A comprehensive review into the effects of different parameters on the hydrogen-added HCCI diesel engine

Abstract

The current study presents research investigations and developments related to the homogeneous charge compression ignition (HCCI) engine.  Research investigations and recent advances, including the role of various operating conditions on HCCI engine combustion phenomena, emissions, and performance, are discussed. There is growing research interest in investigating HCCI engines with diesel fuel to study combustion, emissions, and performance characteristics due to their association with low NO_x emissions. In the published literature, research investigations are also conducted with different fuels ranging from biomass to diesel to gasoline in the HCCI engine showing its capability for utilizing various fuels in coming years. The challenges associated with HCCI combustion are reviewed, and the details of excessive carbon monoxide and unburnt hydrocarbon emissions are discussed. The major parameters affecting the hydrogen addition in HCCI diesel engines are also discussed. Overall, adding hydrogen to a diesel-fueled HCCI engine improves combustion phasing and can potentially increase thermal efficiency while lowering emissions. In addition, the strength, weaknesses, opportunities, and threat analysis is provided and discussed thoroughly.

INTRODUCTION

With the European Union and various nations imposing harsher rules, it is not unexpected that the automobile industry is constantly exploring alternatives to compression ignition (CI) and spark ignition (SI) engines. Homogeneous charge compression ignition (HCCI) engines, which have excellent volumetric efficiency and low nitrogen oxides (NO_x) and soot emissions, are a possible substitute. The HCCI engine combines features from both the SI and CI engines. HCCI engines use an intake charge that is essentially a homogeneous combination of air, fuel, and burnt products, similar to that of a SI engine, but the resultant mixture is compression ignited, similar to that of a CI engine. Contrary to CI and SI engines, an HCCI engine is missing a direct trigger that initiates the combustion process but instead relies on the thermochemical route of the mixture getting compressed. As a result, ignition is affected by the pressure and temperature trajectory of the fuel mixture.

HCCI autoignition is similar to the knocking phenomenon in SI engine combustion, except that HCCI happens volumetrically across the cylinder rather than close to the flame front. The mixture is usually heavily diluted by air and the products of combustion from prior engine cycles, making flame propagation impossible. Dilution reduces NO_x production by lowering the temperature during combustion. The high compression ratios required to autoignite the mixture can result in efficiency comparable to CI engines, assuming that the indirectly controlled ignition begins at the appropriate moment. HCCI engines provide fuel flexibility because they can run on some fuels including gasoline,¹ diesel,² and alternate fuels.³ Although HCCI engines have many of the beneficial qualities of SI and CI engines, they add the challenging jobs of managing the start of combustion (SOC) and possess a restricted load spectrum that cannot meet the heavy load demands necessary for automobiles.⁴ Consequently, a global study is being conducted to look at the different factors that affect HCCI combustion.

The HCCI combustion phenomenon was used for the first time in a two-stroke engine.^5–10 The primary reason behind using HCCI combustion in a two-stroke engine is to control the emission of hydrocarbon (HC) during part load operation. Future works in four-stroke engines have revealed that lower NO_x emission and higher engine efficiency can be achieved by using a higher compression ratio and lean air–fuel mixture.^11–16 Several studies have been performed in a four-stroke engine, where the HCCI combustion phenomenon is studied. In most cases, the studies are conducted on a single-cylinder engine that does not deliver the brake output.

Nevertheless, a brake thermal efficiency (BTE) of 35% was achieved by Stockinger et al.¹⁷ using a brake mean effective pressure (BMEP) of 5 bar at an engine of a capacity of 1.6 L with a four-cylinder arrangement. Furthermore, a BTE as large as 40% was reported with BMEP values of 6 bar.¹⁸ Due to the autoignition phenomenon, the combustion in the whole cylinder is initiated approximately at the same time. The excess amount of air required for combustion was inducted into the cylinder to make a weak mixture for control. Residual gas controlling can be an additional productive way of dilution. Lack of adequate dilution of the mixture will lead to phenomena like fast combustion and knocking, along with higher NO_x emissions. However, excess dilution will lead to incomplete combustion and even misfire. This is known that the beginning of combustion is administered by chemical kinetics.^19–22 Instigation of combustion is governed by various factors like the quantity of fuel, oxygen concentration, pressure and temperature archived, and the product of combustion. Consequently, ignition timing is greatly impacted by a residual fraction of gas, compression ratio, inlet temperature, and air–fuel ratio. One of the major inspirations of the HCCI engine is its low NO_x formation.

HCCI ENGINE HCCI working principle

To control the combustion phenomenon in an HCCI engine, it is important to control the fuel delivery. During the suction stroke, fuel is injected into the combustion chamber of each cylinder with the help of an injector fitted just above the cylinder head. It is independent of the induction of air, which is processed through a suction pipe. After the suction stroke, the air and fuel get completely mixed up in the cylinder's combustion chamber.¹⁴ Then the process of compression follows, which increases the heat inside the cylinder. By the end of the compression stroke, enough heat is generated for the fuel–air mixture's autoignition without requiring any external spark. The power developed by the combustion pushes the piston downward during the expansion process. The whole fuel–air mixture gets combusted simultaneously, thereby, developing equivalent power. It involves less fuel burning and generates less emission. However, the HCCI engine has several limitations too, which are pointed out in Figure 1.

Figure 1. Limitations of homogeneous charge compression ignition (HCCI) engine.

After the completion of the power stroke, the pistons start moving upward, thus eliminating the exhaust gases. However, before the completion of the exhaust stroke, the exhaust valve gets closed and entraps some of the latent heat inside the cylinder. This heat is conserved, and a small amount of fuel is injected inside the cylinder for precharging, which controls the emission and the combustion temperature before the cycle repeats itself. The combustion process differs from the conventional SI and CI engine because it involves low temperature, lean mixture, and no flame propagation. Several typical dissimilarities between HCCI and SI engine and HCCI and CI are shown in Figures 2 and 3, respectively.

Figure 2. Comparison between homogeneous charge compression ignition (HCCI) engine and conventional spark ignition (SI) engine.

Figure 3. Comparison between homogeneous charge compression ignition (HCCI) engine and conventional compression ignition (CI) engine.

Different fuels in the HCCI engine

Generally, an HCCI engine is made to run on a lean mixture and at low-temperature conditions, making it suitable for various fuel applications. Fuels with lower calorific values may be favorably applied in such types of engines due to their high compression ratio. Different kinds of fuels, from biomass, diesel, gasoline, compressed natural gas (CNG), and so forth, are being used in the HCCI engine, showing its capability to deal with the fuel crisis problem. Syngas is one of the most important fuels that is used in HCCI engines due to its ability to be used at a high compression ratio. Syngas is mostly composed of methane (CH₄), carbon monoxide (CO), and hydrogen (H₂), with a slight trace of heavier HCs. Syngas is generated by gasification of waste, biomass, or coal in a high-temperature atmosphere. Syngas is a gaseous fuel produced as a result of partial oxidation of the feedstock under-regulated operational conditions and an oxygen shortage. Based on the gasifying substance, different types of syngas can be generated. Whilst syngas produced with air has a lower heating value (LHV) of 4–7 MJ/Nm³ and is more commonly referred to as producer gas, syngas produced with steam or oxygen has an average heating value of roughly 10–28 MJ/Nm³ and is known as power gas. The HCCI engine performance is significantly impacted by the relatively faster laminar flame speed of syngas compared to conventional fossil fuels. Another significant feature of syngas is its wide spectrum of flammability, which enables the use of the gas even when stoichiometric conditions are not reached; this is very helpful for HCCI engine applications. Ali et al.²³ conducted a numerical study to look at how piston bowl shape affects the emission and combustion characteristics of a syngas-fueled HCCI engine at medium and high loads. To investigate the influence of piston bowl shape, the standard piston was modified by reducing the piston bowl depth and it is observed that despite its greater thermal efficiency, modified pistons are unable to be utilized in the course of combustion because of significant maximum pressure rate (MPRR) and NO_x emissions, especially at high loads. Jamsran et al.²⁴ investigated the effect of different syngas compositions on the performance of the HCCI engine and found an optimum syngas composition (40% H₂, 16.5% CO, and 43.5% CO₂) for which the engine gives the best thermal efficiency. Higher hydrogen percentage is limited by the maximum rise in MPRR and limited operating range. Despite improvement in combustion characteristics, low-temperature combustion leads to almost zero NO_x emission. A numerical analysis of HCCI combustion using syngas yielded a close match to the experimental value. Peak pressure is shown to grow as the hydrogen proportion in syngas composition increases. The engine's lean functioning resulted in a smaller load range for higher engine speed.²⁵

