Fundamental phenomena affecting low temperature combustion and HCCI engines, high load limits and strategies for extending these limits

Progress in Energy and Combustion Science xxx (2013) 1e32

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Progress in Energy and Combustion Science journal homepage: www.elsevier.com/locate/pecs

Review

Fundamental phenomena affecting low temperature combustion and HCCI engines, high load limits and strategies for extending these limits Samveg Saxena a, *, Iván D. Bedoya b a b

Environmental Energy Technologies Division, Lawrence Berkeley National Laboratory, One Cyclotron Road, 90-2138, Berkeley, CA 94720, USA Grupo de Ciencia y Tecnología del Gas y Uso Racional de la Energía, University of Antioquia, Calle 67 No., 63-108 Medellín, Colombia

a r t i c l e i n f o

a b s t r a c t

Article history: Received 30 August 2012 Accepted 22 April 2013 Available online xxx

Low temperature combustion (LTC) engines are an emerging engine technology that offers an alternative to spark-ignited and diesel engines. One type of LTC engine, the homogeneous charge compression ignition (HCCI) engine, uses a well-mixed fueleair charge like spark-ignited engines and relies on compression ignition like diesel engines. Similar to diesel engines, the use of high compression ratios and removal of the throttling valve in HCCI allow for high efficiency operation, thereby allowing lower CO2 emissions per unit of work delivered by the engine. The use of a highly diluted well-mixed fueleair charge allows for low emissions of nitrogen oxides, soot and particulate matters, and the use of oxidation catalysts can allow low emissions of unburned hydrocarbons and carbon monoxide. As a result, HCCI offers the ability to achieve high efficiencies comparable with diesel while also allowing clean emissions while using relatively inexpensive aftertreatment technologies. HCCI is not, however, without its challenges. Traditionally, two important problems prohibiting market penetration of HCCI are 1) inability to achieve high load, and 2) difficulty in controlling combustion timing. Recent research has significantly mitigated these challenges, and thus HCCI has a promising future for automotive and power generation applications. This article begins by providing a comprehensive review of the physical phenomena governing HCCI operation, with particular emphasis on high load conditions. Emissions characteristics are then discussed, with suggestions on how to inexpensively enable low emissions of all regulated emissions. The operating limits that govern the high load conditions are discussed in detail, and finally a review of recent research which expands the high load limits of HCCI is discussed. Although this article focuses on the fundamental phenomena governing HCCI operation, it is also useful for understanding the fundamental phenomena in reactivity controlled compression ignition (RCCI), partial fuel stratification (PFS), partially premixed compression ignition, spark-assisted HCCI, and all forms of low temperature combustion (LTC). Published by Elsevier Ltd.

Keywords: Homogeneous charge compression ignition Low temperature combustion High load Power output Energy conversion Spark-assisted compression ignition

Contents 1. 2.

Characteristics of HCCI, spark-ignited, and diesel engines . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Fundamental phenomena affecting high load HCCI . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.1. Chemical kinetics and fuel properties . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.1.1. Low temperature heat release . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.1.2. Heat release at intermediate and high temperatures . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.1.3. Characteristics of single- and two-stage ignition fuels . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.1.4. Effects of molecular structure on fuel vaporization . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.1.5. 4-Sensitivity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.2. In-cylinder charge . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.2.1. Intake charge conditions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.2.2. Exhaust residuals . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

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* Corresponding author. E-mail address: [email protected] (S. Saxena). 0360-1285/$ e see front matter Published by Elsevier Ltd. http://dx.doi.org/10.1016/j.pecs.2013.05.002

Please cite this article in press as: Saxena S, Bedoya ID, Fundamental phenomena affecting low temperature combustion and HCCI engines, high load limits and strategies for extending these limits, Progress in Energy and Combustion Science (2013), http://dx.doi.org/10.1016/ j.pecs.2013.05.002

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S. Saxena, I.D. Bedoya / Progress in Energy and Combustion Science xxx (2013) 1e32

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4.

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2.2.3. Thermal stratifications . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.2.4. Mixture stratifications . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.3. Combustion timing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.3.1. Factors that determine combustion timing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.3.2. Strategies for controlling combustion timing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.3.3. Effects of combustion timing on engine operating parameters . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.4. Heat transfer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.4.1. Equivalence ratio and temperature effects on heat transfer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.4.2. Combustion timing effects on heat transfer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.4.3. Engine speed effects on heat transfer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.4.4. Intake pressure effects on heat transfer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.4.5. Engine geometry effects on heat transfer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.4.6. Ringing effects on heat transfer . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Emissions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.1. Unburned hydrocarbons and carbon monoxide . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3.2. Nitrogen oxides and soot . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Operating limits and practical considerations for high load HCCI . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.1. Typical load levels achieved in HCCI experiments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.2. Ringing limits . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.3. Misfire limits . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.4. Cyclic variability limits . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.5. Peak in-cylinder pressure limits . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.6. Excessive reactivity limits . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.7. Oxygen availability limits . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.8. Efficiency and emissions limits . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 4.9. Practical considerations for intake pressure boost in HCCI . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Strategies for avoiding & expanding the operating limits for HCCI and LTC . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.1. Delayed combustion timing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.1.1. Why delayed combustion timing avoids ringing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.1.2. Studies demonstrating the use of delayed combustion timing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.2. Partial fuel stratification . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.3. LTC with double fuel injections and distinct heat release stages . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.4. Spark-assisted HCCI/spark-assisted compression ignition . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.4.1. How SA-HCCI works, and fundamental phenomena affecting SA-HCCI . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.4.2. Why SA-HCCI increases load limits and constraints on high power SA-HCCI . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.4.3. Power output levels demonstrated in SA-HCCI . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.4.4. Other benefits of SA-HCCI . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 5.5. HCCI and LTC for power generation in advanced powertrains . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.1. Summary of important concepts . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 6.2. Promising research directions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . Acknowledgments . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . References . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . .

1. Characteristics of HCCI, spark-ignited, and diesel engines For health and environmental reasons, modern engines must meet increasingly tight emissions regulations for urban pollutants such as soot, particulate matter, nitrogen oxides, unburned hydrocarbons and carbon monoxide. Simultaneously, as society increasingly realizes the impact of CO2 on global warming, engines must achieve higher efficiency levels to minimize emissions of the global pollutant CO2. Traditionally, spark-ignited (SI) engines with 3-way catalysts have been effective at minimizing urban pollutants, while diesel engines have been effective at minimizing CO2. With modern aftertreatment technologies, diesel has also become effective at minimizing urban pollutants, however these aftertreatment technologies are currently very expensive. Low temperature combustion (LTC) engines are an emerging technology that allow high operating efficiency to minimize CO2 emissions and present an alternative to SI and diesel. HCCI is one type of LTC. HCCI simultaneously allows low emissions of soot and nitrogen oxides, and using oxidation catalysts HCCI can achieve low emissions of unburned hydrocarbons and carbon monoxide. As a result of the

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cleaner engine-out emissions, HCCI may provide a less expensive alternative than diesel for achieving low levels of both urban and global pollutants. Table 1 presents a comparison of the key characteristics of SI, diesel and HCCI engines. One fundamental defining characteristic of HCCI is how the fueleair mixture is prepared. HCCI typically uses port fuel injection (PFI) or early direct injection (DI) to allow a relatively homogeneous fueleair mixture. Emerging variations of HCCI, as discussed in Section 5.2 for example, use a combination of PFI and DI so that most of the fuel is well mixed while local regions of higher equivalence ratio exist. A second defining characteristic of HCCI is that the mixture is compression ignited. This is accomplished by using a combination of high compression ratios, high intake temperatures, high intake pressures or highly reactive fuels. Power output in HCCI is controlled by varying fuel flowrates directly while maintaining 4
1. The need for maintaining stoichiometric mixtures, as in SI, is unnecessary as 3-way catalysts are not necessarily required. As fuel flowrates are increased in HCCI to achieve higher power output, various strategies must be used to avoid ringing limits and other constraints (discussed in Sections 4 and 5).

Please cite this article in press as: Saxena S, Bedoya ID, Fundamental phenomena affecting low temperature combustion and HCCI engines, high load limits and strategies for extending these limits, Progress in Energy and Combustion Science (2013), http://dx.doi.org/10.1016/ j.pecs.2013.05.002

S. Saxena, I.D. Bedoya / Progress in Energy and Combustion Science xxx (2013) 1e32

Abbreviations

g 4 ATDC CAD CAign BDC BMEP CI DI DME EGR EKG EVC GDI HC HCCI HR HRR IMEP IMEPg

specific heat ratio fueleair equivalence ratio after top dead center crank angle degree crank angle of ignition bottom dead center brake mean effective pressure compression ignition direct injection dimethyl ether exhaust gas recirculation early kernel growth exhaust valve closing gasoline direct injection hydrocarbons homogeneous charge compression ignition heat release heat release rate indicated mean effective pressure gross indicated mean effective pressure

Given that HCCI relies on a premixed charge that is compression ignited, heat release rates in HCCI are governed by chemical kinetics [1e4]. As a result, heat release rates and pressure rise rates in HCCI are typically significantly higher than in SI and diesel engines, where flame propagation speeds or mixing and vaporization rates limit the maximum heat release rate. The high heat release rates in HCCI lead to ringing [5], one of the primary constraints governing the high power output limit. The authors believe that the traditional definitions of SI and diesel will be blurred as LTC engines develop, and engines seamlessly transition between operating modes.1 Throughout the article, HCCI is used as a basis for discussion of fundamental phenomena that are common across many forms of low temperature combustion (LTC) engines (and in many cases, even SI and diesel engines), with a focus on explaining concepts and techniques to improve the achievable load from HCCI engines. This article is intended for multiple audiences and caters to different levels of familiarity with engine technology. The article is written to be tutorial in nature and systematically review important concepts and technologies for the benefit of students and those who seek to learn about HCCI and LTC fundamentals and research directions. For more experienced readers (who may wish to skip to Sections 4 and 5), the article is intended to provide a comprehensive review of prior research in this field and point to some promising opportunities for future research and development. 2. Fundamental phenomena affecting high load HCCI In this section, the important physical phenomena governing HCCI operation are reviewed. An emphasis is placed on physics and chemistry occurring at the high load limits, however these phenomena influence the entire HCCI operating range. Chemical kinetics and important reactions governing HCCI operation and fuel characteristics are first discussed. The effects of variations in intake charge are then reviewed, including intake pressure, temperature, equivalence ratio, and overall charge composition. The importance

1 The authors welcome feedback or questions on this or other topics discussed in this article, or discussion of emerging engine or vehicle powertrain technologies.

IMEPnet ITHR IVO LES LTC LTHR NOx NVO PFI PFS Pin PPCI PPRR PRR PRF PVO RCCI RPM SCR SI TDC Tin

3

net indicated mean effective pressure intermediate temperature heat release intake valve open large eddy simulation low temperature combustion low temperature heat release nitrogen oxides, NO and NO2 negative valve overlap port fuel injection partial fuel stratification intake charge pressure partially premixed compression ignition peak pressure rise rate pressure rise rate primary reference fuel positive valve overlap reactivity controlled compression ignition revolutions per minute selective catalytic reduction spark-ignited top dead center intake charge temperature

of combustion timing is then reviewed, and significant details of the impact of combustion timing on engine operating characteristics are discussed. Finally in this section the impacts of heat losses are discussed, with significant emphasis placed on how heat loss influences engine efficiency over a wide range of operating conditions. 2.1. Chemical kinetics and fuel properties Given that the fueleair mixture in a HCCI engine is generally well-mixed prior to the onset of combustion, the fuel chemistry plays an important role in determining combustion characteristics [6e20] (more than in spark-ignited or diesel engines, where flame propagation [21e25] or mixing and vaporization [26e31] processes play the dominant role). Autoignition of different fuels occurs through a few categories of reaction pathways which can be characterized by the temperature range over which certain reactions occur. Although different fuels exhibit different extents of reaction within certain temperature ranges, the overall reactions that occur in a given temperature range are fairly similar between hydrocarbon fuels [32e37]. The fuel-dependent differences in the extent of reactions within a given temperature range can cause different fuels to behave drastically differently under the same conditions in an HCCI engine. Given that the chemical kinetics plays such an important role in HCCI engine performance, this section is dedicated to describing the reaction pathways followed by different hydrocarbon fuels during their oxidation process in a HCCI engine. Fuels for HCCI can generally be characterized based on whether they display single- or two-stage ignition behavior. Generally, fuels with lower octane ratings such as n-heptane, diesel, PRF80, and dimethyl ether (DME) display two-stage ignition [10,38e49] while fuels with higher octane ratings such as ethanol or iso-octane typically display single-stage ignition [10,11,40,45,50e53]. The difference in heat release behavior in a HCCI engine between a single- and two-stage ignition fuel is depicted in Fig. 1. Two-stage ignition fuels like PRF80 (which is a blend of 80% isooctane, and 20% n-heptane by volume) exhibit a small amount of heat release at temperatures which are lower than the H2O2 breakdown temperature (above 1000 K) [34,55,56]. This low temperature heat release (LTHR) causes a fraction of the fuel to be partially oxidized at temperatures below 850 K [12,34,55]. This

