected as the BSFC for HCCI-DI is

higher than that of CIDI. It can also be observed from Figure 4 that the brake thermal efficiency is the maximum for the

port injection timing of ATDC 190 as compared to other port injection timings due to lower brake specific fuel

consumption at ATDC 190.

Figure 6: Effect of Injection Timings on Indicated Thermal Efficiency of HCCI-DI

It can be seen from Figure 6 that the indicated thermal efficiency (ITE) increases with the load, which is widely

accepted. Furthermore, the ITE for HCCI-DI is higher than that of CIDI, which was expected as the ISFC for HCCI-DI is

lower than that of CIDI. It can also be observed from Figure 6 that there is marginal difference is ITE among different

injection timings. Though, ATDC 110 and ATDC 190 have comparable ITE, which are higher than those at other injection

timings.

It is evident from Figure 7 that the mechanical efficiency increases with the load, which is widely accepted.

Furthermore, the mechanical efficiency for HCCI-DI is lower than that of CIDI due to its higher indicated power and lower



brake power as explained earlier. It can also be seen that the mechanical efficiency is the maximum for the port injection

timing of ATDC 190 as compared to the other port injection timings.

Figure 7: Effect of Injection Timings on Mechanical Efficiency of HCCI-DI

Figure 8: Effect of Injection Timings on Smoke Emissions of HCCI-DI

The Figure 8 shows widely accepted trend for the smoke emissions with load. The smoke opacity is significantly

lower for all cases of HCCI-DI relative to that for CIDI. The lower smoke opacity for HCCI-DI may be due to the absence

of fuel-rich regions produced in the combustion chamber because of better mixing. The mixing is improved due to the

turbulence created by pressurised port injection in the air stream during suction stroke. Furthermore, it can be observed

from Figure 8 that there is no significant difference in smoke opacity among different port injection timings except at very

high loads. In fact, as reported by many researchers that the mechanisms of soot formation and oxidization in partially or

fully HCCI engines are greatly complicated and influenced by very many factors, and not completely understood.

The HCCI combustion, in itself, of premixed charge should produce the low or free soot emissions. However, actually, it is

not easy to provide an ideally homogeneous charge with a practical engine system and this is one of problems to be solved

so that HCCI combustion can be applied for the commercial engine. Especially, this problem is regarded to be more

significant in a diesel fuelled HCCI engine



Cracknell et al. [23] investigated the effect of a broad range of fuel properties on HCCI combustion strategy. They

observed several fuels, at certain speeds and loads, broke the NOx–PM trade-off curve and produced simultaneous

reduction of NOx and soot. But as the load increased, all fuels tended reverting to classic diesel NOx–PM trade-off curve.

Figure 9: Effect of Injection Timings on NOx Emissions of HCCI-DI

Figure 9 shows that the NOx emission increases with increasing load, which is generally accepted. As widely

recognized, the formation of nitrogen oxides is favoured by high oxygen concentration and high charge temperature.

As the in-cylinder temperature increases with the increase in load due to burning of larger amount of fuel, the NOx

emission also increases. Furthermore, it can also be observed from Figure 9 that the NOx emission for CIDI mode is lower

than those for HCCI-DI operations. Though, there is no significant change in NOx emissions for CIDI and the port

injection timings of ATDC 190 except at very high loads. Furthermore, it should be noted that there is significant reduction

in smoke opacity as compared to the marginal increase of NOx emissions especially with reference to port injection timing

of ATDC 190. So, in terms of smoke and NOx emissions, overall HCCI-DI seems to have advantage over CIDI. The lower

value of NOx for port injection timing of 190˚ may be due to the inertia effect of valve closing event leading to the better

mixing and therefore extending the low temperature regions. Further investigations are required to be carried out with

varying amount of port injected fuel because in this case the port injection quantity is too low to have significant effect.

Figure 10: Effect of Injection Timings on HC Emissions of HCCI-DI



As shown in the Figure 4, the HC emission first decreases and then increases at higher loads with increase in the

load. The HC emission is with good agreement with HC emissions of di-methyl ether fuelled HCCI-DI combustion [24].

In general, it was widely accepted that the HC emission came from the wall quench layer of combustion chamber,

ring-crevice storage and the absorption of fuel from oil layers. So it was expected that, as for HCCI-DI combustion, the CO

and HC emissions would be close to that of HCCI because the formation was of the same fundamental reasons. 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 may be generated. And the low exhaust temperature inhibiting from oxidation of HC in exhaust process

also makes the measured HC higher.