Christensen et al.²⁶ experimentally controlled the ignition timing and the rate of combustion by water injection in an HCCI engine. They used three fuel combinations: natural gas, iso-octane, and ethanol with varied intake pressure, fuel–air ratio, and water inoculation rate. Experimental evaluation showed that the ignition timing could be controlled to a certain limit with the use of water injection along with the retardation of the rate of combustion. The maximum working load for the HCCI engine gets improved by applying water as it reduces the peak pressure as well as retards the combustion rate. But exhaust emissions, particularly unburned hydrocarbon (UHC) and CO, increase due to water injection, though NO_x emission decreases in the process. Experimental estimation on the effect of octane rating of primary reference fuel (PRF) was performed by Yao et al.^27,28 on the operational range, emission, performance, and combustion of a revised diesel engine to run on HCCI mode. Based on the results obtained, it is concluded that the duration of combustion and the peak value of the heat release rate (HRR) increases, and combustion timing gets promoted with fuels having a low octane value. The engine can be run at the lower magnitude of indicated mean effective pressure (IMEP) using low-octane fuel and at high IMEP with high-octane fuel. However, the vice-versa is impossible, limiting the higher engine speed. Antunes et al.²⁹ experimented on a modified diesel engine to run on HCCI mode fueled with hydrogen. They concluded that the peak value of pressure and rate of pressure rise is quite higher in HCCI engines running on hydrogen fuel than in conventional diesel engines. Hence, its capability to work on peak load reduces. This needs to be solved by changing the engine design. The port injection of n-heptane in a revised CFR engine run in HCCI mode results in a homogeneous mixture of fuel and air.³⁰ They concluded that the combustion stage and reaction get advanced with upper inlet charge temperature and compression ratio. They compared the HRR of experimental and numerical studies and found them in agreement with each other. Injection of ethanol at the engine port for a four-stroke, two-cylinder HCCI engine showed that the IMEP is quite high with a rich fuel–air mixture, while it substantially decreases with a dilution of the mixture.^31,32 A maximum indicated thermal efficiency (ITE) of 44.80% and combustion efficiency of 97.50% are achieved under an operating condition of 2.5 air–fuel ratios and 120°C of intake charge temperature. The exhaust gas temperature (EGT) was used for the heating of the intake charge with the help of a heat exchanger.³³ Wet ethanol was used as fuel in a four-cylinder Volkswagen engine adjusted to dwell on the HCCI mode. A blend of ethanol prepared from biomass with water was used at 80%–100%. The best performance was achieved for a blend of 80% ethanol with 20% water without needing any external heat source. High in-cylinder pressure and equivalence ratio (φ) were obtained in the HCCI engine working with wet ethanol. Equivalence ratio (φ) is the ratio of the actual fuel/air ratio compared to the stoichiometric fuel/air ratio. Stoichiometric combustion takes place when all the oxygen is used up in the reaction and there is no molecular oxygen (O₂) after the reaction. The NO_x emissions and the negative work due to higher in-cylinder pressure and φ can be controlled by adjusting the injection timing. The injection of CNG with varying fuel–air ratios and suction temperatures was performed in an HCCI engine to evaluate exhaust emissions and combustion characteristics.³⁴ The piston geometry considered was hemispherical. Furthermore, a software tool, CHEMIKN, was used to solve the complex chemical kinetics problem. The cylinder pressure was recorded together with the heat dissipation inside the cinder wall. It was seen that the single zone model does not perform well due to the constant homogeneous mixture throughout the engine operation. The simulation study did not contemplate the cracks inside the cylinder walls, which resulted in an under-estimation of CO emission compared to the experimental results. Moreover, a significant increase in the cylinder pressure and NO_x emission were observed with the rise in the sucked charges to that of the low temperature of the inducted charge.

The performance of an HCCI engine with ethanol as a working fluid with a misfire inside the cylinder is considered the output parameter variation with the misfire cycle.³⁵ No significant misfire effect was observed on output parameters like in-cylinder pressure, crank angle (CA), and HRR. It did not affect the cylinder pressure when the piston crosses the top dead center (TDC). Moreover, an artificial neural network model was used to find the correlation between the misfire cycle and the combustion characteristics to detect the point of the misfire. Knowledge of maximum heat release helps to detect misfires as it has an interconnection with that of misfire cycles.

Pioneering research on HCCI diesel engine

At the outset, researchers around the globe tried to understand the running of conventional fossil fueled HCCI engines by investigating the combustion characteristics and performing minor modifications, namely, port fuel injection, air preheating, and so forth. These are explained in the following sections.

Understanding the combustion phenomena

The HCCI combustion phenomenon was first explored for gasoline engines to achieve the stability of combustion for a two-stroke engine.³⁶ The engine emission and fuel economy can be controlled significantly by arranging the self-ignition of the fuel–air mixture. Combustion gets stable for HCCI engines with gasoline fuel within the speed range of 1000–4000 rpm and a compression ratio of 7.5:1 at different loads ranging from low to high. The experimental investigation of HCCI combustion in a two-stroke opposite piston engine resulted from the ignition occurring at several locations inside the cylinder without any visible flame propagation in the course of combustion.³⁷ Several intermediary compounds of O radicals and CH₂O were traced inside the cylinder with the help of a spectroscopic technique before self-ignition. These were developed due to the low-temperature self-ignition of high paraffinic HC fuel. After the ignition process, a large cluster of OH, H, and CH radicals was observed due to the high combustion temperature. An experimental study on a four-stroke HCCI engine showed that high-temperature characteristics drive low-temperature characteristics to influence the self-ignition phenomenon and the energy release during combustion.³⁸ The combination of this work with that of Onishi et al.³⁶ and Noguchi et al.³⁷ concludes that, unlike the conventional CI/SI engine combustion, HCCI combustion is more or less dependent on the chemistry of pressure, temperature, and composition of the charge inside the cylinder. The combustion in HCCI is largely affected by the uncontrolled ignition process and a smaller range of operations.

Operating HCCI engine with conventional fossil fuel

Using gasoline as a fuel in a four-stroke engine was explored by Najt and Foster³⁸ are further elaborated by Thring.³⁹ The operating range is limited to part-load and there is a severe problem with the control of the timing of the ignition. The discussion on the HCCI engine performance with gasoline is still a burning topic that continues now. Many researchers have preferred to term the HCCI combustion with gasoline fuel as autoignition combustion. But the concern related to diesel engine emission has shifted focus to studying the diesel-fueled HCCI combustion phenomenon, from around 1995. In diesel-fueled HCCI engine study, port fuel injection is a more basic way of producing the combustible mixture, which is studied in earlier research works with diesel fuel.^40,41 Ryan et al.⁴⁰ used a port fuel injection system in their study to provide fuel to the incoming stream of air, and an air preheater is used to preheat the same. The compression ratio being used is in-between 7.5 and 17:1. Furthermore, the study of Grey and Ryan⁴¹ has established that the diesel-fueled HCCI engine suffers from knocking and premature ignition while working under the same compression ratio as that of a diesel engine. An intake temperature of a fairly high value is needed to prevent the deposit of liquid fuel in the intake manifold. Furthermore, the emission of unburnt HC increases and NO_x emission drops significantly. This pioneering work has recognized that chemical kinetics play a vital role in HCCI combustion.

Challenges of HCCI combustion

Depending on the outcome of pioneering research work, it can be concluded that the full-scale use of HCCI combustion for industrial or transportation purposes is restricted by the drawbacks associated with this phenomenon. These drawbacks associated with HCCI combustion are illustrated in this section of the paper.

Homogeneous mixture preparation

Successful preparation of homogeneous mixture and avoidance of fuel collision with engine walls is quite important in obtaining better fuel efficiency, low emission of HC and particulate matters, and the prevention of dilution of fuel. Impingement of fuel on engine cylinder surfaces has been unfavorable regarding HC emissions, even for reasonably volatile organic fuels like gasoline.⁴² Homogeneous mixture characteristics affect the self-ignition properties of the HCCI engine, which controls its combustion phenomenon.⁴³ It is also observed that significantly low emissions of NO_x can be obtained even with a certain amount of inhomogeneous mixture composition inside the engine cylinder. Maintaining the mixture homogeneity is quite difficult, especially for fuel with low volatility like diesel. Furthermore, to obtain a lower smoke emission, a high intake air temperature is to be maintained for port injection of fuel. It is also seen that the HCCI mode of combustion has gained quite a lot of popularity in the past few years due to its associated advantage of high thermal performance along with low NO_x and particulate matter emission operating under part-load conditions.

Cold starting condition

Engine starting in cold weather conditions is one of the significant difficulties faced in the HCCI combustion technique as the average temperature is less than the engine operating temperature, thus resulting in the loss of heat from combustible charge to the engine cylinder walls. As a result, it becomes very difficult to self-ignite the combustible mixture inside the cylinder. Due to that, the engine may need a conventional method for starting the engine before it is finally shifted to the HCCI mode of combustion after a slight engine tune-up. Maintaining homogeneous combustion after that conventional start is quite a difficult task to achieve. The engine operation in cold conditions needs a lot of development to apply HCCI combustion effectively. Engine operation at low loads without losing any benefits in fuel efficacy and emissions is as important a research topic as the extension of engine operation range to a higher load.⁴¹ In the winter season or geographically cold regions, cold starts to become a challenge for HCCI engines. Because the compressed charge loses more heat to the wall at cold start operation.⁴⁴ Also, some modifications on fuel can improve the cold starting condition, the cold start condition can be improved.⁴⁵ in other words, the ozone quickly oxidizes and advances the combustion of all the fuels considered, and is, therefore, a good combustion promoter. Variable compression ratio is another promising technology in the HCCI engine, which gives the possibility to control the auto-ignition timing and extend the low- and high-load limitation due to the high effectiveness in changing the temperature and pressure at the end of the compression stroke in the combustion chamber prior the combustion takes place, consequently this technology can be employed to overcome the cold starting condition.⁴⁶

Range of operation

Along with the already mentioned issues with HCCI combustion, one more issue is the low operating range. Maintaining an extensive working range without losing the benefits associated with HCCI combustion is as important as self-ignition.⁴¹ The uncontrolled combustion phenomenon restricts the use of the HCCI engine at higher loads.⁴⁷ Its operation at a very low engine load is also restricted by the low engine temperature, which is insufficient for the self-ignition of the mixture of fuel and air. Moreover, low combustion temperature also leads to UHC and CO emissions.⁴⁸ So, the worth of this combustion method is limited by the excessive emission and low combustion efficiency, especially during idle operation.⁴³

Maximum pressure rise

Another HCCI's commercial feasibility is hampered by excessive maximum pressure rise rates under high load, resulting in combustion noise and possible engine damage.⁴⁷ Nevertheless, the extension of the HCCI operating range can reduce overall emissions and improve engine thermal efficiency.⁴⁹ By employing advanced control of HCCI combustion techniques the maximum pressure rise rates can be reduced. Most of the studies have focused on exploring one of three techniques: (1) Increasing charge dilution by use of excess air or exhaust gas recirculation (EGR), where investigating the effect of charge dilution with excess air on HCCI combustion. Dilution was achieved by increasing the boost pressure from naturally aspirated conditions to as much as 3 bars. The SOC was kept at the assumed set point with increased intake temperature. The authors observed simultaneous reduction of MPRR and NO_x emissions, without any trade-off on specific fuel consumption⁵⁰ (2) Introducing mixture stratification using late DI in which Dec et al.⁵¹ introduced an additional gasoline DI to the already existing port fuel injection system, aiming to achieve partial stratification of the in-cylinder charge. The authors observed a systematic reduction in the MPRR while increasing the ratio of directly injected fuel.⁵¹ Furthermore, Turkcan et al.⁵² revealed that an increase of stratification via delaying the start of injection of the second injection effectively reduced MPRR, but increased exhaust-gas opacity. (3) Another method of altering mixture reactivity in single-fuel HCCI combustion is via early fuel injection during the negative valve overlap (NVO) period. When the temperature of the recompressed exhaust gases is sufficiently high, the injected fuel undergoes chemical reactions to produce species with different reactivity. They are retained in the cylinder through the next main compression cycle and influence the auto-ignition properties of the air-fuel mixture and affects the main combustion event.⁵³