Please cite this article in press as: Saxena S, Bedoya ID, Fundamental phenomena affecting low temperature combustion and HCCI engines, high load limits and strategies for extending these limits, Progress in Energy and Combustion Science (2013), http://dx.doi.org/10.1016/ j.pecs.2013.05.002

4

S. Saxena, I.D. Bedoya / Progress in Energy and Combustion Science xxx (2013) 1e32

Table 1 Comparison of SI, diesel and HCCI engine characteristics. Spark ignited

Diesel

HCCI

Premixed

Non-premixed

Premixed

Ignition type

Spark ignited

Compression ignited

Compression ignited

Power output control

Airflow control, with near stoichiometric (4 z 1) airefuel ratio

Fuel flow control, with lean (4 < 1) airefuel ratio

Fuel flow control, typically with lean (4
1) airefuel ratio, or charge dilution

Mechanism controlling fuel burning rate

Flame propagation speed

Time for fuel vaporization and mixing

Chemical kinetics

Emission characteristics

Cleaner with 3-way catalyst. Higher CO2

Higher particulate matter, soot, NOx (without aftertreatment). Lower CO2

Higher unburned hydrocarbons, CO. Lower NOx, soot, particulates, and CO2

Fuel/air mixture type

LTHR can cause a temperature rise of between 10 and 20 K in the fueleair mixture within the combustion chamber [38,55,57]. LTHR can cause the hot ignition (shown near 375 CAD in Fig. 1) to occur earlier, or it can cause the requirement of lower intake temperatures for achieving hot ignition at a comparable timing as with a fuel showing single-stage ignition [6,10,58]. 2.1.1. Low temperature heat release Reactions for a straight-chained (normal) alkane will be used in illustrating the reaction pathways followed by a fuel undergoing LTHR, however similar processes are followed by many fuels which exhibit two-stage ignition behavior. At temperatures below 850 K, the primary reaction paths for heat release begin with hydrogen abstraction from the fuel molecules: At temperatures below 850 K, the alkyl radicals (such as the one shown in the products in Eq. (2.1)) can combine with oxygen molecules to produce an alkylperoxy radical [10,59]:

(2.1)

(2.2)

Fig. 1. Heat release rate from an HCCI engine using two different fuels, PRF80 which exhibits two-stage ignition, and iso-octane which exhibits single-stage ignition. Source: Sjöberg and Dec, 2007 [54].

The available bonding site on the outermost oxygen atom in the alkylperoxy radical (shown as a product in Eq. (2.2)) can then reach over and bond with a nearby hydrogen atom as part of an alkylperoxy radical isomerization reaction. The bond between the hydrogen atom and its attached carbon atom will break to leave an open reaction site on the main carbon chain [40,59,60]. There are several possibilities for which hydrogen atom is taken by the oxygen atom, and these different possibilities can be characterized by the intermediate transition state ring that is formed during the isomerization process. If the oxygen atom takes the closest hydrogen atom, a 5-member ring is formed (two carbon, two oxygen and a hydrogen atom). For each successively further hydrogen atom, a larger transition ring is formed. Several of these transition state rings are illustrated in Fig. 2: Once an isomerized alkylperoxy radical is formed, several possibilities exist for the next reaction. One important reaction path is where a similar reaction as in Eq. (2.2) occurs, where another oxygen molecule bonds with the open reaction site on a carbon atom located on the main hydrocarbon chain [10,40]. This process is illustrated for a 7-member transition state ring in Fig. 3: The molecule undergoing a second isomerization reaction (as illustrated in Fig. 3) can react in a branching reaction to release three radicals, two OH radicals, and the hydrocarbon molecule with open O sites. This set of reactions which finally leads to the branching reaction producing two OH radicals is responsible for the rapid heat release observed in the low temperature zone (less than 850 K) [34,38,55]. Other reaction paths are propagation reactions, and thus do not have a net effect on the radical concentration. The reaction of alkyl radicals with oxygen molecules (as in Eq. (2.2), or the reaction preceding the one shown in Fig. 3) is highly temperature and pressure dependent [61]. With higher temperatures, the equilibria of these reactions shift toward the reactants, and by roughly 850 K the addition of O2 to the alkyl radicals is almost completely extinguished [61]. As a result of the equilibria shifting toward the reactants, with higher temperatures the rate of

Fig. 2. Possible alkylperoxy radical isomerization sites, illustrating the different sizes of transition state rings. Source: Westbrook, Pitz, and Curran, 2007 [61].

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Fig. 3. Second isomerization reaction illustrated for a 7-member transition state ring. Source: Westbrook, Pitz and Curran, 2007 [61].

Fig. 4. 2-Methylbutane molecule illustrating locations of primary, secondary and tertiary carbon atoms.

heat release for the LTHR reactions is reduced [39,59]. This causes a region where higher temperatures actually cause lower heat release rates and can be seen in Fig. 1 for PRF80 between 345 and 350 CAD. Above 850 K, for most fuels very little heat release occurs until rapid H2O2 breakdown is encountered above 1000 K [61]. As a result, the majority of heating between 850 and 1000 K comes from piston work during the compression stroke [59].

bonds are where the carbon atom is bonded to only one other carbon atom. A secondary carbonehydrogen bond is one where the carbon atom is bonded to two other carbons, and a tertiary is where the carbon atom is bonded to three other carbons. Primary carbone hydrogen bonds are the strongest and thus they have the highest activation energy to prevent hydrogen abstraction [63]. As a result of the relative bond strengths for carbonehydrogen bonds, a fuel molecule which has more secondary carbonehydrogen bonds (such as a straight-chained alkane) is more likely to allow the abstraction reactions which lead to low temperature heat release [6,10e12,40,50,63e68]. The isomerization reaction illustrated in Fig. 2 occurs when the O2 chain which has bonded to an initially abstracted site reaches around to remove a hydrogen atom from another site. Due to the relative bond strengths, this hydrogen removal occurs more easily by breaking a secondary carbonehydrogen bond as compared to a primary [11,40,50,63,67,68]. Furthermore, the process of an O2 chain reaching to abstract hydrogen occurs more easily in long straight chain molecules because of the long chain length and flexibility of the molecule [69]. Long straight chained alkanes such as n-heptane (n-C7H16) are more conducive to low temperature heat release than branched alkanes such as iso-octane (iso-C8H18) [10,12,40]. This is because of the larger number of secondary bonds, the longer chain length, and greater internal flexibility [10,11,40,50,67,68]. It is for this reason that primary reference fuels with greater n-heptane content (and thus lower octane rating) tend to display a greater extent of low temperature heat release [6,38,55]. Using this knowledge of bond strengths and molecular structure we can analyze different fuel molecules. n-heptane, as already discussed, displays a great deal of low temperature heat release because of its large number of secondary bonds, its longer chain length, and its greater internal flexibility for isomerization reactions [6,43,55,63,70,71]. Iso-octane displays very little low temperature heat release because its branched structure creates more primary bonds making the initial hydrogen abstraction difficult leading to a decreased rate of isomerization [11,50,51,63,64]. Ethanol displays virtually no low temperature heat release partially because it has few secondary carbonehydrogen bonds and it has a very short chain length [52,72,73], but also because of chemical reaction pathways that are unique to alcohols (namely the concerted elimination of alpha-fuel radicals leading to an aldehyde þ HO2 [74]).

2.1.2. Heat release at intermediate and high temperatures Above 850 K, a significant set of reactions involving the fuel molecules are H-atom abstraction by OH (illustrated in Eq. (2.1)) or HO2 [61]. Abstraction with HO2 causes the production of hydrogen peroxide (H2O2) which gradually increases in concentration until the mixture reaches a temperature of approximately 1000 K [34,55,56,59]. Above 1000 K, the oxygeneoxygen bond in H2O2 breaks more readily causing the reaction in Eq. (2.3) [61], which rapidly releases a large amount of OH radicals. H2O2 / OH þ OH

(2.3)

As a result of Eq. (2.3) which proceeds more quickly above 1000 K, the previously stable hydrogen peroxide molecules act as a sudden source of OH radicals which consume the fuel molecules more quickly, leading to rapid heat release [34,40,55,56,59]. The breakdown of hydrogen peroxide into OH radicals and their subsequent reactions causes increasing temperature which further accelerates the breakdown of hydrogen peroxide [1,55]. This leads to an auto-ignition event that increases the temperature to above 1200 K where hot ignition takes over through the branching reaction of Eq. (2.4) [61,62]: H þ O2 / O þ OH

(2.4)

An interesting side effect from the importance of the hydrogen peroxide branching reaction (Eq. (2.3)) is that the intermediate and hot ignition temperatures are relatively independent of the fuel type. Almost uniformly, the intermediate ignition takes place when hydrogen peroxide breaks down near 1000 K, and the hot ignition takes place near 1200 K when Eq. (2.4) dominates. 2.1.3. Characteristics of single- and two-stage ignition fuels The hydrogen abstraction and isomerization reactions shown in Eq. (2.1) and Fig. 2 are critical in the low temperature heat release process (the first stage in two-stage ignition) [40,55]. Thus fuels that are more conducive to these reactions show a higher tendency of exhibiting two-stage ignition behavior [6,10,12,40,63]. The ease with which hydrogen abstraction and isomerization can occur is related to the molecular structure of the fuel molecule, and the type of carbonehydrogen bond that is to be broken [40,63,64]. Chemical bonds between carbon and hydrogen in a hydrocarbon fuel can be characterized by primary, secondary or tertiary bond strengths which depend on the neighboring environments of the carbon atom. As illustrated in Fig. 4, primary carbonehydrogen

2.1.4. Effects of molecular structure on fuel vaporization The environmental advantages of HCCI combustion (low NOx and soot emissions) are achieved using low combustion temperatures and a well-mixed fueleair charge. Fuel volatility is an important property for allowing a well-mixed charge when liquid fuels are used because it directly affects mixture formation. Fuel volatility is governed by intermolecular forces which are determined by molecular size and structure. Hydrocarbons are composed of molecules which are neither positively nor negatively charged in the ends. Therefore, there is not significant electrostatic attraction force caused by permanent dipole interactions between

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these molecules. The so called dispersion forces determine the molecular attraction in hydrocarbons. These kinds of attractions are produced by the change in electron density in the molecular periphery and its strength depends on the surface area. For higher surface area the intermolecular forces increase and consequently more energy is needed to achieve the normal boiling point. Molecular size determines the surface area, and as a consequence, fuel volatility decreases with increasing number of carbons for straight chained hydrocarbons. For the same molecular size, chain branching reduces the molecular surface area, thereby reducing the intermolecular forces and increasing fuel volatility [75]. Diesel is mostly made of straight-chained and lightly branched hydrocarbons, containing from 10 to 19 carbon atoms and boiling at in the range of 180e370  C [76]. Diesel exhibits low volatility and needs to be heated to near 200  C to avoid impingement in the intake stroke and be well mixed with the oxidant before the autoignition event in HCCI engines. Because of the high boiling point and high ignitability of diesel, DI is preferred instead of port-fuel injection for using this fuel in advanced combustion engines. 2.1.5. 4-Sensitivity The 4-sensitivity of a fuel is a measure of the impact of varying equivalence ratio on combustion timing when all other factors are held constant. Fuels that are highly 4-sensitive will experience larger changes in combustion timing as the equivalence ratio is changed. Prior research has shown that fuels exhibiting more twostage ignition behavior tend to be more 4-sensitive [77e80]. Sections 2.2.4 and 5.2 discuss how the 4-sensitivity of certain fuels can be used to extend the high power output limits of HCCI engines by intentionally creating mixture stratifications to cause more sequential ignition which avoids excessive ringing [77,79,81e86]. An important consideration in measuring the 4-sensitivity of a fuel is that when equivalence ratio is increased (up to stoichiometric) the residual and wall temperatures will also increase. It is important to isolate the equivalence ratio effects from the temperature effects to properly quantify the 4-sensitivity of a fuel, thus Dec et al. [78e80] developed a 19-1 firing cycle to accomplish this objective. In this strategy, 19 engine cycles use a controlled equivalence ratio, while the 20th cycle (where measurements are taken) uses the equivalence ratio of interest. Thus, by maintaining a fixed equivalence ratio for the 19 preceding cycles, the residual and wall temperatures will remain the same regardless of the equivalence ratio of the 20th cycle of interest [77]. Fig. 5 shows the 4-sensitivity of gasoline, iso-octane and PRF73 for a range of intake pressures. Fuels which have a strong negative slope exhibit increased reactivity (and earlier ignition) as equivalence ratios are increased, thereby making them highly 4-sensitive. In Fig. 5(a), for naturally aspirated conditions it is apparent that PRF73 is highly 4-sensitive, showing significant advancement in ignition timing with increased equivalence ratio [78]. Iso-octane