In general, the overall combustion event of compound HCCI combustion consists of premixed combustion of port

injected fuels and diffusion combustion of directly injected fuels. Specifically, at low premixed ratios, diffusive

combustion plays a vital role; thus, CO and HC emissions are at a low level, which is a feature of the DICI combustion

mode. When the premixed ratio increases, premixed fuel/air mixture concentrations before auto-ignition increase

correspondingly, then the overall combustion event is dominated by HCCI combustion. Therefore, the CO and HC

emissions start to increase. D. S. Kim et al. [12, 13, 24] reported that the increase of HC in partial HCCI or HCCI-DI

combustion are caused by a crevice effect and a flame quenching near the wall. Increasing rate of HC is measured to be

steeper for diesel premixing because of its poor vaporization. Effects of charge heating did not appear remarkable until

Tin= 80 ˚C. Gray et al. [25] found that heating the inlet charge could prevent the diesel fuel injected in through the intake

port from being transported into the combustion chamber in the form of liquid fuel jet, although the heating temperature

was far below the initial boiling point (175 ˚C). They suggested that the heating temperature of intake air should be over

130 ˚C at an injection pressure of 0.4 MPa with a gasoline port fuel injector used in their experiment.

Figure 11: Effect of Injection Timings on CO Emissions of HCCI-DI

It is clear from the Figure 11 that CO emission is the lowest for CIDI mode as compared to the HCCI-DI mode,

which was expected. As for HCCI-DI combustion, the CO emissions would be close to that of HCCI because the formation

was of the same fundamental reasons. Furthermore, the CO emissions decrease with the increase in load except at very

high loads. In general, it is widely accepted that CO emissions are controlled primarily by the fuel/air equivalence ratio and

the reaction from CO to CO2 is sensitive to the bulk gas temperature. Higher CO emission results in the loss of power of

the engine. Different factors can be at the origin of its formation, insufficient residence time, too low or too high

equivalence ratios are the part of those reasons. The formation of CO is much more complex. Unlike HC, CO is one of the

intermediate products of combustion reaction and its formation is controlled by the chemical kinetics. CO emissions



decrease due to increase in in-cylinder combustion temperature. The in-cylinder combustion temperature increases with the

load, which therefore decreases the CO emissions. The temperature decreases in the strong heat transfer regions such as the

regions of cylinder wall and piston surface, which therefore results in extending the low temperature regions of HCCI-DI

combustion. This leads to stronger flame quenching in the strong heat transfer regions for stratification combustion and

more CO cannot be oxidized to CO2. The CO emission is with good agreement with CO emissions of di-methyl ether

fuelled HCCI-DI combustion [24].

CONCLUSIONS

The following conclusions can be drawn from the present research work:

• The HCCI-DI combustion mode has the potential to reduce the smoke emissions maintaining the NOx emissions,

brake thermal efficiency and mechanical efficiency close to those of CIDI combustion mode.

• The indicated specific fuel consumption was lower, whereas the brake specific fuel consumption was found to be

higher for HCCI-DI mode as compared to CIDI mode. The indicated thermal efficiency was higher, whereas the

brake thermal efficiency and the mechanical efficiency were lower for HCCI-DI as compared to CIDI

combustion.

• There was significant reduction in smoke opacity over full load range at all injection timings for HCCI-DI relative

to CIDI mode. The NOx emission for HCCI-DI was higher in comparison to CIDI. But at port injection timing of

ATDC 190, there was only marginal increase in NOx and that too was at higher loads for HCCI-DI as compared

to CIDI. Therefore, in terms of NOx and smoke emissions, HCCI-DI has an advantage over CIDI mode. Over full

load range and for all injection timings of port injection, the HC and CO emissions were higher for HCCI-DI

relative to CIDI, it was expected that, as for HCCI-DI combustion, the CO and HC emissions would be close to

that of HCCI because the formation was of the same fundamental reasons.

• The port injection timing of ATDC 190 was found to be the best as compared to other injection timings for

improving the performance and the exhaust emissions of HCCI-DI mode of combustion. The brake specific fuel

consumption, the brake thermal efficiency, the HC, CO and the NOx emissions for ATDC 190 were very close to

those of CIDI, but the smoke emission for ATDC 190 is considerably lower than CIDI.

FUTURE RESEARCH DIRECTIONS

The HCCI-DI mode can be optimised by varying port injected fuel quantity and port injection timings for

improving performance and exhaust emissions.

ACKNOWLEDGEMENTS

The authors are grateful to the authorities of Maulana Azad National Institute of Technology (MANIT), Bhopal

(India) for granting permission for using their laboratory facilities.

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Documents

Performance and Emission Characteristics of Dual Injection in Compression Ignition (CI) Engine