Excessive CO, UHC emissions, and noise release

Another major trouble associated with HCCI combustion is the large-scale release of CO, UHC, and noise release. The common thing that happens in the case of homogeneous combustion is that some fuel gets trapped during compression stroke inside a narrow crack on the cylinder walls, thereby escaping the combustion. Moreover, the combusted fuel temperature is insufficient to burn that uncombusted fuel. Hence, they go through the expansion phase and finally come out of the engine as UHC. Furthermore, the combustion temperature inside the cylinder is much lower (below than 1400 K) when operated under the part load. This results in a lack of capability to transform CO into carbon dioxide (CO₂), resulting in a steep decline in combustion efficiency.⁵⁴ Lower combustion efficiency and trouble with the ignition control restrict HCCI combustion's usefulness in part-load operation. While at higher loads, the pressure rises in the cylinder leading to excessive noise and engine damage, which further limits its potency.⁵⁵

The problem associated with the control of combustion phasing

The main problem associated with HCCI combustion is the control of combustion phasing. Contrary to the conventional combustion method, where a natural technique controls the combustion, there is no natural technique in HCCI combustion control. It is influenced mainly by the self-ignition quality of the air-fuel mixture, which in turn depends on the air–fuel mixture's quality and combustion phasing of the HCCI engine. Combustion phasing in the HCCI engine depends on the factors like⁵⁶: engine heat transfer, the temperature of the engine, latent heat, inlet temperature, compression ratio, homogeneous mixture, rate of residual reactiveness, the concentration of the fuel, and self-ignition properties among others engine relying conditions.

PARAMETER AFFECTING THE HYDROGEN ADDITION IN HCCI DIESEL ENGINE

The suitability of diesel as a fuel for HCCI engines is restricted owing to its low volatility.⁴¹ The problems such as low thermal efficiency, high HC emissions, and inappropriate combustion phasing are encountered.⁴⁴ Hydrogen, an attractive, clean fuel with high thermal efficiency, has been widely employed in running automotive and stationary engines.⁴⁵ Hydrogen can be the best candidate for adding to internal combustion engines alone or partially by replacing usual HC fuels to enhance performance and mitigate emissions.⁴⁶

Effect of percentage induction of hydrogen

A cooperative fuel research (CFR) engine was adapted to run on HCCI mode by applying hydrogen.⁴⁷ n-Heptane and two other fuels with Cetane numbers ranging from 36.6 to 46.6 were used. The rise in hydrogen quantity delays the combustion phases, and the combustion duration gets shortened. Variation in combustion phasing and combustion duration impacts the engine power output. The thermal efficiency and IMEP increase with the rise in the use of hydrogen at a constant 50% accumulated heat release (AHR) point (CA50) due to a decrease in combustion duration and retardation in combustion phasing. Delay in combustion phasing helps operate the engine at a higher compression ratio, with the rise in output power. Similarly, due to a decrease in combustion duration, maximum energy is released at the TDC, and complete combustion occurs, thus improving the thermal efficiency and the power output. Both indicated specific CO emission and CO emission per unit mass of fuel consumption decrease with an increase in hydrogen addition. However, indicated specific HC emission decreases. HC emission per unit mass of fuel consumption is not much affected by an increase in hydrogen addition. A performance and knock resistance study of an HCCI engine was performed by using CH₄ and its blend with hydrogen as fuel using a one-dimensional simulation technique.⁵⁷ It is concluded that adding hydrogen decreases the knocking tendency significantly and improves engine torque output, brake efficiency, and engine emissions. It is seen that the maximum improvement in brake torque output is found to be around 0.5%–0.7% for a hydrogen mass fraction in the range of 1.5%–2.5%. The growth in brake torque output is faster for the lower value of hydrogen content in the range of 1%–1.5%, while the growth is not so significant for a higher percentage of hydrogen content. It may be attributed to the improvement in fuel combustion due to the quick flame propagation rate, which is overshadowed by the effect of ignition delay at a higher hydrogen percentage. Brake-specific fuel consumption (BSFC) is improved due to the hydrogen's low density and higher heating value. With the increase in hydrogen addition, brake efficiency increases continuously along with a continuous decrease in HC, and CO emission due to complete combustion, however, after a hydrogen percentage of 2%, NO_x emission increases due to higher temperature of combustion resulting from high LHV value of H₂. A hydrogen–diesel dual-fuel study with hydrogen energy share (HES) varying from 0% to 20% is performed for the engine load range of 25%–75%.⁵⁸ The engine's BTE is marginally influenced by HES as observed in Figure 4. It is concluded that at lower load and HES, the combustion efficiency is quite low. This is because the fuel outside the spray region gets consumed due to the propagation of flame being in the lean zone, which results in combustion efficiency as low as 74.9% with HES of 4 at 10% load. This, in turn, results in a decrease in BTE and brake-specific energy consumption (BSEC). However, at higher load and HES, the decline in BTE and IMEP is not as high as in the case of low load and HES (Figures 5 and 6). CO emissions decrease by 7%, 11%, and 21%, for 20% HES at different loads of 25%, 50%, and 75%, respectively. Increased hydrogen content, in general, increases the NO_x emission due to higher in-cylinder temperature and pressure generation, however, at low load and low HES, NO_x level is found to decrease. At a low HES level of 5%, HC emission is found to increase, while with the increase in HES, the HC emission seems to decrease significantly.

Figure 4. Effect of percentage induction of hydrogen on brake thermal efficiency.

Figure 5. Effect of percentage induction of hydrogen on brake-specific fuel consumption.

Figure 6. Effect of percentage induction of hydrogen on indicated mean effective pressure.

In general, the use of hydrogen in a diesel-based HCCI engine offers appropriate combustion phasing and has the prospect of better thermal efficiency and lower emissions. Hydrogen is proved to deliver fairly better thermal efficiency owing to the retarded combustion. The harmful emissions such as nitric oxide will be limited. It is mostly caused by retarded combustion with lower HRR. Furthermore, hydrogen enrichment has added benefits for a diesel engine at a higher air/fuel ratio without EGR. The one that remains a threat to adding hydrogen is the prospect of knocking off a diesel-based HCCI engine. It may be further amplified by enriching hydrogen. The analyses of the strength, weaknesses, opportunities, and threat (SWOT) for the percentage induction of hydrogen are summarized in Table 1.

Table 1 SWOT analysis of percentage induction of hydrogen.

Strength	Weakness	Opportunity	Threat	Ref.
– The thermal efficiency will be improved due to combustion process retarding. – Hydrogen induction leads to enhance the power output. – In HDHCCI mode of operation; as the hydrogen energy ratio increased the BTE raised with the. – The smoke level has reduced as the HER is increased. – Increasing H2 leads to a reduction in both BSCO and BSHC emissions – Extremely low NOx is seen in exhaust.	– Large volumes of H2 and EGR are required to inhibit knock. – At BMEPs of 2 and 3 bar, the maximum hydrogen energy ratios were 41% and 47%, respectively. – A low HER resulted in knock, whereas a high HER resulted in misfiring. – Mass weighted values of HC and CO were not highly influenced by the amount of hydrogen inducted. – By hydrogen energy ratio of 24.2% the maximum brake thermal efficiency (BTE) that is reported was 31%. – At the BMEP of 3 bar, the engine operation without EGR was shown possible only with HER in the range of 26%–47% while above this HER there has been seen misfire and unstable combustion. – Since the engine will be violently knocking it is not viable to operate in DHCCI mode at BMEPs of 3 and 4 bar without employing the EGR.	– Harmful exhaust emissions will be reduced significantly. – Higher energy ratio of hydrogen always led to improve the BTE. – H2 required less air than diesel for combustion. – To achieved the best efficiency, while the amount of hydrogen increased the injection timing of diesel is retarded. – Reduction in the concentration of the OH radicals occur as a result of the reactions with H2. – Till an HER of 20%, it is beneficial to apply EGR in the HDHCCI mode. – The operation in the HHCCI mode is solely possible with very lean mixtures, i.e., an excess air ratio in the range of (3–6). – Due to reduction in the combustion rate by increasing HER, the emission of nitric oxide levels decreased significantly at all operating conditions. – As the HER increased the CO2 emission raised which is the indication of complete combustion process.	– The amount of hydrogen used had no effect on the amount of HC and CO emitted. – EGR improved thermal efficiency in the HDHCCI engine only at low flow rates hydrogen. – Increase in hydrogen supply and EGR slowed combustion and caused misfire. – The greatest IMEP that could be attained in the HHCCI mode was 3.5 bar. – The greatest IMEP of HHCCI engine can be 3.5 bar. – Running engine above 4 bar BMEP is not likely due to excessive banging.	[40, 42, 50, 51]

Abbreviations: BSCO, brake-specific carbon monoxide; BSHC, brake-specific hydrocarbon; EGR, exhaust gas recirculation; HER, hydrogen energy ratio; IMEP, indicated mean effective pressure; SWOT, strength, weaknesses, opportunities, and threat. Effect of compression ratio

Numerical analysis for the compression ratio in the range of 15–20 and steam injection at the rate of 0%, 10%, and 20%, showed a decrease in indicated power with increasing steam injection.⁵⁹ A maximum reduction of 17% is achieved for a compression ratio value of 20. The indicated torque generated in an engine is estimated through IMEP, which affects the engine's mechanical efficiency. An increase in IMEP is achieved by increasing the compression ratio for all different steam injection amounts. It is found that the power output decreases with an increase in steam injection beyond 20% due to a drop in the temperature inside the cylinder. CO emission is found to drop by 40% and 70% due to steam injection of 10% and 20%, respectively. Similarly, CO₂ emission decreases by 10% and 20% for steam injection values of 10% and 20%, respectively. Two different fuels Fusel40 (37.5% fusel oil and 62.5% n-heptane) and RON40 with octane number (ON) 40 are used for the proper comparative analysis.⁶⁰ Different performance parameters, namely ITE, BSFC, and emission parameters, namely HC and CO, are evaluated at a constant intake temperature of 350 K and a varying compression ratio of 11–13, respectively. It is seen that the engine performance is not much affected by Fusel40 though it has a LHV. A minimum BSFC of value 228.6 g/kWh is obtained for an engine speed of 800 rpm, and a compression ratio value of 13, with an increase in compression ratio, a decreasing trend for BSFC is observed. Thermal efficiency is largely dependent on CA50, which should be around 7–11°CA ATDC for maximum thermal efficiency. However, with an increase in compression ratio, the combustion phasing is delayed, resulting in decreased thermal efficiency. A maximum ITE value of 44.6% at compression ratio 12% and 43.4% at compression ratio 12 is obtained for RON40 and Fuesel40, respectively. HC and CO emission of 435 ppm and 0.2% and 401 ppm and 0.1% is obtained for RON40 and Fuesel40, respectively. The use of RON20 and RON40, operating with varying compression ratios of 9:1 to 12:1 in an HCCI engine, provided a maximum thermal efficiency of 37.3% with a compression ratio of 11:1 for RON20.⁶¹ Increasing compression ratio beyond 11:1 resulted in an increase in negative piston work due to more advancement of combustion phasing, which decreased thermal efficiency. IMEP increases with an increase in compression ratio due to an increase in oxidation at the end of compression and flame velocity. With the increase in compression ratio, emissions of HC and CO decrease with in-cylinder temperature; however, the NO_x emission increases for the very same reason. The maximum range of operation is obtained for RON20 with a compression ratio of 10:1, having the lowest BSFC of 210 g/kWh. An increase in compression ratio beyond that raises both the tendency of knocking and BSFC.