and gasoline at naturally aspirated conditions, however, are less 4-sensitive and do not show an increase in pre-ignition reactions with increased equivalence ratio [77,79,82]. As a result of the decreased specific heat ratio, g, with increasing fueling, gasoline and iso-octane show delayed ignition timing with increasing equivalence ratio [79,81]. Fig. 5(b) shows the 4-sensitivity for gasoline over a range of intake pressures. The results show that as intake pressure increases, gasoline becomes more 4-sensitive. This indicates that gasoline has pre-ignition reactions, primarily intermediate temperature heat release (ITHR), that increase with increasing intake pressure [11,79,81,87]. 2.2. In-cylinder charge 2.2.1. Intake charge conditions The intake temperature, intake pressure, equivalence ratio and charge composition have a significant effect on ignition timing, heat release and other important factors. The impact of variations in each of these factors is discussed briefly here. Intake temperature influences the in-cylinder temperature, and higher in-cylinder temperature increases the rate at which chemical reactions progress. Thus, higher intake temperatures will cause earlier combustion and faster heat release rates [64,88,89]. Intake temperature can be used as a control parameter for influencing combustion timing, and this technique has been demonstrated in many prior studies [72,81,90e94]. One approach for rapidly controlling the intake temperature is a technique known as fast thermal management, where control of intake air through different flow paths is used to change the intake temperature rapidly. This approach has been used by Saxena et al. and Martinez-Frias et al. in their studies [90,91,94], and a simplified schematic of their experimental system and control methodology is shown in Fig. 6. Higher intake pressures cause increased mixture reactivity when the same equivalence ratio and residual fraction are maintained [11,12,55,57,63,89,95]. Thus, higher intake pressures typically cause earlier combustion timing for a fixed intake temperature, or a lower intake temperature requirement to maintain fixed combustion timing (shown in Fig. 30) [11,12,14,57,63,88,89]. Higher intake pressures allow for the same equivalence ratio and dilution level to be achieved while increasing the absolute amount of fuel injected to each engine cycle, thus higher intake pressures allow higher load to be achieved [96,97]. For many fuels that exhibit LTHR or ITHR, increases in intake pressure cause a larger fraction of heat release to occur from LTHR or ITHR [55,79,81,98], also potentially causing increased 4-sensitivity with higher intake pressures [77,79,81e83]. In order to prevent excessive heat release rates which can damage an engine and cause excessive noise, HCCI uses diluted mixtures. Lean equivalence ratios are one way to achieve the required dilution however stoichiometric airefuel ratios can also be

Fig. 5. 4-Sensitivity of gasoline, iso-octane and PRF73, a) under naturally aspirated conditions, and b) at a range of intake pressures [79].

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lowered [11,12,14,55,57,63,64,88,89] (see Fig. 30). Under these situations, the use of external EGR is important for achieving these high power outputs without having to cool the intake air below ambient temperatures.

Fig. 6. Schematic and control methodology for intake temperature control of combustion timing.

used with EGR dilution. Lean ratios are discussed here and EGR is discussed in more detail in Section 2.2.2. Varying equivalence ratio impacts combustion characteristics in several ways, for instance through ignition timing, heat release rate, and wall and residual temperatures [48]. For highly 4-sensitive fuels, higher equivalence ratios cause earlier combustion timing (for 4
1), although the extent of equivalence ratio sensitivity of combustion timing is highly fuel dependent [77e80]. For fuels that are not very 4-sensitive, higher equivalence ratios can cause delayed combustion timing because more fuel results in lower specific heat ratios causing the requirement of more compression heating to reach ignition temperatures [99]. Fuels demonstrating two-stage ignition have been shown to be highly 4-sensitive, while single-stage ignition fuels do not readily show this 4-sensitivity [77,79,81e 83,86]. Another effect of higher equivalence ratio (for 4
1) is that it causes higher wall and residual temperatures which influences heat loss. 2.2.2. Exhaust residuals A second method to provide charge dilution for avoiding excessive heat release rates that can damage an engine is using exhaust residuals as non-reactive diluents. Two different EGR techniques can be used, which have almost opposite effects: 1) internal EGR, and 2) external EGR. In the internal EGR approach, hot exhaust gases are retained within the combustion chamber by closing the exhaust valve early during the exhaust stroke to cause a small recompression event during the NVO period [15,100e105]. The amount of retained EGR is varied by changing the EVC timing. With earlier EVC, more hot residuals are retained resulting in a hotter overall charge at BDC intake. Through this approach of controlling the BDC intake temperature through varying EVC timing, a variation of thermal control can be performed. One important consideration, however, is that the high temperature and increased residual mass fraction have counteracting effects on chemical reactivity and ignition timing. The increased non-reactive residual mass fraction will act to delay combustion timing [53,89,100,105e110], while the increased temperature will act to advance combustion timing [100e103,111,112]. Typically, the temperature effect will be the dominant factor [15,100e102,111,112]. In the external EGR approach, exhaust gases from the exhaust manifold are passed through a cooling device before being recirculated back to the intake manifold. As a result, the residual mass fraction effect can be used without the accompanying temperature effect as in internal EGR. The amount of EGR can be varied by changing the exhaust backpressure to cause more flow of exhaust gases into the intake manifold, with a backpressure valve for example. This approach of using external EGR for charge dilution is particularly important for achieving high load in HCCI for operating conditions at high intake pressures or with more easily ignitable fuels (thereby avoiding the excessive reactivity limit discussed in Section 4.6) [41,53,100,105,107,108]. With high intake pressures and high equivalence ratios to achieve high load, for example, the intake temperature to achieve delayed combustion timing is significantly

2.2.3. Thermal stratifications Thermal stratifications within the in-cylinder charge naturally occur for several reasons: 1) inhalation of charge with non-uniform temperature [113], 2) non-uniform heat transfer during the intake and compression stroke [85,113e118], 3) turbulent mixing for lowresidual engines [116], and 4) incomplete mixing between the fresh charge and hot residuals for engines with high levels of retained residuals. These temperature gradients cause different rates of chemical reaction throughout the charge and can cause sequential autoignition which is useful in lowering the maximum heat release rate [85,116,119e124]. A multi-zone model coupled with CFD analysis [125e127] can be used to study thermal stratifications and its effects on enabling more gradual heat release. Fig. 7 shows the temperature distribution computed using CFD analysis near the end of the compression stroke [127]. As Fig. 7 shows, the core fluid region near the center of the combustion chamber is the hottest and the airefuel mixture gets progressively cooler near the boundaries and in the crevices. These temperature gradients naturally occur as a result of heat loss to the cylinder walls, head and piston during the compression stroke [85,116e118]. In determining the effects of these thermal stratifications upon the overall heat release rate, it is important to consider the amount of mass present within each simulated zone. The hottest region will ignite first with subsequent ignition occurring in colder regions [79], however as seen in Fig. 7, the largest volume (and therefore the most mass) is encompassed within the hottest region. For illustrative purposes, Fig. 8 shows the zone temperature and zone mass fraction assuming the temperature map of Fig. 7 is divided into 40-, 20- and 10-zones grouped by temperature range [127]. Only the 10 hottest zones are shown in each case. As shown in Fig. 8, more mass is present in the hotter zones for HCCI engines [1,113]. It is interesting to note that a larger number of zones in a simulation allows better resolution of the thermal stratifications and the mass present in each zone. Fig. 9 shows the temperature profile of the 10 hottest zones in a 40-zone model [127]. As illustrated in Fig. 9, sequential autoignition occurs with the hottest zones igniting first. Since more mass is present in the hotter zones (Fig. 8), more heat will be released when earlier ignition occurs in the hot zones. Thus, these hotter zones have the largest impact on determining the maximum heat release and maximum pressure rise rate that influence the ringing intensity [127]. As a result of this sequential autoignition, any strategy that enhances the thermal stratifications, particularly near the hottest regions, should be effective at also lowering the ringing intensity (see

Fig. 7. Thermal stratifications at 3 CAD bTDC show that core region is the hottest. Intake conditions Pin ¼ 2 bar abs, Tin ¼ 438 K, 4 ¼ 0.29, for biogas fuel (60% CH4, 40% CO2).

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Zone Average Temperature [K]

1400

40 zones 20 zones 10 zones

1200

100

Solid lines = Temperature Dashed lines = Mass fraction

90 80

1000

70

800

60 50

600

40

400

30 20

200

Zone Mass Fraction [-]

8

10

0

0 0

1

2

3 4 5 6 7 8 10 hottest zone number

9

10 11

Fig. 8. Zone temperature and mass fraction for multi-zone model with 10, 20 and 40 zones at 3 CAD bTDC. Source: Bedoya, Saxena, Cadavid, Aceves, Flowers, Dibble, 2012 [127].

Section 4.2 for a detailed overview of ringing). This strategy has been explored in experimental studies, and at least four strategies can be used to artificially enhance the thermal stratifications: 1) increasing the heat transfer to the walls by using lower engine coolant temperatures [128], 2) increasing the heat transfer to the walls through more in-cylinder turbulence, with higher swirl for example [128], 3) using charge cooling from vaporization by using direct fuel injection at delayed injection timings [129], and 4) allowing more time for the development of naturally occurring thermal stratifications using delayed combustion timing. The first two strategies of promoting thermal stratifications through enhanced heat transfer to the walls were demonstrated as being effective in lowering the maximum heat release rates and ringing intensities. The higher rates of heat loss, however, also resulted in lower power output and lower fuel economy. When the fueling rate was increased to compensate for the lower power output, the benefits of enhanced thermal stratifications were lost [128]. The third strategy, the use of vaporization charge cooling from direct injection is less effective in lowering the peak heat release rates for fuels with single-stage ignition. However, for 4-sensitive fuels, the mixture stratifications that arise from direct injection can potentially play a much larger role in lowering the maximum heat release rate. The fourth strategy, the use of delayed combustion timing, is perhaps the most effective for allowing greater thermal stratifications and lower heat release rates [114,128,130]. Zone#29

1900

Zone#30

Temperature [K]

Zone#31

1700

Zone#32 Zone#33

1500

Zone#34 Zone#35 Zone#36

1300

Zone#37 Zone#38

1100

Zone#39 Zone#40

900 -10

0 10 20 Crank Angle [CAD aTDC]

30

Fig. 9. Crank angle resolved zone temperature showing sequential autoignition with hottest zones igniting first. Source: Bedoya, Saxena, Cadavid, Aceves, Flowers, Dibble, 2012 [127].

2.2.4. Mixture stratifications HCCI in its strictest sense has a completely homogeneous charge, meaning that there are no mixture composition stratifications. In reality, however, even PFI engines may contain mixture stratifications if there is inadequate time and turbulence for fuele air mixing. Alternatively, DI engines can use late injection timing to intentionally create mixture stratifications to take advantage of the charge cooling effects of late injection or the chemical kinetic effects of richer regions. Section 2.1.4 introduced the fuel property 4-sensitivity. A 4-sensitive fuel will have richer regions ignite earlier than leaner regions [79,85]. For these fuels, mixture stratifications using DI can be intentionally created to allow a sequential autoignition which can be used to achieve high power output without excessive rates of heat release and excessive ringing [79,81e83,86]. The use of mixture stratification in HCCI is commonly referred to as partial fuel stratification (PFS).2 Under PFS, which is discussed in Section 5.2, a majority of the fuel is injected in the intake manifold or directly injected in the cylinder very early to allow premixing, while a small amount of fuel (maximum 20%) is directly injected late to create areas with locally higher equivalence ratio. The effectiveness of PFS in lowering heat release rates and ringing intensity is highly dependent on the fuel’s 4-sensitivity [77,79,81e83,86]. Fuels with high 4-sensitivity are more conducive to PFS [77e80]. Fuels exhibiting two-stage ignition are more 4-sensitive and allow lower peak heat release rates with PFS [77,79,81e83,86], while fuels exhibiting single-stage ignition are not 4-sensitive and result in higher peak heat release rates and more NOx emissions when using PFS [77,79,81e83]. For 4-sensitive fuels, the use of direct injection fraction and direct injection timing can be used as control parameters for controlling the extent of fuel stratification and for tailoring the heat release rate under particular operating conditions. For fuels with two-stage ignition, PFS allows lower peak pressure rise rates (and lower ringing intensities) [79,81e83,86]. This occurs because regions with higher localized equivalence ratio tend to ignite earlier with subsequent ignitions occurring in regions with lower localized equivalence ratio [77]. It has been shown that the overall fuel reactivity, rather than the exact fuel composition determines the 4-sensitivity of a given fuel and therefore also determines its performance with PFS [78]. Fuels containing molecules with eCH2eCH2eCH2e chains, which allow 6-member low strain CeCeCeOeOeH rings (see Section 2.1.1) tend to display two-stage ignition and are more 4-sensitive [6]. The 4-sensitivity of fuels is not highly sensitive to engine speed, thus PFS can be used successfully over a wide range of engine speeds, however 4-sensitivity does depend on the intake temperature and intake pressure [78]. 2.3. Combustion timing 2.3.1. Factors that determine combustion timing Combustion timing in HCCI and many LTC engines is predominantly determined by the chemical kinetics. As a result, the entire time history of temperature, pressure, composition and other factors has an effect on determining ignition timing [13,32,131]. This complex interaction of parameters makes combustion timing more difficult to control in HCCI as compared with spark-ignited and diesel engines. Higher intake temperatures and intake pressures cause earlier combustion timing as they both cause faster chemical kinetics [12,64,88,89]. For 4-sensitive fuels, higher equivalence ratios (up to

2 Partial fuel stratification is also referred to in the literature by other names which describe very similar engine operating strategies, partially premixed compression ignition for instance.

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stoichiometric) tend to advance combustion timing because of the enhancement of charge reactivity. However, fuels without 4-sensitivity show an opposite trend, where higher equivalence ratios tend to cause more delayed combustion timing, because the lower specific heat ratio from higher equivalence ratios causes the requirement of more compression heating to reach ignition temperatures. However, in actual operation, for both kinds of fuels the use of higher equivalence ratios also increases the in-cylinder wall and residual temperatures, thereby causing advanced combustion timing [77,129]. In effect, for fuels with little to no 4-sensitivity there are competing effects where lower specific heat ratios will tend to delay combustion while higher in-cylinder wall and residual temperatures will tend to advance combustion. Fuel type can have a significant influence on ignition timing, even beyond 4-sensitivity. Fuels with significant amounts of LTHR or ITHR typically ignite earlier [12,40], while fuels showing only single-stage ignition typically ignite later. It is important to note that the temperature at which hot ignition occurs for different fuels is not substantially different [34,40,66], however the existence of LTHR or ITHR allows the in-cylinder mixture to reach this hot ignition temperature sooner, causing earlier ignition timing [6,40]. Fuels displaying LTHR and ITHR (see Sections 2.1.1 and 2.1.2) also have the capability of sustaining much later ignition timing before encountering misfire. This ability for achieving later combustion timing occurs because the low and/or intermediate temperature heat release helps to sustain a sufficiently hot in-cylinder temperature during the expansion process occurring after TDC (when incylinder charge expansion would normally cause charge cooling) thus allowing late combustion without misfire [86]. The use of EGR can cause different effects on combustion timing depending on the method for introducing EGR. The use of external EGR which has been cooled causes delayed combustion timing since the residual gases act as non-reactive diluents within the in-cylinder mixture [53,100,105,108e110]. Internal EGR, using NVO for example, can cause earlier combustion timing because the residual gases are at a much higher temperature [100e103,111,132]. Finally, engine speed also influences combustion timing. Higher engine speed allows less time for the chemical reactions to occur, however higher engine speed also results in less time for heat loss allowing a higher in-cylinder temperature. Typically, if all other factors are equal, higher engine speeds will result in delayed combustion timing. 2.3.2. Strategies for controlling combustion timing Two major categories of control techniques have been used in HCCI engines. The first relies on temperature control, and the second relies on fuel reactivity control. The temperature control techniques are discussed in Sections 2.2.1 and 2.2.2, and thus will only briefly be revisited here. Higher in-cylinder temperatures result in faster chemical kinetics and thus earlier combustion timings. One strategy for temperature control uses intake air heating, or heat recovery from exhaust gases. The second strategy for temperature control uses EVC timing to vary the amount of hot exhaust gases that are retained within the cylinder. Earlier EVC allows more hot residuals to be retained in-cylinder thereby creating a hotter in-cylinder charge and earlier combustion timing [101,102]. The fuel reactivity control technique is commonly referred to as reactivity controlled compression ignition (RCCI) [112,133e137] or dual-fuel HCCI. The RCCI approach relies on injection of two separate fuels, one which is highly reactive (with two-stage ignition for instance) and another which is less reactive (with only singlestage ignition for instance). By varying the relative amounts or timing of injection of the two fuels, combustion timing can be controlled. For example, in an RCCI approach where both gasoline and diesel fuel are port injected into an RCCI engine, combustion

9

timing can be advanced with injection of more diesel fuel to take advantage of its LTHR. Alternatively, a small amount of diesel fuel can be directly injected in-cylinder just before the desired combustion timing to create a pilot ignition which ignites the rest of the mixture, this approach is commonly called diesel LTC [138e143]. One recent innovation in HCCI control is the ability to use a reactivity controlled ignition, like RCCI, with a single fuel. One implementation of this concept for a gasoline-fueled HCCI engine has been demonstrated in recent studies [144e147]. The strategy relies on a combination of multiple direct injections and NVO to create fuel regions with different reactivity. Early EVC and late IVO are used to create a recompression event during an NVO period. A small initial DI event occurs during this NVO period, causing some fuel to react and reform due to the heat of the hot retained residuals. This forms a more reactive mixture of intermediate reactants. The intake valve then opens during the intake stroke to inhale air, and then the intake valve closes for the compression stroke. During the intake or compression stroke, the remainder of the fuel is added during a second DI event (although PFI can work too) into the combustion chamber. By controlling the relative amount of fuel injected in the first and second DI events, the overall fuel reactivity can be varied to change the combustion timing. More fuel during the first injection causes earlier combustion timing. 2.3.3. Effects of combustion timing on engine operating parameters Combustion timing is one of the most important factors influencing overall engine operating parameters. Combustion timing directly influences power output, efficiency, ringing intensity, heat transfer, combustion efficiency, emissions, and peak in-cylinder temperature and pressure. Having excessively early combustion timings, before TDC for example, causes excessive power losses, efficiency losses, high ringing intensity, higher heat losses, potentially higher NOx emissions, and higher in-cylinder temperature and pressure. Conversely, excessively delayed combustion timing can cause lower power and efficiency because of low combustion efficiency and unused expansion ability [128]. The low combustion efficiency for excessively delayed combustion timing results in high emissions of unburned hydrocarbons and carbon monoxide [128,148]. Detailed results exploring the influence of combustion timing on these parameters were presented by Saxena, et al. [90,91], and these results are briefly reviewed here. Figs. 10e13 show experimental results of various HCCI engine parameters over a range of combustion timings and equivalence ratios. All of these results are for an intake pressure of 1.8 bar absolute with an engine speed of 1800 RPM, using gasoline fuel. Fig. 10 shows the power output as gross IMEP for a range of combustion timings. Focusing briefly on the contour for 4 ¼ 0.45 it is apparent that power output follows a parabolic shape in relation to combustion timing, with the highest power output occurring at intermediate combustion timings. Lower power output occurs at early combustion timings because of negative work (if heat release occurs before TDC) and higher heat loss (discussed in detail in Section 2.4.2) [149]. With very late combustion timings, power output is lower because of incomplete combustion and unused expansion ability [128]. It is important to note that as the equivalence ratio is increased, the combustion timing at which the highest power output occurs is later because heat loss begins to play a more dominant role on reducing power output. For instance, by comparing the combustion timing at which the highest power output occurs for 4 ¼ 0.40, 4 ¼ 0.45 and 4 ¼ 0.50 in Fig. 10, it is apparent that the combustion timing for highest power output is later with higher equivalence ratios. Fig. 11 shows the ringing intensity for a range of combustion timings and equivalence ratios. Fig. 11 shows that delayed combustion timing can be used to avoid

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Maximum of mass averaged incylinder temperature (K)

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Combustion Timing, CA50 (deg. ATDC) Fig. 10. Experimental results of power output (Gross IMEP) for different equivalence ratios and combustion timings at Pin ¼ 1.8 bar and 1800 RPM [90].

Fig. 12. Experimental results of mass averaged peak in-cylinder temperature for different equivalence ratios and combustion timings at Pin ¼ 1.8 bar and 1800 RPM [90].

excessive ringing. Prior research [54,81,90,91,128] has shown the feasibility of using delayed combustion timing for enabling high power output while avoiding excessive ringing. This strategy is discussed in detail in Section 5.1. Fig. 12 shows an estimate of mass averaged peak in-cylinder temperature for a range of combustion timings and equivalence ratios. The mass averaged peak in-cylinder temperature is estimated using ideal gas law from the inducted mass and in-cylinder pressure measurements. The experimental results show that the peak in-cylinder temperature is lower with delayed combustion timing. The lower temperature with delayed combustion timing occurs because combustion occurs in a larger volume [109,128,148]. This lower temperature also results in lower heat loss, which is discussed in more detail in Section 2.4.2. Fig. 13 shows the combustion efficiency for a range of combustion timings and equivalence ratios. The experimental results show that combustion efficiency is lower with delayed combustion timing. The higher unburned hydrocarbon and carbon monoxide emissions that cause the lower combustion efficiency are caused by the lower in-cylinder temperatures (shown in Fig. 12). Fig. 14 shows simulated results of how thermal efficiency changes with combustion timing for cases where combustion occurs before TDC. Two plots are shown by including or excluding the effects of heat losses. The bottom line, showing thermodynamic losses only, shows that combustion between 0 and 5 CAD bTDC

does not have a significant impact on efficiency. The top line, which includes thermodynamic and heat losses, shows that when heat losses are included, combustion before TDC significantly decreases efficiency. Fig. 15 shows the change in thermal efficiency with delays in combustion timing. The results of Fig. 15 are gathered from a simulation which neglects chemical kinetics and heat transfer, thus isolating for thermodynamic efficiency losses. The thermodynamic efficiency losses with delayed combustion timing are caused by a lower effective expansion ratio leading to more available work being lost to the exhaust gases [99]. In this section, the factors affecting the combustion timing for peak power output were introduced. Fig. 10 shows that the highest power output occurs at intermediate combustion timings, and further delays in combustion timing should be used to produce the highest power output while avoiding excessive ringing as equivalence ratio increases. Power and efficiency losses occur at early combustion timing for two reasons, 1) increased heat loss due to higher in-cylinder temperatures and more time for heat transfer (discussed in Section 2.4.2), and 2) negative work if combustion occurs before TDC (shown in Fig. 14). Power and efficiency losses occur at late combustion timings for two reason, 1) an increase of incomplete combustion due to low in-cylinder temperatures (shown in Figs. 12 and 13), and 2) thermodynamic losses with delayed combustion timing (shown in Fig. 15).

0.45

2

Ringing Intensity (MW/m )

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12 10 0.40 8 0.35

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Ringing Limit

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2 0 0

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10

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Combustion Timing, CA50 (deg. ATDC) Fig. 11. Experimental results of ringing intensity for different equivalence ratios and combustion timings at Pin ¼ 1.8 bar and 1800 RPM [90].

Fig. 13. Experimental results of combustion efficiency for different equivalence ratios and combustion timings at Pin ¼ 1.8 bar and 1800 RPM [90].

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50%

Thermal Efficiency (no heat loss)

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No Heat Loss with Heat Loss

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S. Saxena, I.D. Bedoya / Progress in Energy and Combustion Science xxx (2013) 1e32

45%

65% 0

1

2

3

4

5

Combustion Timing (CAD bTDC) Fig. 14. Efficiency loss from negative work, with and without the effects of heat loss with Pin ¼ 2 bar, Tin ¼ 400 K, 4 ¼ 0.45, 1800 RPM, Twall ¼ Thead ¼ Tpiston ¼ 400 K, CR ¼ 17.