The compression ratios ranging from 17 to 20 have proved feasible and sustainable for adding H₂ in HCCI engines at both 80°C and 100°C of intake temperatures. In addition, enriching hydrogen in natural gas is a realistic approach for controlling combustion phasing in an HCCI engine. Moreover, it should be noted that since dimethyl ether (DME) has a high cetane number and low evaporating temperature, it is an ideal option to use in an HCCI engine. This may provide a stable solution with the least nitrogen oxide (NO_x) emission. The investigations of SWOT for varying the compression ratio in an HCCI engine are illustrated in Table 2.

Table 2 SWOT analysis of variable compression ratio on HCCI engines.

Fuel	Strength	Weakness	Opportunity	Threat	Ref
Hydrogen	– Compression ratios range from 17 to 20 has proved feasible and sustainable for adding H2 in HCCI engines at both 80°C and 100°C of intake temperature	– Operation at compression ratios of more than 18.5 is limited due to high thermal losses – Operation at low compression ratio is also limited by decreasing efficiency, due to high chemical and sensible enthalpy losses, and by decreased operational range – For CR ranges from 8.8 to 12.8 it is necessary to preheat the inlet air at the lowest compression ratios to initiate HCCI combustion	– Hydrogen HCCI has proved feasible at compression ratios around 18 with intake heating to 100°C	– Thermal losses steadily increase for compression ratios above 18–19	[55]
Natural gas—Hydrogen	– If other engine thermodynamic conditions are fixed, increased compression ratio would advance ignition timing – Peak rate of pressure is somewhat higher for higher compression ratios	– Very low rate of heat release, or not at all, affected by compression ratio	– Compression ratio should be selected as high as possible to minimize NOx emissions at high load, but low enough that control authority is not killed at peak load	– Maximum cylinder pressure is lower at low compression ratios at late combustion timing	[9]
Dimethyl ether	– The dimethyl ether HCCI engine can realize zero NOx emission and smokeless combustion under the compression ratio of both 10.7 and 14 – The maximum cylinder pressure of high load is noticed when high compression ratio of 14 is applied in HCCI engine	– At CP of 14, the beginning point of combustion is strongly tied to the concentration of mixture as temperature at the end of compression rises, it proceeds further vs. crank angle as engine load rises – As the negative compression work increases, mechanical losses increase, and fuel economy and thermal efficiency drop	– Higher the compression ratio, the engine can be fueled with a much thinner mixture – For enhancing the stability and fuel economy, while working at lower load range, high CR should be adopted in HCCI engines – Higher the CR, higher is the compression temperature and pressure	– At CR of 14 the engine operation is noisy at high-loads	[56]

Abbreviation: SWOT, strength, weaknesses, opportunities, and threat. Effect of equivalence ratio

A diesel engine adapted to run on HCCI mode using hydrogen as a fuel showed that the engine can run with the leanest fuel–air ratio and produce high BTE compared to that of a conventional diesel engine.²⁹ A maximum BTE of 45% is achieved while the engine is run at 2200 rpm and at an equivalence ratio of 0.33, which is quite high compared to a conventional diesel engine. NO_x is formed by the oxidation of atmospheric nitrogen, a temperature-dependent process. The use of an HCCI engine reduces NO_x emission due to low operating temperature, which is further reduced with decreased value of equivalence ratio. A similar reduction of NO_x emission for HCCI engines for lean combustion is also confirmed by Senthur et al.⁶² At a higher value of equivalence ratio (φ ≥ 0.33), the NO_x emission is found to be quite high. However, a decrease in φ decreases NO_x emission sharply. The emission of CO and UHC are found to be constant at all loads. Different blends of biodiesel with diesel influence the performance of a modified HCCI engine at various air-fuel ratios (2.1–5.6).⁶³ The occurrence of combustion restricts the lower limit, and the higher limit is restricted by knocking. The EGT is found to increase with engine load due to the inclusion of more fuel in the combustion. A higher ISFC is observed at a higher load due to more fuel supply. With the increase in biodiesel content, ISFC increases; this may be attributed to biodiesel having a lower calorific value. With the decrease in φ, fuel supply decreases; hence the rate of combustion decreases, but it leads to an earlier SOC, which results in energy loss during compression. Therefore, the best possible thermal efficiency (ITE) is obtained for medium-load operating conditions. According to Banke et al.,⁶⁴ a higher equivalence ratio of additives escalates the early release of heat and fast-tracks the core combustion. The traces with DME and n-heptane show a very distinct low-temperature heat release (LTHR) followed by a negative temperature coefficient near TDC. It is seen that with the HCCI mode of operation, an engine can be run at a wide range of φ.⁶⁵ The engine runs satisfactorily at φ of 0.5 to a very low value of 0.2. With the rise in φ, peak cylinder pressure, the peak value of HRR, IMEP and work done per cycle per kg fuel increases. The work output per unit of fuel decreases sharply with a drop of φ beyond φ = 0.3. It is reported that the peak pressure and NO_x emission are among the most sensitive factors to φ variations. When the value of φ crosses 0.5, NO_x emissions significantly increase and go beyond 1000 ppm. Also, it has been observed that the most optimal value φ lies within the range of 0.15–0.37. Varying the equivalence ratio has the following SWOT effects, as reported in Table 3.

Table 3 SWOT analysis of equivalence ratio on HCCI engines.

Strength	Weakness	Opportunity	Threat	Ref
– For fairly dilute equivalence ratios (0.2–0.5), peak efficiency occurs – By raising the equivalence ratio from 0.4 to 0.65, the ignition timing is advanced up to 15°CA and the maximum cylinder pressure is enhanced – The optimal equivalence ratio was found to be between 0.15 and 0.37	– As the equivalency ratio falls at a constant temperature, the sensitivity of phasing to fuel flow and temperature changes rises, till there is misfire – The range of acceptable equivalency ratios is somewhat limited – With the exception of operating sites at very low equivalence ratios of about 0.2, anticipated efficiency decreases as the equivalence ratio rises – Phasing is severely influenced by beginning temperature for equivalence ratios greater than about 0.3	– At low equivalence ratios (low load), the operation necessitates a rapid increase in beginning temperature, but efficiency falls as a result – An advancement of the ignition timing is observed with an improved maximum cylinder pressure by increasing the equivalency ratio – Drops in CO and HC are reported when the equivalency ratio has been increased – As the equivalency ratio is increased, combustion phasing should advance	– As the equivalency ratio rises, the peak pressure also rises – To minimize high peak pressures, hydrogen HCCI will be limited to dilute reactant mixes – The most sensitive factor to the dilution level is peak pressure – As the equivalent ratio rises, the temperature of residual gases rises dramatically – NOx emissions have grown as a result of increasing the equivalency ratio	[46, 55]

Abbreviation: SWOT, strength, weaknesses, opportunities, and threat. Effect of piston bowl shape

Different piston geometries and their effects on the performance and combustion were explored for a single-cylinder four-stroke HCCI engine run by iso-octane.⁶⁶ Three piston crown geometries named A, B, and C for clarity are used for the study. The engine speed and compression ratio are fixed at 1500 rpm and 11.7:1, respectively. Piston A has a concave shape head with a small conical at the middle of the base diameter of 10 mm, and height of 4.7 mm. This conical shape helps in swirling motion, with better air–fuel mixing inside the combustion chamber. Piston B has a three-stepped conical crown with the internal two separated from the external one by a wall of various outer and inner slant angles. Piston C has a head similar in shape to a Mexican hat. IMEP increases with the advancement of the combustion phasing due to better combustion efficiency, which may be attributed to the fact of better air–fuel mixing due to better-swirling action and turbulent kinetic energy. A numerical exploration was performed to analyze the influence of the flat, square bowl, and toroidal piston shapes on the performance and combustion of an HCCI engine run by syngas.⁶⁷ The engine operates at a fixed compression ratio of 17.1, for low, medium, and high loads. First, the optimization of the shape of the square bowl is done by varying the ratio of depth and squish area from 34% (A) to 20% (B), 10% (C), and 2.5% (D) and matched with the flat piston. Out of all the pistons, Piston D shows low MPRR due to retardation of combustion; also, it shows the maximum combustion and thermal efficiency compared to A, B, C, and flat pistons. A heavy-duty diesel engine (Scandia D12) was modified to run in HCCI mode through port injection of ethanol and air pre-heating.⁶⁸ It is seen that the range of operating load can be raised by a piston having a square bowl shape. This is because the combustion rate decreases in square bowl combustion, leading to a rise in combustion duration and a lesser rise in peak pressure. A drop in the rate of combustion occurs owing to the rise in heat transfer through the cylinder wall due to higher turbulence during the intake stroke. The indicated efficiency at a limited rate of pressure rise is reported to be similar for both square bowl and disc-shaped pistons for high-load operation; however, at lower load, efficiency decreases with a rise of CO and UHC. The spherical piston bowl facilitates complete combustion more than the other bowl types. The SWOT investigations of different piston bowl shapes on HCCI engines are stated in Table 4.

Table 4 SWOT analysis of piston bowl shape on HCCI engines.

Piston bowl shape	Strength	Weakness	Opportunity	Threat	Ref
Reentrant	– The maximum in-cylinder pressure of 13.6024 MPa at 719.6 CA is reported	– During the initial stages of the cycle, the wall heat transfer losses in the reentrant piston bowl is more	– Lower temperatures in this type of Piston reduce the formation of NOx	– The majority of the turbulence has increased heat loss through the cylinder's walls	[69, 70]
Spherical	– A maximum in-cylinder temperature of 1211.259 K at 719.0 CA was obtained – A reduction of 2.96% in CO emissions was obtained with spherical piston bowl – Higher CO2 emissions are recorded in the case of spherical piston bowls as the conversion of CO to CO2 is high. This is due to high cylinder temperatures – An increase of 9.16% in CO emissions was obtained with a spherical piston bowl	– The maximum peak pressure is happening very near to TDC, in which 13.863 MPa at 719.9 CA was obtained	The higher in-cylinder temperatures represent the increased turbulence in the combustion chamber, creating a well homogeneous mixture leading to an ideal volumetric combustion Higher in-cylinder temperatures facilitate the conversion of CO to CO2 Higher in-cylinder pressures spherical piston bowl managed to produce higher piston work	After 730°CA, the wall heat transfer losses in the spherical piston bowl are predominant. A spherical piston bowl has produced more NOx	[69, 70]
Square-bowl piston	– The higher turbulence level enhances the wall heat transfer significantly	– Has a lower peak heat release and longer combustion duration	– Ignition in both geometries still occurs at about the same time	– Prolonged combustion duration has been observed	[71]

Abbreviation: SWOT, strength, weaknesses, opportunities, and threat. Effect of intake temperature

For a hydrogen-fueled HCCI engine, the ignition timing variation is approximately linear to the inlet air temperature. However, when the air inlet temperature rises, the engine's power production falls. The IMEP and the BTE decrease as the air inlet temperature rises. It results in a drop in volumetric efficiency.²⁹ Furthermore, raising the intake temperature from 80°C to 100°C reduced anticipated efficiency, except for operating sites with very low equivalence ratios of less than 0.2.⁴⁹ The engine could run in HCCI mode at 120°C with an air–fuel mixture as rich as a relative fuel-to-air ratio (φ) of 0.5. However, to perform HCCI combustion successfully a leaner mixture having an equivalence ratio of more than 4 could not be employed. The large cyclic variation in IMEP values is the reason for this. It could also be a sign of an engine misfire. The intake air temperature was subsequently raised to 140°C, and the air–fuel mixture operating window for HCCI operation was set at 2.5–5.0.²⁴ The autoignition timing has advanced as the intake temperature has increased from 95°C to 125°C, and the combustion duration has lowered. Figure 7 demonstrates this.⁷² Within 2.5–4 bar of BMEP and 100–135°C of temperature, the engine can run in the biogas–diesel–HCCI (BD-HCCI) mode. A 40%–57% biogas energy ratio can be used at a charge temperature of 135°C.⁷³ With a lower inlet temperature, less NO_x is produced. This demonstrates that NO is temperature-sensitive, even at temperatures below the thermal NO limit.²² The temperature of the intake air plays a vital role in combustion. The intake temperature can be employed as an effective combustion phasing control variable. For a 10°C rise in temperature, while the other variables remain unchanged, the rates of common reactions increase by two to four times. Table 5 shows a summary of the SWOT analyses that were conducted.

Figure 7. Variation of cylinder pressure and ROHR under different intake and coolant temperatures with the same 50% burn point.74

Table 5 SWOT analysis of intake temperature on HCCI engines.

Strength	Weakness	Opportunity	Threat	Ref
– By increasing of intake temperature, the pressure in cylinder will increase – Increasing the intake air temperature accelerates exothermic reaction – As intake air temperature increases, heat release of low temperature exothermic reaction increases – As the intake temperature rises, so does the temperature of the compression phase and the fuel mixture, resulting in a more uniform fuel mixture and better combustion – NOx emissions will be reduced by lowering the temperature and delaying ignition timing	– Efficiency decreases as intake temperature rises, with the exception of operating settings with very low equivalence ratios of less than 0.2 – At low equivalence ratios (low load), the operation necessitates a rapid increase in beginning temperature, but efficiency falls as a result – When the intake temperature exceeds 140°C, the maximum value of cylinder pressure and heat release occurs before the TDC, especially with lower surplus air, and this might increase negative compression work, lowering the IMEP – At a temperature of 525 K, the ignition time would be sufficiently delayed that misfiring would occur.	– At a lower surplus air ratio, the IMEP is greatest at 140°C, and it diminishes as the intake temperature rises or falls – Higher the temperature of the incoming charge, higher is the IMEP with greater excess air ratio – A higher intake temperature raises the molecular activation energy of the mixture, which may boost chain reactions, speed up chemical reactions, and intensify burning – A higher intake temperature indicates that more energy is entering the cylinder and may be turned to work during the combustion phase – At higher intake temperatures, a faster chemical reaction rate and HRR can be obtained, potentially cutting the combustion time in half – A higher intake temperature may supply enough ignition energy for a lean fuel mixture and provide combustion stability in lean situations – A lean fuel combination with a suitable higher intake temperature could be a useful technique for improving HCCI engine combustion – In common reactions, the chemical reaction rate accelerates as the temperature rises	– As intake air temperature increases, cylinder charge decreases due to low air density – Initial temperature strongly affects peak pressures – As intake air temperature decreases, the ratio of low-temperature exothermic reaction decreases – Intake temperature significantly affects ignition timing – Increasing initial temperature or may cause combustion phasing to advance – By retarding the ignition timing, CO and HC emissions will be increased	[46, 48, 59]

Abbreviations: HRR, heat release rate; IMEP, indicated mean effective pressure; SWOT, strength, weaknesses, opportunities, and threat. Effect of valve timing

As NVO duration increases from 162°CA to 182°CA, the residual gas content rises by ∼12.5%. At this range, the variation of pressure, HRR, and predicted cylinder temperature are increased with the increase of NVO in the combustion zone. At this range, the temperature increases by ∼20 K, especially at inlet valve closing (IVC), and the period of combustion advances from 8.3 to −0.2° after TDC.⁷⁵ With the IVC angle of −143°ATDC, the CA of 10% heat release (CA10) is retarded from −24.2 to −23°ATDC and then to −21.4°ATDC, with the injected fuel mass rising from 31.2 to 62.4 and 93.6 mg. The increased injected fuel mass decreases the fuel–air charge's specific heat ratio. Hence, compression temperatures decrease with a delay in the ignition time. The larger global equivalence ratio of the mixture, owing to the injection of 93.6 mg results in a faster combustion reaction. The injection is performed at CA50, which is around 0.5 CAD higher than in the 62.4 mg fueling scenario. On the other hand, with a fixed IVC angle of −85°ATDC, CA50 of 93.6 mg fuel supply provides a delay of about 1.5 CAD in the CA50 concerning the 62.4 mg case.⁷⁶ The explorations on SWOT for different valve timings on HCCI engines are identified in Table 6.

Table 6 SWOT analysis of the effect of valve timing on HCCI engine.

[Table omitted. See PDF]

Abbreviations: IMEP, indicated mean effective pressure; NVO, negative valve overlap; SWOT, strength, weaknesses, opportunities, and threat. Effect of speed of the engine

The ability to adjust combustion phasing and prolong diesel HCCI operating ranges is affected by increasing engine speeds from 821 to 1727 rpm at high loads.⁷⁶ Further engine speed resulted in a decrease in maximum brake torque. It is discovered to be analogous to CI engines. With increasing engine speeds, the brake-specific CO and NO_x emissions values increase and decrease, respectively.¹⁴ According to Niklawy et al.,⁷⁷ at 200 N m load, the temperature increases from about 1000–1500 K at 2500 rpm during the low-temperature reaction and then increases to 2050 K at the end of the high-temperature reaction. The combustion phase is slowed, and the HRR reduces as the engine speed rises. Furthermore, the use of very high engine speed resulted in stifled combustion. However, as the engine speed increases, the rate of heat release in time increases, suggesting that the reaction is accelerated.⁷⁸ According to Sahin,⁷⁹ safe HCCI engine operation without knock and misfire can have within 0.45–0.55 of equivalence ratio (φ) value for the speed range of 800–1800 rpm. However, the addition of CNT (within 50–150 ppm) increases the equivalence ratio range (0.29–0.63) but reduces the speed range of 800–1400 rpm. The variations in engine speeds have the following SWOT effects on HCCI engines, as mentioned in Table 7.

Table 7 SWOT analysis of the effect of engine speed on HCCI engine.

Strength	Weakness	Opportunity	Threat	Ref.
– At higher speed the 50% AHR point gets retarded in a diesel HCCI engine – BMEP increases with speed upto a certain range of speed (1800 rpm) – At high engine speed, the heat release rate increases	– The results were obtained with fixed fuel injected quantity beyond that range brake torque again decreases – Increasing speed beyond a certain range retardeds the combustion phase and reduces the engine efficiency	– No information on the investigation of the HCCI engine at very late IVC of around −100°ATDC is available – Higher speed delays the combustion phasing and makes it easier to run under high load operation – At a wider range of speed, turbo-charging will cause special constraints – Consideration of VGT or electrically assisted turbochargers would be useful	– At very high speed, the 50% AHR point may shift towards the expansion stroke, which may reduce IMEP – In multi-cylinder HCCI engine, injector communication should be appropriate, otherwise, misfire will occur on some cylinder – Combustion may quench at a very high speed	[10, 49, 61]

Abbreviations: IMEP, indicated mean effective pressure; SWOT, strength, weaknesses, opportunities, and threat. Effect of injection timing

The higher the ON of the fuel from 25 to 90,⁸⁰ the initiation of ignition is later. Compared to conventional injection timing and using 10% reformed EGR (REGR) with ultra-low sulfur diesel (ULSD), combustion with delayed injection timing and using 10% REGR with ULSD reduces NO_x. However, smoke levels were consistent with the baseline engine operating (without REGR).⁴⁶ Advancement of the timing caused a reduction in the combustion rates when the engine was operated in the diesel-HCCI mode. Furthermore, even at a BMEP of 2 bar, delaying it causes a severe knock. HRR curves for various injection timings are presented in Figure 8.⁴⁴ Table 8 summarizes the SWOT analysis of different injection timings on HCCI engines. In another study by Roslan et al.,⁸¹ NO_x emission reduces with retarded port injection timing of ABE10 (10% of acetone–butanol–ethanol) run HCCI-DI, while HC increases for the same condition. This is primarily because of the reduction in the combustion duration for the retarded timing of the port injection. According to Panda and Ramesh,⁸² advancing the injection timing from 300°bTDC to 360°bTDC is found to increase HC and CO emissions but reduces NO_x and smoke emissions.