2.4. Heat transfer Heat losses from the hot in-cylinder gases to the cylinder walls, head and piston cause power and efficiency losses. Heat losses can be approximated with a modified version of the Woschni heat transfer coefficient [115,150]:

0:8 h
i C2 Vd TIVC 0:8 0:73 W=m2 K h ¼ C0 h0:2 P T S þ ðP  P Þ C m 1 p c 6 PIVC VIVC (2.5) In the heat transfer coefficient, h, in Eq. (2.5), hc is the instantaneous cylinder clearance height, P and T are the instantaneous cylinder pressure and temperature, Sp is the mean piston speed, Vd is the cylinder displacement volume, TIVC, PIVC and VIVC are the incylinder temperature, pressure and volume at intake valve close, and Pm is the instantaneous in-cylinder pressure for a motoring engine at the same conditions. Using this heat transfer coefficient, an estimate of the heat transfer rate can be calculated:

h i Q_ ¼ hAsurr ðT  Tsurr Þ W

(2.6)

In the heat transfer rate in Eq. (2.6), h is the heat transfer coefficient, Asurr and Tsurr are the surface area and temperature of the surroundings (i.e. piston, cylinder walls, and head), and T is the incylinder gas temperature. 66%

Thermal Efficiency

65%

11

In the following subsections, the effects of various engine parameters upon heat losses are discussed. The basis for these discussions is the heat loss equations in Eqs. (2.5) and (2.6), and a single-zone model of HCCI engine combustion using these heat loss equations (neglecting chemical kinetics3 by using the Otto cycle) is constructed to model the heat loss effects as different engine parameters are changed. By neglecting chemical kinetics, the model only captures thermodynamic and heat transfer losses so that these two effects can be studied independently of other engine characteristics. Combustion timing can be specified as an input independently of other operating parameters, and it is assumed that all heat release occurs instantaneously. 2.4.1. Equivalence ratio and temperature effects on heat transfer Equations (2.5) and (2.6) show that heat losses increase as the in-cylinder temperature increases. In-cylinder temperature typically increases at higher power output conditions. Specifically, higher temperatures are observed when using higher equivalence ratios (up to stoichiometric), earlier combustion timing, higher intake temperatures (assuming a fixed equivalence ratio), and potentially also depending on fuel composition (for example, with fuels that show low or intermediate temperature heat release). Fig. 16 shows a plot of the changes in heat loss with higher equivalence ratio (and therefore higher peak in-cylinder temperatures). Fig. 16 shows that as equivalence ratio is increased (while remaining fuel lean, 4 < 1), the amount of heat loss increases linearly. This increased heat loss is caused by higher in-cylinder temperatures. Interestingly, however, with increasing equivalence ratio below stoichiometric the efficiency losses from heat transfer decrease despite the increase in heat loss. This decreased contribution of heat loss on overall efficiency is because the increase in fuel energy with higher equivalence ratio outweighs the increase in energy loss from heat transfer. 2.4.2. Combustion timing effects on heat transfer Earlier combustion timing increases heat losses for two reasons: 1) higher in-cylinder temperatures, and 2) more time for heat loss to occur [115]. Higher in-cylinder temperatures are achieved with earlier combustion timing because the combustion occurs within a smaller volume, thereby causing both higher in-cylinder pressures and temperatures. Fig. 17 shows a plot of the increasing heat losses with earlier combustion timing for three different equivalence ratios. For a range of equivalence ratios, Fig. 17 shows that both the actual heat loss and the percent efficiency loss from heat transfer decrease with delayed combustion timing. As discussed in Section 2.4.1, Fig. 17 also shows that with increasing equivalence ratio the overall contribution of heat loss to overall efficiency losses decreases, despite an increase in the actual amount of heat loss. 2.4.3. Engine speed effects on heat transfer Engine speed has two separate effects which combine to influence heat transfer. First, the heat transfer coefficient, h (in Eqs. (2.5) and (2.6)), increases as RPM0.8 where RPM is the engine speed. Second, the time available for heat transfer decreases as RPM1. Thus, these two factors combine such that heat transfer varies with engine speed as Q NRPM0:2 [99]. The change in heat loss over a range of engine speeds is shown in Fig. 18.

64%

63%

62%

61%

60% 0

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15

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Combustion Timing (CAD aTDC) Fig. 15. Efficiency loss from delayed combustion timing (neglecting heat loss) with Pin ¼ 2 bar, Tin ¼ 400 K, 4 ¼ 0.45, 1800 RPM, Twall ¼ Thead ¼ Tpiston ¼ 400 K, CR ¼ 17.

3 Chemical kinetics are neglected in the single-zone model discussed here by simply input into the model the crank angle at which ignition occurs, and assuming that all heat release occurs instantaneously at the specified timing. Although this is an unrealistic scenario, it is useful for the heat loss discussions in this section.

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Fig. 16. Heat loss vs. peak in-cylinder temperature with Pin ¼ 2 bar, Tin ¼ 350 K, CAign ¼ 2 CAD ATDC, 1800 RPM, Twall ¼ Thead ¼ Tpiston ¼ 400 K, CR ¼ 17.

Heat Loss per Cycle (J)

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CAign) against the case with fixed CAign (varying Tin) shows that the heat losses are far more sensitive to changes in ignition timing than changes in intake temperature. Although the amount of heat loss increases linearly with intake pressure for all three cases, it is interesting to note that case 1 with fixed intake temperature (variable combustion timing) is the only one which shows an increasing contribution of heat loss to overall efficiency losses. All three cases experience higher heat loss from increased peak in-cylinder temperature because more fuel is present as intake pressure increases, however case 1 is the only one which also experiences higher heat loss due to earlier combustion timing from increased intake pressure. 2.4.5. Engine geometry effects on heat transfer As engine size increases, heat losses also increase since more surface area is available for heat transfer. However, the efficiency loss due to heat transfer decreases because the surface to volume ratio is also decreasing (thus heat transfer becomes a less dominant effect overall). To demonstrate the effects of engine geometry on heat transfer, Fig. 20 shows the change in actual heat loss and the efficiency penalty due to heat losses as engine size is increased. The

%Efficiency Loss from Heat Transfer

2.4.4. Intake pressure effects on heat transfer Intake pressure influences heat loss through several terms within Eq. (2.5). Higher intake pressures will cause substantially increased in-cylinder pressures, however the intake temperature requirement will decrease with increasing intake pressures [11,14,55,63,88,89]. This lower intake temperature causes lower incylinder temperatures during the compression portion of the cycle. After combustion, however, in-cylinder temperatures can be higher if equivalence ratio is maintained constant because of the release of more fuel energy. The complex effects of intake pressure are modeled for three scenarios in Fig. 19: 1) assuming intake temperature remains constant with changing intake pressure, thereby causing the combustion timing to change, 2) assuming intake temperature changes with varying intake pressure such that a fixed combustion timing is maintained, and 3) assuming that intake temperature and combustion timing both remain constant with changing intake pressure (an unrealistic scenario, but useful in isolating for the effects of intake pressure). The input conditions for the first two cases (namely intake temperature and combustion timing) with varying intake pressure are taken from Fig. 30, which is based on experimental studies on an HCCI engine by Saxena [90,91]. The results from Fig. 19 indicate that heat loss increases linearly with increasing intake pressure, however the rate of increase varies for the different cases. Comparing the case for fixed Tin (varying

Fig. 18. Heat loss vs. engine speed with Pin ¼ 2 bar, Tin ¼ 350 K, CAign ¼ 2 CAD ATDC, 4 ¼ 0.50, Twall ¼ Thead ¼ Tpiston ¼ 400 K, CR ¼ 17.

15

Ignition Timing (CAD ATDC) Fig. 17. Heat loss vs. ignition timing with Pin ¼ 2 bar, Tin ¼ 350 K, 1800 RPM, Twall ¼ Thead ¼ Tpiston ¼ 400 K, CR ¼ 17.

Fig. 19. Heat loss vs. intake pressure for three scenarios: 1) Fixed Tin ¼ 350 K with changing Pin, 2) Fixed CAign ¼ 2 CAD ATDC with changing Pin, and 3) Fixed Tin ¼ 350 K, and fixed CAign ¼ 2 CAD ATDC with changing Pin. For all cases, 4 ¼ 0.45, 1800 RPM, Twall ¼ Thead ¼ Tpiston ¼ 400 K, CR ¼ 17.

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smallest engine volume corresponds to the engine used in studies by Saxena [90,91], while the largest volume is similar to the engine used by Dec et al. [54,72,77,81,114,128,131]. For the changes in engine displacement on the x-axis of Fig. 20, the bore diameter was gradually increased while the stroke was maintained as being 1.2 times larger than the stroke. A compression ratio of 17 was maintained for all cases. 2.4.6. Ringing effects on heat transfer Prior studies have shown that heat losses in an HCCI engine are enhanced under conditions with ringing [151,152]. The increased heat loss is caused by the pressure waves coinciding with ringing causing a breakdown of the thermal boundary layer near the cylinder boundaries. In their study, Tsurushima et al. [151] suggest modifying the heat transfer coefficient (similar to the one shown in Eq. (2.5)) by adding a term that accounts for the pressure rise rate. Given that engines operating at boosted intake pressure conditions are more tolerant to increased pressure rise rate before showing ringing, further study is necessary to relate the heat transfer coefficient with the ringing intensity. 3. Emissions HCCI typically has low emissions of nitrogen oxides, soot, and particulate matter, and higher emissions of unburned hydrocarbons and carbon monoxide. Two particular characteristics of HCCI govern the emissions: in-cylinder temperatures and charge composition (discussed in Sections 2.2.1 and 2.2.2) [113,153e157]. 3.1. Unburned hydrocarbons and carbon monoxide

340

14.7%

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14.0%

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13.3%

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12.7%

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12.0% 0.55

0.65

0.75

0.85

HC fuels begins at roughly 1000 K [1,34,40,55,56,59,61], and Fig. 12 shows that several of the low equivalence ratio conditions at these operating conditions approach that threshold temperature. As a result, many of these lower equivalence ratio points have high unburned HC emissions (shown in Fig. 21). Fig. 22 shows the CO emissions for a range of equivalence ratios and combustion timings. CO-to-CO2 oxidation requires temperatures of at least 1500 K as OH radical concentrations are too low below this temperature [2,58,158,165e168]. As expected, CO emissions are low for early combustion timings and high equivalence ratios as in-cylinder temperatures are high enough to allow rapid CO-to-CO2 oxidation. The 4 ¼ 0.20, 0.25 and 0.30 plots in Fig. 22 show a parabolic trend with combustion timing that differs from the higher equivalence ratios. The trend shows that with very late combustion timings, the CO emissions for these lower 4 cases decrease. This behavior is caused by in-cylinder temperatures being near or below the 1000 K H2O2 breakdown temperature, where radicals released from rapid H2O2 breakdown (and subsequently radicals from hot ignition, Eq. (2.4)) cause HC molecules to break down to form CO (see Fig. 12) [1,34,108,154,159,161,167e169]. When the temperatures are near or below 1000 K, partial oxidation of HCs into CO does not occur thereby causing higher emissions of HCs and lower emissions of CO. Summarizing the HC and CO results from Figs. 21 and 22, it is clear that in-cylinder temperature has a significant impact on these emissions. Near or below 1000 K, unburned HC emissions are very high since the rapid breakdown of H2O2 leading to hot ignition does not occur quickly enough [34,125]. For this same temperature range, CO emissions are low as very little partial oxidation of HCs into CO occurs [1,158,162,170]. At a temperature range between 1000 K
T
1500 K, HCs are oxidized more effectively, but CO emissions are high since the CO-to-CO2 oxidation temperature threshold has not been reached [2,58,154,159,165,168]. Above 1500 K, both HCs and CO are oxidized more rapidly causing lower emissions of both these products [2,58,154,159,165]. For the lowest emissions of HCs and CO, conditions leading to the highest in-cylinder temperatures should be used, particularly high 4. For the remaining HC and CO emissions leaving the engine it is possible to use aftertreatment processes [171], with an oxidation catalyst for example. Oxidation catalysts are most effective at removing HCs and CO when higher exhaust gas temperatures are available. Fig. 23 shows exhaust temperatures for higher equivalence ratio cases for a range of equivalence ratios (with ethanol fuel). The results show that the use of delayed combustion timing

% Efficiency Loss from Heat Transfer

Heat Loss per Cycle (J)

In-cylinder temperature is a key factor in determining unburned HC and CO emissions characteristics [1,34,58,108,113,126,153,154, 158e164]. Lower temperatures generally cause higher emissions of unburned HCs and CO, however there are certain temperature ranges where HC emissions increase while CO decreases. Figs. 21 and 22 show the unburned HC and CO emissions for a range of equivalence ratios and combustion timings. The trends from the two figures indicate that both HC and CO emissions generally increase with lower equivalence ratios and delayed combustion timings as these factors cause lower in-cylinder temperatures (see Fig. 12) [109,128,148]. An important consideration for the HC and CO emissions characteristics is the hot ignition and CO-to-CO2 oxidation threshold temperatures. Prior research has shown that the rapid breakdown of H2O2 leading to hot ignition for

13

0.95

Displacement Volume (L) Fig. 20. Heat loss vs. engine displacement sizewith Pin ¼ 2 bar, Tin ¼ 350 K, CAign ¼ 2 CAD ATDC, 1800 RPM, 4 ¼ 0.45, Twall ¼ Thead ¼ Tpiston ¼ 400 K, CR ¼ 17.

Fig. 21. Experimental results of unburned hydrocarbon emissions for different equivalence ratios and combustion timings at Pin ¼ 1.8 bar and 1800 RPM [90].

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Soot is an important concern for diesel engines. It is formed in the rich regions of a fuel spray when fuel breakdown begins while insufficient oxygen is available. Given that the fuel is well mixed prior to ignition in an HCCI engine, negligible levels of soot emission occur. In diesel engines, the high emissions of NOx and soot require the use of expensive aftertreatment systems [26,27,134,170,171,176e 178]. Modern diesel vehicles use a combination of aftertreatment technologies, including lean NOx traps, SCR devices, particulate filters and/or oxidation catalysts [171,179e182]. These aftertreatment systems are extremely expensive and can cost as much as the diesel engine itself [171,182]. Thus, one significant benefit of HCCI is that it can achieve efficiency levels (and thus low CO2) comparable with diesel engines [99,183,184], without need for expensive NOx and soot removal aftertreatment systems [70,109,110,185e188].

Carbon Monoxide (ppm)

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Combustion Timing, CA50 (deg. ATDC) Fig. 22. Experimental results of carbon monoxide emissions for different equivalence ratios and combustion timings at Pin ¼ 1.8 bar and 1800 RPM [90].

4. Operating limits and practical considerations for high load HCCI 4.1. Typical load levels achieved in HCCI experiments

allows higher exhaust temperatures to be achieved, thereby making an oxidation catalyst more effective at removing HCs and CO. For higher equivalence ratio cases, this delayed combustion timing is also critical for avoiding excessive ringing [54,148,151] and excessive heat losses [115], thus there are multiple benefits for using delayed combustion timing. 3.2. Nitrogen oxides and soot NOx and soot emissions are grouped together in this section because they are both low in an HCCI engine. Nitrogen oxide emissions, NO and NO2, are predominantly formed through the Zeldovich chemical path through the breakdown of the strong triple bond between nitrogen atoms in N2 [156,172e174]. This chemical pathway has an activation temperature near 1800 K [175]. Given that HCCI takes advantage of highly diluted charges, with either lean mixtures or high fractions of exhaust residual gases, peak in-cylinder temperatures remain low in comparison to diesel or spark-ignited engines. As a result, numerous studies [72,79,81,90,91,176] have shown that high power output (sufficient for passenger transportation applications) can be achieved from HCCI engines while producing very low levels of NOx emissions, well below the US2010 limits of 0.27 g/kWh.

Exhaust Temperature (°C)

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Combustion Timing, CA50 (deg. ATDC) Fig. 23. Experimental results of exhaust temperatures for different combustion timings at high equivalence ratios, 1800 RPM with ethanol fuel.

In internal combustion engines power output is determined by load and engine speed. For a fixed engine speed, different load levels are achievable in spark ignited, diesel and HCCI engines and each engine type is controlled using different methods. Engine load is usually reported using the IMEP or BMEP because these variables allow comparison between different engine sizes and technologies. In automotive applications, the maximum BMEP is around 12 bar for light duty SI engines and 18 bar for light duty CI engines [189,190]. For heavy duty CI engines for road applications the maximum BMEP can be as high as 25 bar [191]. Maximum load for HCCI engines tends to be considerably lower when compared with SI and diesel engines because the highest load is constrained by the limits discussed in Sections 4.2e4.9. The high load limit can be defined as the IMEP (gross or net) at which various limits are encountered, particularly the ringing limit. In naturally aspirated HCCI engines the maximum IMEP is close to 5 bar operation in a speed range of 1000e1500 RPM [81,109,192e 194]. This load only represents around 40% and 30% of the achievable load in SI and diesel engines respectively. Additionally, the maximum load is not conserved at all operating engine speeds leading to a very limited HCCI operating window in automotive applications, as shown in Fig. 24 [195]. Pioneering research in HCCI combustion has targeted extension of the high load limits as one of the most important challenges for widespread utilization of this technology. This challenge remains even today, however Section 5 of this article reviews many promising strategies to extend the high load limits. In early studies developed at the Lund Institute of Technology the achievement of 14 bar IMEPnet at 1000 RPM [96], 16 bar IMEPg at 1000 RPM [196], and 16 bar BMEP at 1800 RPM [197] using boost pressures close to 2 bar and compression ratios between 17:1 and 19:1 was reported. From the Heavy Truck Clean Diesel (HTCD) Program, supported by US-DOE, it was reported that a load level of 20 bar BMEP was achieved while operating at three engine speeds (1200, 1500, and 1800 RPM), boost pressure around 3 bar, and a compression ratio of 8:1 [198]. The mentioned studies defined high load limits based on the maximum pressure rise rate (PRR), however the allowable PRR depends on operating variables such as engine speed and boost pressure, therefore, the study lacked a more suitable method to define the load limit. For recent research accomplished by Dec et al. at Sandia National Laboratories [79,81], load limits are defined using a metric called ringing intensity (RI) rather than maximum PRR as criterion for safe operation. RI (see Section 4.2) is more useful to define the

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maximum IMEP because it compares the PRR to peak in-cylinder pressure and engine speed, producing a more accurate metric of tolerable combustion noise. Using delayed combustion along with a boost pressure of 2.25 bar, compression ratio of 14:1 and engine speed of 1200 RPM, the highest IMEPg observed was 16.34 bar [81]. Using the same experimental setup, but a lower boost pressure of 1 bar, partial fuel stratification (see Section 5.2) allowed the high load limit to be extended up to 13 bar IMEPg [79]. Many of the above mentioned studies in this section are based on extending the load limits for single cylinder engines limited to a narrow range of engine speeds and using external compressors or superchargers. For operation more representative of a vehicle, studies are conducted on multi-cylinder engines and over a wider range of engine speeds. Under these conditions and with turbocharging, Hyvönen et al. [199] and Johansson et al. [200] showed that the high load limit can be further constrained. The following sub-sections discuss the main limits constraining high load HCCI operation. 4.2. Ringing limits Ringing in HCCI engines is the occurrence of pressure waves that can cause an undesired audible noise in extreme cases [201]. These pressure waves can damage an engine over time [202,203]. Ringing occurs in conjunction with excessive pressure rise and heat release rates [204e206], and thus steps must be taken to avoid the harsh combustion regimes that cause these rapid pressure rise rates. Fig. 25 shows an in-cylinder pressure trace for an operating condition which exhibits intense ringing. The pressure oscillations observed between 1 and þ10 CAD are the physical manifestation of ringing. The pressure oscillations shown in Fig. 25 are caused by pressure waves which propagate throughout the combustion chamber [202,204]. Although HCCI is idealized as being a uniform autoignition event of the entire in-cylinder airefuel mixture at once, in reality ignition occurs sequentially over a short time at different spatial locations throughout the mixture [84,202,207,208]. These sequential ignition events are caused by differences in localized temperatures which arise during the compression stroke. Laser diagnostic studies in optical-access HCCI engines [114,130] and CFD studies [113] have quantified the spatial distribution of hotter and cooler zones in motoring and firing engines. These studies suggest that the temperature difference between the coldest and hottest

Fig. 24. Speed and load range in New European Driven Cycle (NEDC) for a fourcylinder, hybrid SI-HCCI engine. Tests were done in a test bed which simulates speed and load engine regime as on a real chassis dyno test. The intensity scale of the map represents the consumed fuel mass. The different operation modes are highlighted by their operation borderlines. The 100 km/h point, the 120 km/h-point as well as idle and a lot of the accelerations are outside the HCCI operation range [195].

Pressure (Bar)

S. Saxena, I.D. Bedoya / Progress in Energy and Combustion Science xxx (2013) 1e32

15

130

110

90

70

50

30 -20

-10

0

10

20

30

40

Crank Angle Degree (ATDC) Fig. 25. In-cylinder pressure for a condition exhibiting intense ringing.

zones in the core fluid at TDC can be as large as 100 K. These thermal gradients typically do not arise until the last 20e30 crank angle degrees before TDC [114,116,118,121]. The fact that auto-ignition in HCCI engines occurs sequentially from the hottest to the coldest portion of the charge contributes to decrease the peak pressure rise rates which could be even higher in a thermally homogeneous charge. Therefore, thermal stratifications in the core fluid region of the combustion chamber allow the use of higher equivalence ratios to achieve higher power output before encountering the ringing limits when compared with a truly thermally homogeneous charge prior to auto-ignition [85,116,119,120]. Ringing coincides with higher load conditions; however, sequential ignition from thermal gradients will likely arise at all operating points (including lower loads) [116]. A prior study [207], using chemiluminescense images of n-pentane auto-ignition in a Rapid Compression Machine (RCM), suggests that the ringing observed in HCCI engines at high loads originates from the localized development of the hot stage of ignition. “Hotspots” randomly localized throughout the combustion chamber auto-ignite, leading to high temperature combustion zones. The interaction between the high temperature zones may be critical for creation of pressure waves causing ringing. Furthermore, a study [209] explored the origin of knock for iso-octane fueled HCCI combustion using threedimensional CFD simulations (including LES) coupled with detailed chemical kinetics. It was found that when the pressure trace begins to fluctuate as evidence of high ringing intensity, multiple “hotspots” appear in the core region of the charge (T > 2000 K) leading to increased heat release rates and local pressure fluctuation. As the rest of the charge combustion takes place at typical temperatures for HCCI conditions (T < 1800 K), the heat release rates and pressure rise rates are lower than in the hottest zones. The multi-spot pressure discontinuity creates interactive pressure waves that lead to high HCCI ringing. Experimental studies show that the pressure oscillations indicative of ringing, however, do not arise significantly at lower load conditions (with lower fuel content) [90]. At low load conditions, there is unlikely to be enough acoustic energy released from the ignition of localized hotspots to cause the interaction of pressure waves that leads to ringing at higher load conditions. Additionally, at low load conditions the pressure waves from sequential ignition may be dampened enough to make them undetectable with in-cylinder pressure sensing. At higher load conditions (with higher fuel content), however, enough acoustic energy is released from these sequential ignitions to cause detectable pressure waves. Furthermore, at higher load conditions, it is possible that there is

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constructive interference where the pressure wave from an early localized ignition causes ignition in different locations in a way that amplifies the early pressure waves [206]. Finally, it is possible that the higher equivalence ratios at high load conditions influence the temperature sensitivity of fuel reactivity to further facilitate sequential ignitions which cause ringing. Ringing differs from knocking as observed in SI because knocking is caused by end-gas autoignition. In HCCI, autoignition is desired and the fuel-in-air mixture autoignites relatively uniformly as compared with the flame propagation that occurs in SI. Both knocking and ringing cause pressure waves that can be measured using in-cylinder pressure transducers, however ringing typically has the majority of its acoustic wave energy in the 5e6 kHz frequency range, while knocking contains more energy in the 8e 25 kHz range [203,210,211]. The amplitude of pressure pulsations in HCCI can be an order of magnitude larger than knocking in SI engines, however cylinder linings can significantly dampen wave energies below 8 kHz [5]. This dampening effect can make it more difficult to detect ringing using externally mounted sensors. A well accepted correlation4 to quantify ringing behavior using in-cylinder pressure measurements was derived by J.A. Eng (shown in Eq. (4.1)). Numerous studies have demonstrated the ability of the ringing intensity correlation in Eq. (4.1) to quantify the extent of ringing over a wide range of engine operating conditions [5,72,81]. An acceptable limit for the maximum allowable ringing intensity used in some studies is 5 MW/m2 [72,81].

 1 RIz 2g

dP dt max Pmax

b

2 pffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi gRTmax

(4.1)

The primary method for avoiding ringing in an HCCI engine is to use highly diluted mixtures [128,205,212]. This can be accomplished with high fractions of EGR, or with lean equivalence ratios. The side effect of using only diluted mixtures to control ringing is that it causes lower power output as compared with diesel or SI engines, and thus other strategies (aside from only charge dilution) must be used to control ringing while using less diluted mixtures to increase power output. One alternative strategy is to use highly delayed combustion timing [70] since there is a higher rate of charge expansion due to faster downwards piston motion for a given engine speed at later crank angles after TDC. The faster charge expansion at later combustion timings counteracts the rapid heat release (and pressure rise) from autoignition, thereby causing lower ringing. Another strategy for reducing peak heat release rates (and therefore ringing intensities) is to use moderate amounts of temperature or mixture stratifications. Temperature stratifications cause spatial variations in the rates of chemical reaction to allow the hotter spatial zones to autoignite prior to the cooler zones [85,119,120,124,128]. Mixture stratifications can have a similar effect as temperature stratifications for creating more sequential ignition, however this strategy is only effective for fuels which exhibit 4-sensitivity (usually fuels with LTHR or ITHR) [77]. Temperature stratification can be difficult to control in a production engine, however mixture stratifications can be easily achieved in an engine with DI by using multiple injections or late injection strategies. The use of delayed combustion timing, temperature stratifications and mixture stratifications are discussed in more detail in Sections 5.1 and 5.2.