Figure 8. Effect of injection timing on heat release rate with a hydrogen energy ratio of 23.9%.40

Table 8 SWOT analysis of the effect of injection timing on HCCI engine.

Strength	Weakness	Opportunity	Threat	Ref.
– The use of delayed injection timing resulted in a 4% reduction in fuel usage – Tests with retarded injection timing and reformed EGR (REGR) with ULSD can reduce NOx emissions to that of the standard timing – Higher hydrogen supply enhanced the thermal efficiency – At a higher ON of fuel, the initiation of the ignition is significantly delayed – First-stage ignition to occur earlier for fuels with a high Cetane number	– Using 20% REGR had no significant effect on NOx and smoke for both engine conditions – An increase in NOx was seen when REGR was increased to 20%, and injection timing was delayed – With 20% REGR, both smoke and NOx were higher at high load. Retarding the fuel injection leads to producing knock even at lower BMEP – With later ignition time, UHC emissions grow almost linearly – A higher degree of smoke was noted due to a large amount of fuel impinges on the bowl surface of the piston	– To reduce NOx and smoke emissions, varied injection timing is needed for different REGR values – The fuel collides with the walls, resulting in increased HC emissions. As a result, injection orientation has a great deal of flexibility – As the equivalence ratio and the ON rise, the duration of second-stage combustion reduces	– Inappropriate REGR and injection timing adjustments can rise in smoke and NOx – The maximum achievable hydrogen energy ratio is reduced at larger loads and high EGR ratios – Details on HCCI engine performance at part loads are yet to be determined with different pilot fuel	[40, 42, 51]

Abbreviations: EGR, exhaust gas recirculation; SWOT, strength, weaknesses, opportunities, and threat. Effect of intake pressure

At low IMEP conditions, the rise of intake pressure leads to a drop in coefficient of variation of IMEP (COVIMEP) first, then COVIMEP decreases as intake pressure increases, and finally, COVIMEP increases with the further rise of intake pressure.⁸³ The phasing of the 50% AHR point is not affected by increasing the boost from the naturally aspirated level of 102.5 kPa (baseline) to 150 kPa and then to 200 kPa. For all three boost conditions simulated here, an IVC time of around −110°ATDC would result in a 50% AHR point around the TDC (0°ATDC). On the other hand, the 50% AHR curves show a considerable widening as the IVC is delayed from −143°ATDC to −100°ATDC.⁷⁶ A higher boost is needed to reach the same peak pressure at a lower compression ratio. Furthermore, the boost required to attain peak pressure is more than that of the lower peak pressure.⁸⁴ According to the calculations, raising the boost pressure from 1 to 2 bar slows combustion.⁸⁵ However, Solmaz et al.⁸⁶ reported that a rise in intake manifold pressure enhances volumetric efficiency, resulting in quick initiation of combustion at earlier CAs. It is responsible for a high rate of chemical reaction in the cylinder. As a result, maximum heat release and in-cylinder pressure increase. Polat et al.⁸⁷ also reported that CA50 was advanced up to 9°CA bTDC at 160 kPa of intake manifold absolute pressures and the combustion duration gets shortened. With 2-bar boost pressure, NO_x is one order of magnitude lesser than the naturally aspirated condition. Because a considerable portion of the gas forced into the gaps has already burned, it does not contribute to HC emissions as the pressure in the cylinder rises.²² Table 9 elaborates on the SWOT investigations performed at different intake pressures on HCCI engines.

Table 9 SWOT analysis of the effect of variation of intake pressure on HCCI engine.

Strength	Weakness	Opportunity	Threat	Ref.
– Higher the engine speed and load, higher boost pressure – Increasing boost pressure increases the relative air–fuel ratio – Turbocharging is liable for high pump losses – Even at lower intake temperature, increases of a boost to 1 and 2 bar have shown the peak cylinder pressure to rise by 11 and 18 MPa, respectively, albeit with a different fuel composition – Increasing boost pressure rises IMEP	– Higher boost is needed to hit the same peak pressure at a lower compression ratio – Increasing boost pressure decreases efficiency, but specific HC and CO increase at the same load – The increase in boost pressure from 1 to 2 bar retard the occurrence of combustion – For ethanol and natural gas, the use of boost pressure varies the combustion efficiency nonlinearly at CR of 17	– At high loads, the combustion time must be chosen as soon as feasible while staying within the limitations of peak pressure and rate of pressure increase – The turbocharger must be fitted with rotor speed monitoring capabilities. It will show where the engine is functioning in the turbo map – The performance is needed to be studied at different boost levels with the same fuel – Scope of using EGR with boost can be explored to improve IMEP	– Without heating of inlet air, the use of only intake boost by turbocharger may not keep CO emission down, especially at low loads – With or without boost, the simple kinetic model will always over-estimate the heat release during combustion – Increase of boost pressure from 1 to 2 bar has a minor effect on HC emission	[18, 63, 64]

Abbreviations: IMEP, indicated mean effective pressure; SWOT, strength, weaknesses, opportunities, and threat. Effect of EGR

The HC fuel was injected into a catalytic reformer attached to the EGR system and was investigated by Tsolakis and Megaritis.⁴⁶ The mixture of gas formed in it is recirculated to the engine in the form of REGR. Use of 10% REGR results in a drop in smoke and NO_x emissions. In the presence of ULSD, 20% REGR extends a decrease in smoke during engine run. At the same time, the NO_x is found to decrease only at lower loads. Asad et al. have come across the concern of the high reactivity of diesel fuel.⁸⁸ EGR is applied to delay the SOC. Higher values of EGR rates result in larger COVIMEP. Furthermore, COVIMEP increases with a rise in the EGR rate.⁸³ This is affected by the replacement of air by the EGR gases. This retarding effect is strong for the two-stage ignition fuels (PRF80 and PRF60) but quite weak for the single-stage fuels (iso-octane and gasoline). According to Asghari et al.,⁸⁹ the presence of EGR in the range of 36%–44% increases the chance of misfire even at a higher value of φ. The NO_x emissions obtained at different EGR levels are mentioned in Table 10. Gasoline showed the lowest sensitivity to “O₂.”⁹⁰ The employment of the EGR system reduced EGT by 12.8% and nitrogen oxide (NO_x) emission by 20% at full load conditions but showed an adverse effect on BTE, BSEC, as well as on emission parameters like CO₂, CO, UHC, and smoke opacity.⁹¹ Similar observations are reported by Bhurat et al.,⁹² where a partially HCCI engine produced 18% lower NO_x emissions along with a penalty in HC and CO emissions compared to a diesel run in the presence of 10% EGR. The maximum BTEs are observed to increase by up to 4% with EGR than without EGR.⁴⁴ The use of different EGR rates resulted following SWOT effects in HCCI engines, as mentioned in Table 11.

Table 10 Various NO_x emissions obtained by using exhaust gas recirculation.

Ref.	Fuel	Engine Specification	% EGR	Emissions (NOx)
Stanglmaier et al.42	ULSD	6.1 bar IMEP @ 1500 rpm	EGR: 10, 20	580, 510 ppm
			REGR: 10, 20	620, 720 ppm
Yousefi et al.65	ULSD	6.8 bar IMEP @ 1500 rpm	60.5, 66.8, 69.6	0.8, 0.108, 0.15 g/kWh
Aljaberi et al.66	ULSD, Algae Biodiesel, H2	0, 1.3, 2.6, 3.9, 5.2 kW	10%	400, 300, 180, 150, 130 ppm

Abbreviations: EGR, exhaust gas recirculation; IMEP, indicated mean effective pressure; ULSD, ultra-low sulfur diesel.

Table 11 SWOT analysis of the effect of exhaust gas recirculation on HCCI engine.

Strength	Weakness	Opportunity	Threat	Ref.
Maintaining COVIMEP below 3.5% reduces the threat of misfire	An increase in the rate of EGR raises COVIMEP	Use of COVIMEP at high load and high EGR	Combustion stability gets hampered due to lean mixture, resulting from a high EGR rate	[62]
The presence of EGR trims down EGT by 12.8% and NOx by 20% at full load	EGR is detrimental as far as BTE, BSEC, and different emissions, namely CO2, CO, UHC, and smoke opacity	Scope of exploration of engine performance with biodiesel-hydrogen HCCI mode with higher EGR rates (10%–60%).	Hydrogen enrichment results in higher EGT compared to DF	[66]
The reduced combustion efficiency at high EGR was aided by improving the combustion phasing at 66.8% EGR	With 61% EGR, the combustion is quite advanced (CA50:−11 ° ATDC), and the improper rate of pressure rise of ∼20 bar/°CA.	The high temperature of intake at high EGR can counter the NOx and combustion phasing benefits. A high EGR cooling is needed	The advanced combustion phasing (before TDC) reduces the thermal efficiency and limits the engine loads	[65]
Using real EGR can change the relative sensitivities of fuels used PRF80 is more sensitive to real EGR than iso-octane. The reverse occurs by using complete stoichiometric products (CSP)	The single-stage ignition fuels (iso-octane, gasoline) suffer a profound impact due to EGR cooling than two-stage ignition fuels (PRF80, PRF60) The fuels with two-stage ignition display better sensitivity to the drop of O2 in air due to EGR	The retarding effect of EGR changes with both fuel types and EGR This effect of EGR can be separated into thermodynamic and chemical effects	The presence of UHC, CO, and NO in EGR enhances the chance of suppression of autoignition	[52]

Abbreviations: EGR, exhaust gas recirculation; IMEP, indicated mean effective pressure; SWOT, strength, weaknesses, opportunities, and threat. COMBUSTION OF HYDROGEN IN THE PRESENCE OF DIFFERENT PILOT FUELS IN HCCI DIESEL ENGINE

The application of H₂ in the company of different pilot fuels has significant effects on the performance, combustion and emission characteristics of HCCI run of diesel engines. Different engine operating features, namely, the quantity of hydrogen induction, compression ratio, local and global equivalence ratio, temperature, and pressure intake, the timing of valve operation, engine speed, timing of injection of both pilot and primary fuels and last but not the least EGR are discussed extensively in the following sections.

Effect of percentage induction of hydrogen

The influence of the hydrogen to diesel energy ratio on the performance, combustion, and emission of HCCI engine running on diesel (DHCCI) is investigated at several brakes' mean effective pressures. It is seen that an increase in hydrogen input improves thermal efficiency by delaying the combustion process. At 0% hydrogen energy ratio, in DHCCI mode, maximum in-cylinder pressure, peak pressure, and maximum HRR occur.⁴⁴ This is due to the quick combustion of fuel mixtures within a short stipulated time. This, in turn, gives rise to an increase in pressure inside the cylinder and a higher HRR. Higher the hydrogen induction, the peak pressure and HRR decrease as hydrogen dilutes the mixture and retards the combustion process.⁴⁷ The addition of hydrogen in the HCCI engines has two effects: dilution effect and chemical effect. The dilution effect is owing to a decrease in the fuel-air ratio, whereas the chemical effect is attributable to the hydrogen and OH radical reactions. The addition of hydrogen in general delays the combustion process and reduces the combustion duration and for the same, the dilution effect is more prominent than the chemical effect. For an HCCI engine run with PRF-85, a combustion delay of up to 5°CA is observed by Rezgui.⁷⁰

For an HCCI engine run with hydrogen HRR is found to be quite high.²⁹ On one side hydrogen makes it possible to use a much leaner mixture in the combustion process. On the other side, a higher value of the maximum cylinder pressure (40% higher) and rate of pressure rise limits the engine operation to only part load. This also arise the need for design modification for the stability of the engine. Furthermore, inlet air preheating is necessary to promote the autoignition of HCCI engine fuel. The addition of hydrogen on HCCI engines fueled with natural gas does not change the operating range much however, 30% oxygen and its combination increase the operating range by 48% and 40.19%, respectively.⁷⁴ The addition of hydrogen starts SOC early and reduces the chance of ignition delay and localized combustion. The effect on the advancement of SOC is more for oxygen addition than that of hydrogen addition. Addition of hydrogen in an HCCI engine running on CH₄ as a primary fuel shows the advancement of ignition due to the fast combustion of hydrogen. Furthermore, the chemical reaction taking place also gets promoted due to the rise in the temperature of the reaction. Ignition gets advanced by 2°CA for a hydrogen induction percentage of 20%, whereas the peak in-cylinder pressure increases by 1.5% and 3% for a hydrogen induction percentage of 20% for 1000 and 1500 rpm operating speeds, respectively. Combustion duration (CA90-CA10) decreases by 12% with an increase in hydrogen induction of 15% due to a quicker reaction rate of hydrogen at all operating speeds.⁹³ Variations of HRR, in-cylinder pressure, combustion duration, and combustion phasing with percentage hydrogen induction are illustrated in Figures 9–12.

Figure 9. Effect of percentage induction of hydrogen on heat release rate.

Figure 10. Effect of percentage induction of hydrogen on in-cylinder pressure.

Figure 11. Effect of percentage induction of hydrogen on combustion duration.

Figure 12. Effect of percentage induction of hydrogen on combustion phasing.

Effect of compression ratio

The pressure inside the cylinder is largely dependent on the compression ratio value. Even a high power output value can be obtained using a higher compression ratio. The HCCI engine combustion largely relies on the mixture's homogeneity and changes at different compression ratios. Hence, it greatly influences the combustion characteristics of an HCCI engine. Rahbari et al.⁹⁴ described that a high compression ratio is useful in igniting the combustible mixture as it increases the temperature inside the cylinder. This can help the combustion of varying fuel quality in HCCI engines.⁹⁵ Yoon and Park investigated the combustion properties of an HCCI engine fueled with CH₄.⁹³ HCCI combustion, in general, needs a high compression ratio for the auto-ignition requirement of the combustible mixture. However, the use of high compression ratio results knocking phenomenon and is limited by the use of CH₄ with high octane value. Furthermore, a significant rise in the cylinder pressure is observed for all different operating speeds. The combustion duration gets shortened at a higher compression ratio up to a value of 18. The peak pressure is increased by almost 84% for a compression ratio of 20, whereas a 9% increase in HRR for a compression ratio of 16.

The effect of compression ratios of 10.7 and 14 using DME in an HCCI engine provided smokeless combustion and nearly no NO_x production. It is also discovered that the HCCI combustion process in the presence of DME has two-stage mode features at both compression ratios. However, it is more pronounced at a compression ratio of 10.7.⁵¹ The HCCI combustion within the compression ratios 15 and 20 demonstrated that a compression ratio of 20 delayed the ignition angle much past the 13°ATDC.¹³ In the presence of supercharging with a 3 atm intake pressure, each compression ratio of 14, 16, and 18 at 1200 rpm resulted in the indicated efficiency above 50% and NO_x emission below 100 ppm.⁹⁶ To get autoignition at TDC, the required compression ratio values for the operation with n-heptane, iso-octane, and gasoline are 11, 21.5, and 22.5, respectively. The best gross indicated efficiency is close to 43%, with a compression ratio of 21.5 while running on iso-octane.²⁶

Effect of equivalence ratio

A low equivalence ratio is suitable because the HCCI engine is designed to operate on a lean mixture. A higher equivalence ratio is limited by the absence of an oxidizer and impacts the combustion phasing. A higher equivalence ratio leads to higher chemical energy within the combustion chamber. As reported by Martinez-Frias et al., a high equivalence ratio may not necessarily mean high work output or high BMEP.⁹⁷ This is owing to the loss of energy in the conversion process. Najafabadi et al.⁵² have performed an experimental study to find the influence of equivalence ratio and intake air temperature on an HCCI engine. They concluded that the higher the equivalence ratio, the combustion phasing gets advanced along with the rise in peak in-cylinder pressure. The advancement in the combustion phasing is around 15°CA, with the increased value of the equivalence ratio. Pedersen and Schramm⁹⁸ varied the equivalence ratio along with the engine speed on an HCCI engine. They observed an advancement in the ignition timing when a richer mixture is used. This results in the rise in combustion temperature owing to the heating of the charges. However, heating is not so prominent when the mixture is lean to advance the ignition timing. So the ignition gets delayed when a lean mixture is used, as also observed by Tanaka et al.⁹⁹ They concluded that the low-temperature combustion phasing gets advanced with the rise in speed.

For an HCCI engine run with a natural gas increase in equivalence ratio results in the advancement of combustion timing due to higher fuel abundance in the combustion chamber. This further results in an increase in HRR and peak in-cylinder pressure.¹⁰⁰ However, a lower equivalence ration (below φ = 0.3) also results in a misfire. For a biogas run HCCI engine equivalence ration in the range of 0.25–0.4 is found to provide a better burning rate and results in an increase in temperature inside the cylinder.¹⁰¹

Effect of intake temperature and pressure

Wang et al. explored the impact of input parameters like air–fuel ratio and intake air temperature in controlling the combustion in an HCCI engine run by natural gas.¹⁰² The higher the intake air temperature, the higher the HRR and combustion pressure, as also observed by Yang et al.¹⁰³ The highest value of HRR and combustion pressure shift towards lower CA values with increased air temperature. The higher temperature of the intake charges causes faster oxidation of the natural gas mixture, thus causing a rise in pressure inside the cylinder and higher heat release. Moreover, an advancement in the mixture auto-ignition temperature is observed owing to the higher temperature inside the cylinder, causing the initiation of combustion of the fuel-air mixture before the piston reaches TDC. It is observed that with an augmented temperature of intake air from 180°C to 220°C, SOC shifted from 2.65°CA ATDC to −2.82°CA ATDC, and the duration of combustion advances from 6.15°CA to 4.02°CA. By increasing the overall gas temperature of 50 K, the HCCI combustion timing can be advanced up to 20°CA.⁵²

A study on the influence of stratification of incoming charge temperature on the HCCI combustion showed that under the given condition, the combustion processes demonstrated larger inhomogeneity with an inlet temperature of 125°C. Moreover, the in-cylinder pressure was lower at the larger stratification condition.⁷² Furthermore, the relative HCCI ranking of gasoline with a RON of 98 was constant at high intake temperatures with an intake pressure of 1.0 bar.¹⁰⁴ A two-cylinder revised port injection diesel engine was adapted to run on HCCI mode with ethanol as a fuel. The inlet air was supplied at 120°C, 140°C, and 160°C temperatures respectively. Results showed that the COV in the maximum pressure rise rate is less than 10% for all the test conditions and increases with the increase in inlet air temperature. It is observed that the average value of the maximum rate of pressure rise is more than 10 bar/CAD for some of the operating conditions and the maximum repeatability of CAD of a maximum rate of pressure rise up to 60% for all engine operating conditions.³¹ These researchers further experimented with varying intake air temperatures (120–150°C) and achieved the maximum ITE of 44.78% at an intake air temperature of 120°C. They also obtained the maximum IMEP of 4.3 bars. The maximum gas exchange efficiency obtained for ethanol HCCI combustion is 97.47%, and the maximum combustion efficiency obtained for ethanol is 97.45% for the given engine operating conditions.³²

Effect of valve timing

Variable valve timing (VVT) is the phenomenon, where a proper mechanism controls the valve opening and closing to control the supply of fuel. Thus it controls the performance and emission of the engine. It can be applied in the HCCI engine to successfully control engine combustion and emission. The engine compression ratio can be adjusted devoid of any alteration to the engine dimension with the help of the VVT approach by delaying the IVC time. A delay in IVC can effectively decrease the temperature inside the cylinder and ignition lag, improving the operating load. The NVO technique is extensively used to entrap the residual gases inside the cylinder, which helps in preheating the intake charges, thus improving the engine performance, combustion, and ignition characteristics.

The experimental investigation was performed on a single-cylinder HCCI engine running on a load of 3 bar and an operating speed of 2000 rpm.⁷⁵ Results showed that as the duration of NVO increases from 162°CA to 180°CA, the residual gas fraction increases by 12.5%, the temperature at IVC increases by 20°C, and combustion phasing advances from 8.3° to −0.2° after TDC. The simulation of variable intake valve timing on diesel HCCI combustion at varying load, speed, and boost pressure suggests that at the baseline fueling level of 31.2 mg, an IVC timing near −100°CA ATDC will produce a 50% AHR point near the TDC.⁷⁶ It is also revealed that for optimal combustion phasing near the TDC, with an IVC timing of −85°CA ATDC, the addition of boost pressure to 200 kPa increases IMEP from 4.35 to 5.13 bars.