4 At the time this review article is being published, the authors are aware of a handful of unpublished studies that seek to improve the ringing intensity correlation or propose a new correlation to define ringing limits. The use of the b term to relate maximum pressure rise rate with the amplitude of pressure fluctuations is one particular area that may require improvement.

4.3. Misfire limits Misfire occurs when insufficient energy is provided to allow the fuel to reach its hot ignition point (see Section 2.1.2 for an explanation of hot ignition). As a result of the failure to ignite, in-cylinder pressures after top-dead-center remain relatively low, resembling a motoring pressure trace. Little to no power is produced from a misfiring cycle, and a large amount of unburned fuel exits the engine through the exhaust manifold. Fig. 26 illustrates the difference between burning cycles and misfiring cycles. Misfire introduces a further constraint upon maximum load. As will be discussed in Section 5.1, the use of delayed combustion timing is very important for avoiding ringing and excessive efficiency losses from heat transfer. In some cases, it is desirable to delay the combustion timing to be near the point of misfire. In these cases a given cycle can misfire because charge expansion with downwards piston motion produces a quenching effect on the chemical reactions prior to the occurrence of hot ignition [213]. As discussed in Section 5.1, the ITHR behavior of gasoline is critical for allowing delayed combustion timing since it counteracts the cooling effect from charge expansion with downwards piston motion after TDC. For gasoline, this ITHR behavior increases with lower intake temperatures and with higher intake pressures up to a maximum at 1.8 bar (absolute) [72,81,86,87,99,176,214]. 4.4. Cyclic variability limits HCCI was initially considered as having low cycle-to-cycle variations when compared with SI. However, more detailed studies revealed that cyclic fluctuations of the most important combustion parameters can be quite large under certain operating conditions [148,215e217], especially at highly delayed combustion. As mentioned above, the ringing limit can be avoided through the use of late combustion; however, cyclic fluctuations of combustion timing become stronger after certain levels of delayed combustion phasing thereby causing power output constrains. Cyclic variability of combustion parameters in HCCI engines are produced by thermal stratifications, mixture stratifications, fluctuations in the charge air to fuel ratio, and fluctuations in diluent composition and temperature [218]. These cycle-to-cycle changes in charge temperature and composition are inevitable but they have a stronger effect on combustion timing and combustion duration when using delayed combustion timing. When the ignition event takes place in the expansion stroke, it is more sensitive to small variations in charge temperature at TDC. These changes in the ignition point cause larger effects in combustion duration and combustion completeness at the most delayed combustion phasing. Some cycles tend to ignite too late leading to high residuals temperature and incomplete combustion (which provides a pool of unburned fuel and radicals) that produce advanced ignition timing and higher energy release in other cycles. As a consequence, unstable combustion occurs, characterized by partial burn cycles with unacceptable IMEP reduction, knocking cycles, and increased HC and CO emissions [128]. Fig. 27(a) illustrates the changes in the standard deviation of CA10 related with the average CA10 and Fig. 27(b) shows the response of the standard deviation of IMEPg to changes in combustion phasing. Cyclic variability in HCCI engines depends on fuel composition. Fuels with lower octane ratings tend to exhibit lower cyclic fluctuations at the same combustion phasing [54,219]. This is because of the reduced random fluctuations in heat transfer (which are caused by the lower in-cylinder temperatures) and the increased LTHR (which produces a higher temperature-rise rate around TDC, making the hot-ignition less sensitive to changes in the compressed gas temperature).

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Fig. 26. In-cylinder pressure traces and gross IMEP for 300 consecutive cycles at an unstable operating point. Demonstrates burning cycles, partially burning cycles and misfiring cycles.

Combustion phasing also plays an important role on determining the nature of cyclic fluctuations. Three different patterns can be identified: normal, periodic, and weak/misfired. These patterns of cycle-to-cycle variations for the start of combustion (SOC), the maximum pressure (Pmax), and the IMEP are shown in Fig. 28. In order to have low cyclic variations, periodic and misfire oscillation patterns should be avoided and the magnitude of the normal variations should be kept as low as possible. These goals are possible to achieve if the start of combustion is controlled within the narrow range of 0e5 CAD aTDC. 4.5. Peak in-cylinder pressure limits The peak in-cylinder pressure limit is a design consideration that is specific to a given engine. Fig. 29 shows a pressure trace for a 4-cylinder Volkswagen TDI engine modified for HCCI operation, the peak pressure limit is 130 bar (absolute). The limit is determined by the high pressure capabilities of engine components such as the piston rings and head gasket which can be damaged by excessive in-cylinder pressures. Prolonged exposure to excessive pressure for these components can cause leaks which will allow increased blowby and poor compression, and ultimately result in unburned fuel escaping from the engine. Excessive in-cylinder pressures can be caused by a number of factors, including: 1) high intake pressures combined with high equivalence ratios which lead to high peak combustion pressures [96], 2) excessive ringing, which can cause the upper peak of the pressure pulsations (illustrated in Fig. 25) to exceed the incylinder pressure limits, 3) early combustion where the heat

release and pressure rise from the combustion event occurs in a smaller volume, thereby producing high pressures [90], and 4) a misfiring cycle followed by a burning cycle which can cause more fuel to be present in a given cycle [54]. The peak in-cylinder pressure can be kept within safe limits through strategies that avoid these conditions. Aside from simply using lower equivalence ratios, and lower intake pressures, which will also cause lower power output, the use of more prolonged heat release [79,81,83] and delayed combustion timing [86,119,120,128] are important techniques in avoiding peak in-cylinder pressure limits. More prolonged heat release will cause the pressure rise from combustion to be more gradual, thus allowing charge expansion from downwards piston motion to play a more significant role in avoiding excessively high pressures. The use of delayed combustion timing allows the pressure rise to occur at a time when the cylinder volume is greater, and takes advantage of the faster rate of charge expansion that occurs with later combustion timings (illustrated in Fig. 34). 4.6. Excessive reactivity limits There are many techniques for HCCI engines to provide the energy required for autoignition. Some methods include the use of higher intake temperatures, higher intake pressures, higher compression ratios, retaining hot EGR to heat the fueleair mixture, or using fuels with greater reactivity. As higher intake pressures are used, the requirement for higher intake temperatures is decreased in order to achieve combustion at a desired timing [11,14,55,57,63,88,89]. If sufficiently high intake pressures are used for gasoline or fuels exhibiting

Fig. 27. (a) Cycle-to-cycle variation of the phasing of hot ignition, measured as CA10, as a function of average CA10. (b) Standard deviation of IMEPg divided by (IMEPg  IMEPg,motored) as a function of CA50 [128].

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Maximum In-Cylinder P Limit y

120

Cylinder 1 Cylinder y 2 Cylinder 3 Cylinder 4

C4 4 100

C2, C3

In-Cylinder ylinder Pressure (B (Bar)

80 80 60 C1 (motored)) 40 20 0 -20

-10

0

10

20

30

40

Crank Angle Degree, 0=TDC Fig. 29. Pressure trace showing VW engine in-cylinder pressure limits.

LTHR or ITHR, another constraint is introduced because intake temperatures below ambient would be required to avoid combustion before TDC [14,64,70]. Fig. 30 illustrates how the intake temperature requirements decrease toward the excessive reactivity limits when intake pressure is increased for gasoline-fueled experiments. Aside from simply avoiding the use of high intake pressures (which would subsequently constrain the maximum power output), one strategy for avoiding the excessive reactivity limit is to use cooled EGR to dilute the in-cylinder mixture. Prior studies have used intake temperature as a technique for controlling combustion timing [81,106,107,112,220] only up to the point where intake temperatures of 60  C are required because this is a typical temperature for air exiting the compression side of a turbocharger [81]. For conditions requiring lower intake temperatures, for example when using higher intake pressures, EGR is used to induce the required combustion timing delays. Fuel type is an important consideration governing the occurrence of the excessive reactivity limit. Fuels with greater amounts of LTHR or ITHR (such as n-heptane [12,43,55,70,71,73], diesel,5 or gasoline at high intake pressures [44,79,81,87,99]) tend to encounter the excessive reactivity limit more easily because their low temperature chemistry causes the requirement of less intake heating to achieve a desired combustion timing [70]. Fuels with little or no LTHR and ITHR (typically single-stage ignition fuels like iso-octane [50,51,64], ethanol [52,53,73], or methane [221]) may never encounter the excessive reactivity limit because even at high compression ratios and high intake pressures significantly elevated intake temperatures are required to reach ignition. For these singlestage ignition fuels, other limits are likely encountered before the excessive reactivity limit becomes a constraint. 4.7. Oxygen availability limits The oxygen availability limit is applicable to conditions where EGR is used for charge dilution. The use of external (cooled) EGR is one method to avoid the excessive reactivity limit (discussed in Section 4.6) for fuels with high reactivity [100]. Cooled EGR provides a dilution effect which counteracts the increased reactivity from higher intake pressures [53,103,105,108e110,222] As a result, the use of increasing amounts of charge dilution with cooled EGR allows intake pressures to be increased without having to decrease Fig. 28. Cyclic variations patterns of SOC, Pmax, and IMEP. (a) Normal variation, (b) periodic variation and (c) variations with weak (misfire). The symbol W/M in part (c) represents weak/misfired ignitions [218].

5 It is important to note that diesel typically requires high intake temperatures to allow complete vaporization of the fuel, however from a chemical kinetics point of view, diesel’s higher LTHR causes lower intake temperatures to achieve ignition.

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Intake Temperature (K)

charge (i.e. not utilizing any of the strategies for extending the high load limits discussed in Section 5), the ringing limit will likely be encountered prior to the NOx limit. 4. Soot and particulate matter emissions are typically very low for HCCI, however these factors can become important if significant amounts of PFS are used to extend power output limits (discussed in Section 5.2).

Intake temperaturedecreases as intake pressure increases

420

400

1.0

380

19

1.2 1.4

360 1.6

340 Intake Pressure (Bar abs) P=1.0 P=1.2

320

P=1.4

P=1.6

P=1.8

300 0

P=2.0

3

1.8

Ambient intake temperature 6

9

4.9. Practical considerations for intake pressure boost in HCCI

2.0 12

15

Combustion Timing, CA50 (deg. ATDC) Fig. 30. Intake temperature requirements for maintaining constant combustion timing at different intake pressures 4 ¼ 0.45, 91-Octane commercial grade gasoline, CR ¼ 17, 1800 RPM.

intake temperatures below ambient, while still maintaining sufficiently delayed combustion timings. High power output can be achieved in HCCI, for instance a recent study demonstrating up to IMEPg ¼ 16.34 bar with gasoline [81] or 18.1 bar with E10 [99], by increasing the intake pressure and using EGR to delay ignition timing. This strategy can be used up to the oxygen availability limits, where the fueleair ratio becomes stoichiometric and further fuel injection would not result in increased power output. Once sufficiently high intake pressures are reached, the increased amount of EGR required to maintain delayed combustion timing does not allow for further increases in power output as any additional fuel would not burn completely because of a lack of available oxygen. An important consideration is that peak in-cylinder pressure limits (Section 4.5) for boosted intake are likely to be reached prior to the onset of oxygen availability limits. For an engine with a 170 bar in-cylinder pressure limit, results [1] showed that there was still potential to use EGR to increase the power output beyond IMEPg ¼ 16.34 bar, however this would result in excessively high incylinder pressures. 4.8. Efficiency and emissions limits In addition to considering the limits discussed in Sections 4.2e 4.7, the overall efficiency and emissions at a given operating point should be considered as further limits. Operating points with poor overall efficiency or excessively high emissions should be avoided and an engine control system should be designed which considers these factors as part of its control strategy. Sections 2 and 3 discussed some efficiency and emissions characteristics that should be considered, including: 1. Efficiency and power output are lower when the combustion timing is too early, or too late, and with increasing equivalence ratio the optimal efficiency and optimal power output occur with later combustion timings [90,91,149,223,224]. 2. Unburned HC and CO emissions are higher for conditions resulting in lower in-cylinder temperature, such as with low equivalence ratios or excessively delayed combustion timing [1,70,108,109,113,154,158,162,225]. 3. Nitrogen oxide emissions are higher for conditions resulting in higher in-cylinder temperature, such as with high equivalence ratios or excessively advanced combustion timing [48,110,156, 173,226,227]. For HCCI engines using a truly homogeneous