Effect of speed of the engine

During the HCCI engine run, control of the combustion depends largely on the engine speed. With the increase in engine speed, the duration of combustion decreases. Hence, the energy loss through the wall also decreases. Furthermore, it is important to initiate the combustion early for a complete combustion process. The higher the engine speed, the lower the loss of energy through the walls and LTHR, as reported by Chang et al.¹⁰⁵ and Szybist and Bunting.¹⁰⁶ The maximum values of cylinder pressure and HRR increase at the higher engine speed of an HCCI engine fueled with methanol.¹⁰⁷ At higher engine speeds, the loss of energy during the compression stroke decreases due to a quick reduced period of cyclic operation. This leads to a reduced loss of energy through the cylinder wall. As a result, the temperature and pressure of combustion increase significantly. This is also responsible for the advancement in combustion phasing.

Experimental investigation of engine speed variation between 700 and 1300 rpm resulted in a maximum IMEP 419.31 at 138 kPa boost pressure and 1300 rpm engine speed.¹⁰⁸ It is also observed that the maximum COVIMPEP value of 2.48 was obtained in the 101 kPa inlet pressure condition at 1300 rpm. Also, it was revealed that the maximum BTE of 36.83% was obtained at 138 kPa intake pressure under the same engine speed. However, according to Aroonsrisopon et al., the HCCI engine is more reactive to intake temperature variation than the equivalence ratio at different speeds.¹⁰⁹ The effect of engine speed variations (N = 600 and 900 rpm) showed that the maximum deviation of low-temperature reaction duration, zone of negative temperature coefficient, and high-temperature reaction duration were 4.6°, 1.7°, and 3.2°, respectively, at higher speed.¹¹⁰ The influence of engine speed variation on the combustion properties of HCCI engines fueled with n-butanol was estimated by Iida et al.¹¹¹ They obtained the maximum IMEP in HCCI mode was 3.5 bar at 2000 rpm engine speed. The minimum ISFC achieved 175 g/kWh, comparable to a DI engine. Furthermore, Sjöberg and Dec reported that a higher intake temperature is needed for ethanol than gasoline to maintain CA50 at 372°CA at lower engine speeds.¹¹² It is also shown that HRR is maximum for gasoline at 392 rpm engine speed.

Effect of injection timing

Various experimental investigations are performed to measure the influence of different injection timing and duration on the HCCI engine working on multiple injection techniques. Results showed that in the diesel HCCI mode, thermal efficiency with MP fuel injection is higher (15%) compared to SP injection (11%) at an IMEP of about 3 bar.¹¹³ It was also revealed that the HC and smoke emissions are lower with MP injection by about 56% and 24%, respectively, compared to those of SP injection. It is also observed that the NO_x emission level is very low (about 6 ppm) in the case of the SP injection and is elevated to about 17 ppm with the MP injection. A much higher and more advanced HRR is obtained with multiple fuel injection techniques, which increases the temperature inside the cylinder and hence the NO_x level. For the same reason, higher NO_x was also observed by Sjöberg et al.¹¹⁴ Compared to the single fuel injection technique, higher thermal efficiency and lower fuel intake are achieved with multiple fuel injection techniques, resulting in better combustion phasing and greater HRR.

Studies were performed to understand the effect of second injection timing on the combustion and emission of an alcohol-gasoline blended HCCI engine corresponding to low and high equivalence ratios.¹¹⁵ Four different fuel blends of methanol and pure gasoline are used for experimental purposes. When injection timing is delayed beyond 25°CA, a decrease in P_max and maximum HRR occurs, leading to a delay in combustion phasing. It was again confirmed by incorporating the CFD study by Turkcan et al.¹¹⁶ In addition, a higher value of P_max and maximum HRR is achieved for fuel with a high equivalence ratio to that of a low equivalence ratio due to larger fuel combustion taking place at multiple locations throughout the cylinder. The effect of direct injection and EGR have been studied by Jang et al.¹¹⁷ in a DME-fueled HCCI engine to evaluate their influence on combustion characteristics. The starting point of heat release is retarded with direct injection than with port injection; however, the burning duration is the same for both injection techniques.

Effect of EGR

The HCCI combustion can be best regulated by applying EGR. Internal and external EGR are the two methods of doing it. Internal EGR can be achieved by modifying the duration of overlap of the exhaust valve, while external EGR can be achieved by regulating the EGR and the backpressure of the exhaust valve. The engine burning rate, ignition timing, and NO_x emission can be improved with the help of EGR, as suggested by Nakano et al.¹¹⁸; Li et al.¹¹⁹; Hassan et al.¹²⁰; Maurya and Saxena.¹²¹ The EGR has specific impacts on HCCI combustion, such as the impact on pre-heat, dilution, heat capacity, and chemical reaction.

As far as the low value of CO is concerned, it does not leave much effect on HCCI combustion. However, rise in the added quantity of CO, the peak pressure and HRR increase. This is owing to the increase in the chemical energy needed for the oxidation of CO. It is observed that the reaction rates of respective iso-octane, iso-butane, and propane with OH increase at the addition of 50 ppm NO by 28%–65%. Moreover, it is revealed that by adding 300 ppm NO, the reaction rate of iso-octane with OH decreases by 40%.¹²² The maximum IMEP of the fuel having ON of 80 is 0.58 MPa, while that of ON 40 fuel is 0.47 MPa, as shown by Yao et al.²⁷ The highest ITE up to 45% is observed for the PRF with an ON of 60. In another study, it is shown that with the increase of PRF ON, the IE increases, and the highest value of IE obtained is close to 45%, which occurs at ON above 60. It was also found that the increase of EGR delays the ignition timing, slows down the combustion reaction rate, reduces the temperature and pressure in the cylinder, and decreases the NO_x emission.¹²³ In similar studies by Dubreuil et al.,¹²⁴ it is shown that for the cool flame ignition delay, 100 ppmv of NO has the maximum effect and the change in CAD is of the same order as for 50% of EGR (about 2°CA) and for the main flame ignition delay, the effect of EGR is about twice that of NO. Furthermore, higher EGR has a considerable impact on soot formation. It first increases with an increase in EGR temperature and then finally decreases after a point. The best possible performance in terms of soot emission is achieved at an EGR percentage of 58%.¹²⁵

The cooled EGR has significant impacts on the combustion stage and duration. The high value of EGR reduces combustion speed and increases the combustion duration. It can be found that the EGR rate has a very slight effect for n-heptane and RON25, even the EGR increased up to 45%. While the CO emissions for RON50 and RON75 exhibit rapid increases with the increase of the EGR rate to 40%. Mainly, when the EGR rate increases up to 45%, partial combustion of RON75 occurs, and CO emission increase sharply.⁸⁰ Employing an exhaust gas fuel reforming technique influences the combustion characteristics of the HCCI engine.⁴⁶ In this technique, hydrogen-rich gas is produced online, then delivered into the cylinder as a REGR. The net HRR is found to increase at a higher percentage of REGR from 4 to 6.5 (Figure 13). This technique improves the engine performance and emission, thus reducing the emissions of harmful gases like smoke, NO_x, and so forth. The phenomenon is almost reversed with higher engine load, speed (4 bar IMEP, 1300 rpm), higher HRR, advanced combustion phasing, and a decrease in combustion duration, as represented in Table 12.

Figure 13. Effect of REGR on the heat release rate and cylinder pressure at different operating conditions.42

Table 12 Effect of reformed exhaust gas recirculation on combustion duration and ignition delay for different operating.⁴²

Engine operation	Premixed ratio (rp)	REGR (vol%)	Combustion duration (°CA)	Ignition delay (°CA)
3.6 bar, 1100 rpm	0	0	27	18
	0.18	6.5	24	18
4 bar, 1300 rpm	0	0	27	19
	0.12	4	25	20

CONCLUSIONS

The challenges associated with the HCCI engine combustion phenomena are reviewed and presented. The operating parameters have a strong role in deciding the efficacy of combustion. The dominant parameters influencing engine performance are the variations in percentage induction of hydrogen, equivalence ratio, piston bowl shape, intake temperature, valve timing, engine speed, injection timing, intake pressure, EGR, and compression ratio. Furthermore, there is growing research interest in investigating HCCI engines with diesel fuel to study combustion, emissions, and performance characteristics.

In general, the use of hydrogen in a diesel-run HCCI engine offers the appropriate distribution of combustion phases. This has the prospect of achieving improved efficiency and emission characteristics for conventional diesel engines. Hydrogen is proven to deliver significantly greater thermal efficiency because of its higher calorific value and ability to retard the combustion process.

Optimization of injection pressure is an important parameter that often remains unnoticed. Lower injection pressure results in inappropriate mixing of charges. Very high injection pressure results in over-penetration, and wets wall, thereby limiting combustion efficiency.

Some researchers reported that operating the HCCI engine at lower CR is helpful. This is because lower CR distributes combustion phasing appropriately, limits the maximum value of pressure rise rate and cylinder pressure within a controllable range, and broadens the operating range of the HCCI engine.

Harmful emissions such as nitric oxide will be limited, primarily because of combustion retardation and the lower value of HRR. Further, enriching the hydrogen quantity and running the engine at a higher air-fuel ratio without EGR provides better thermal efficiency. Furthermore, higher EGR rates also increase NO_x owing to the increased temperature of the intake charge. The one that remains a threat to adding hydrogen is the tendency of diesel to knock off an HCCI engine, which may get augmented due to the enrichment of hydrogen quantity. To minimize soot emission in the HCCI engine, multiple injections with optimized scheduling are necessary.

RECOMMENDATIONS

From the review of published research literature on HCCI engines, it is noted that further research investigations are necessary to understand the role of operating parameters such as compression ratio on the performance of HCCI engines. Analysis of appropriate combustion phasing of HCCI engines run with hydrogen enrichment at different compression ratios, both numerically and experimentally, has tremendous scope for research.

Research investigations and further analysis of data are required to understand the emissions and their variation with different amounts of hydrogen induction for the HCCI engine. As shown through the SWOT analysis in Table 3, further investigations are necessary to clearly understand the details of CO and HC emissions within a wide range of engine speeds and equivalence ratios. Further investigations are required to understand the propensity of knock for a diesel-based HCCI engine at various operating conditions, especially with enriching hydrogen quantity.