As previously mentioned in Section 4.1, naturally aspirated HCCI engines have a maximum IMEP near 5 bar, which is too low to cover all the operating requirements for automotive applications. Increasing the intake pressure is a way to extend this limit; however, the allowable boost pressure is constrained by several practical considerations. Engine boosting is achieved using superchargers or turbochargers, and the practical considerations for both strategies are discussed below. When superchargers are used to provide elevated intake pressure, the allowable pressure ratio across the compressor may be constrained by the crankshaft power needed to drive the compressor. For instance, Olsson et al. [228] have showed that in the range of desired intake pressures (Pin ¼ 1e3 bar absolute), the penalty on HCCI engine BMEP from supercharging can be higher than the actual boost pressure achieved because of parasitic losses. In turbo-charging the power to drive the compressor is provided by the expansion of exhaust gases across a radial turbine, hence causing lower parasitic losses compared with supercharging. The main drawback for this boosting strategy is the increased backpressure in the exhaust manifold, which has to be lower than the desired intake pressure to make turbo-charging feasible. According to Hyvönen et al. [199] the increase in maximum BMEP using turbo-charging is between 32 and 97% when compared with supercharging. Their results are illustrated in Fig. 31 and clearly show the higher potential of using turbochargers for enabling higher load from HCCI engines. The application of efficient turbo-charging over the entire operating range of HCCI engines requires research and development of new turbocharger designs. In HCCI, the total trapped incylinder mass at IVC timing is lower than compared with diesel engines because the charge temperature is typically higher to enable auto-ignition. As a result, the mass flow rate through the turbocharger expansion turbine after exhaust valve opening is decreased. Additionally, and more importantly, combustion temperatures are substantially lower to avoid NOx formation leading to lower exhaust gas temperatures, less available energy in the exhaust gas, and higher gas density. Less available energy in the exhaust gas leads to low turbocharger expansion turbine power output. If a turbine designed for diesel engines is used in HCCI operation, the turbine operating efficiency is very low at part load conditions, and therefore the allowable boost pressure out of the compressor is also reduced. The use of smaller turbines allows for higher pressure ratios and high turbine efficiency; however the backpressure is also increased leading to lower BMEP or low thermal efficiency. For instance, Olsson et al. [197] experimented with a 48% reduction of the turbine inlet area for multi-cylinder operation and achieved a maximum BMEP of 16 bar with a boost pressure of approximately 2 bar. However, at this load, the backpressure was close to 4 bar causing increased pumping work and a reduction of the brake thermal efficiency of up to 35%. Additionally, at this load high peak pressure rise rates and high peak pressure became limiting factors. In another study, Johansson et al. [200] tested a turbocharged multi-cylinder HCCI engine using an over-sized

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enhanced natural thermal stratifications (by enhancing heat transfer using colder coolant temperatures) vs. the use of delayed combustion timing, and concluded that delayed combustion timing is most effective at reducing peak pressure rise rates. The downsides of highly delayed combustion timing, however, are that it causes thermodynamic losses (shown in Fig. 15) [99], it can cause less stable combustion (or complete misfire) [54,230], and can require intake temperatures that are below ambient temperatures (discussed in Section 4.6).

Fig. 31. Maximum BMEP related vs. engine speed for HCCI operation with different intake charge strategies (naturally aspirated, turbocharged, and supercharged [199].

turbine which allowed for a maximum boost pressure of approximately 0.7 bar. Fig. 32 shows the results achieved for IMEPnet at different engine speeds operating with natural aspiration and turbo-charging. Even though the IMEPnet appears to be low, the results achieved sufficiently low levels of ringing intensity, NOx emissions, maximum in-cylinder pressure, soot emissions, and cyclic variability, thereby making boosted HCCI attractive for further exploration. 5. Strategies for avoiding & expanding the operating limits for HCCI and LTC Section 4 discussed the different limits that confine the maximum load in HCCI engines. One of the primary limiting phenomena constraining maximum power output is the need to prevent excessive ringing [70,90,91,151,203,205,228,229], which is discussed in Section 4.2. Avoiding excessive rates of heat release and excessive pressure rise rates is the primary means for avoiding excessive ringing [5], thus the strategies for increasing maximum power output discussed in Sections 5.1 and 5.2 focus on avoiding excessive heat release rates. 5.1. Delayed combustion timing The use of delayed combustion timing is one strategy for reducing pressure rise rates and thereby avoiding ringing [54,128,148,151]. Sjöberg et al. [128] compared the effects of

Fig. 32. Comparison of maximum IMEPnet at different engine speed for turbocharged (Turbo) HCCI and naturally aspirated (NA) HCCI [200].

5.1.1. Why delayed combustion timing avoids ringing Delayed combustion timing is useful in reducing ringing because: 1) more time is available for the natural development of thermal stratifications, which lowers the peak heat release rate, 2) faster charge expansion allows lower pressure rise rates, and 3) peak in-cylinder temperatures are lower. Each of these three contributing factors is discussed in this section. Charge temperature is a critical factor in determining the rate and timing of heat release. In Section 2.2.3, it was explained that thermal stratifications are naturally formed during the intake and compression stroke. As shown in Fig. 33, at later combustion timings there are substantially more thermal stratifications (because more time is available for the stratifications to naturally develop) [114,118,119], and these spatial temperature differences influence the localized timing and rate of heat release. Thus, increased thermal stratifications in the core region from further delayed combustion timing allow more sequential ignition with lower peak heat release rates. The lower peak heat release rates subsequently cause lower pressure rise rates to enable a lower ringing intensity. Faster charge expansion at later combustion timings causes greater rates of pressure decrease, counteracting the large pressure rise from autoignition. Fig. 34 shows the in-cylinder pressure and pressure rise rate for a naturally aspirated motored condition. The pressure rise rate from Fig. 34 shows that when combustion is delayed after top dead center, the rate of pressure decrease is higher with later crank angle degrees, with a minimum occurring near 11 CAD ATDC. Delayed combustion timing allows for lower peak in-cylinder temperatures, which also leads to a lower ringing intensity (see Eqn. (4.1)). Lower peak in-cylinder temperatures occur for two reasons: 1) the larger cylinder volume at delayed combustion timing causes lower temperature once autoignition has occurred, and 2) the more sequential ignition caused by enhanced natural thermal stratifications at delayed combustion timing allows some charge expansion to counteract the temperature rise from heat release. 5.1.2. Studies demonstrating the use of delayed combustion timing The control of intake temperature has been shown as one effective means of controlling combustion timing for different equivalence ratios, engine speeds and intake pressures [58,72,81,92,129,144,231]. Higher intake temperatures generally cause advanced combustion timings, while reduced intake temperatures delay timing. Fuel chemistry (particularly whether the fuel is a single-stage or two-stage ignition fuel) also plays an important role in determining heat release rates and the extent with which combustion timing can be delayed [86,223]. Tanaka et al. [6] showed that straight chained hydrocarbon fuels with e CH2eCH2eCH2e structures more readily show two-stage ignition behavior and shorter ignition delay times. In their work, they identified means of independently controlling the burn rate and the ignition delay that a particular fuel would exhibit in HCCI conditions. Several important studies by Sjöberg and Dec [54,58,77,81,232] have revealed specific strategies that can be employed for enabling delayed combustion timing for single-stage

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Fig. 33. Temporal sequence of single-cycle in-cylinder temperature maps. 1200 RPM, Pin ¼ 100 kPa, Tin ¼ 170  C, 4 ¼ 0.40. Figure shows that later crank-angle timings after TDC exhibit more thermal stratification. Source: Dronniou and Dec, 2012 [118].

maintaining delayed combustion timing were lower than expected temperatures out of a turbocharger’s compressor. Beyond this level of boost pressure, increasing EGR fractions were used in maintaining delayed combustion timing (as late as 19 CAD ATDC). Using this strategy, a gross IMEP of 16.34 bar was achieved for gasoline at Pin ¼ 3.25 bar abs, with a stoichiometric equivalence ratio, and 60% 40

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and two-stage ignition fuels. Specifically, for highly delayed combustion near the point of misfire, combustion stability is highly sensitive to random fluctuations in charge temperature, and higher temperature rise rates just before hot ignition have been shown effective in dampening the effects of these temperature fluctuations [54]. With increasingly delayed combustion, the temperature rise rates before hot ignition are lowered because of faster charge expansion. Fuels that exhibit LTHR or ITHR, such as PRF80, or gasoline at high intake pressures, have been demonstrated in maintaining higher temperature rise rates because of higher reactivity just before hot ignition. Numerous studies, including early research from Christensen et al., have shown that boost pressure allows higher load before the onset of excessive ringing is encountered [5,72,81,96,196,197,228,232]. In their research, Christensen et al. [96] showed fuel dependencies on maximum IMEP, showing 14 bar maximum net IMEP for natural gas and 9.7 bar net IMEP for iso-octane. Subsequent studies [81,196,233] have also shown that EGR provides an effective means of enabling higher load by smoothing heat release and pressure rise rates through dilution and enabling further delayed combustion timings. Many of the results discussed above were combined in a singlecylinder gasoline-fueled study by Dec and Yang [81]. In their study, intake temperature was decreased as boost pressures increased allowing delayed combustion timings, however beyond Pin ¼ 180 kPa abs the intake temperature levels required for

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Crank Angle Degree (CAD ATDC) Fig. 34. Motored pressure and pressure rise rate for naturally aspirated conditions.

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EGR (see Fig. 35) [81]. Even higher load, up to a gross IMEP of 18.1 bar, was demonstrated using E10 (90% gasoline, 10% ethanol) [99]. These high power operating points were demonstrated with little ringing, high overall efficiency, good combustion stability and very low levels of NOx (below US2010 NOx limits). It was found that a key factor in enabling highly delayed combustion timing was keeping bulk gas temperatures rising despite high rates of charge expansion by using ITHR reactions which for gasoline are enhanced by increased boost pressures. These ITHR reactions could be increased up to boost pressures of 180 kPa. Fuels that do not exhibit much LTHR or ITHR cannot accommodate very late combustion timings because there is no low temperature heat release to prevent the quenching effect from charge expansion. The use of delayed combustion timing is very effective for avoiding ringing limits to allow high load in HCCI engines. Aside from this, late combustion timing also allows higher exhaust temperatures if relatively complete combustion has occurred so that oxidation catalysts can be used more effectively for removing HC and CO emissions (discussed in Section 3.1). Additionally, late combustion timings significantly reduce the efficiency losses from heat transfer, which is particularly important at high load conditions (discussed in Section 2.4.2) [224]. One disadvantage, however, of late combustion timing is that it results in thermodynamic losses because there is less expansion volume available to utilize the work potential of the product gases (shown in Fig. 15). 5.2. Partial fuel stratification For fuels with sufficient 4-sensitivity, PFS enables sequential ignition to proceed from richer regions to leaner regions (although overall 4
1 remains) [77,79,81e83,86,229]. Section 2.1.5 introduced the concept of 4-sensitivity, while Section 2.2.4 introduced the influence of mixture stratifications. The reduced ringing from PFS effectively enables higher power output conditions to be achieved before ringing becomes a limiting factor. PFS for fuels displaying two-stage ignition and high 4-sensitivity has been shown as effective in preventing excessive ringing to allow higher load from HCCI engines [82,83,234]. Since PFS reduces the extent of ringing, less combustion timing delay is needed for preventing ringing, and thus PFS can also allow for improved fuel economy (roughly 2e2.5% fuel economy improvement) [111]. In an engine using PFS, PFI or early DI is used for a majority of the fuel to create a relatively homogeneous mixture. During the compression stroke, a small fraction of fuel (up to 20%) is added

Fig. 35. Extension of high load limits for gasoline fuel demonstrated using the use of delayed combustion timing. Load limits are shown as the gross IMEP at the ringing limit of 5 MW/m2 for a variety of intake pressure both with and without the use of EGR [81].

with DI to create regions that have a locally higher equivalence ratio [79]. As a result of the 4-sensitivity of certain fuels, a sequential ignition event is created as richer regions ignite earlier (note that 4
1 is always maintained). Ignition propagates toward the leaner regions to create smoother heat release as compared with conventional HCCI. This sequential ignition event allows lower peak pressure rise rates, and thus lower ringing intensity, thereby allowing more fuel to be injected to achieve higher load. Fig. 36 illustrates the operating principle in an engine using PFS. Although PFS has been shown to be effective for highly 4-sensitive fuels with two-stage ignition, there are two limits which constrain the usefulness of PFS. First, when overly delayed DI timing is used in PFS, distinctly separate ignition events can occur between the premixed charge and the DI charge. Under these conditions, the maximum pressure rise rate of the premixed charge ignition will constrain the operating limits. Second, when too high a DI fuel fraction is used in PFS, the heat release rate of the DI portion can be too high [80] and/or high local temperatures may lead to excessive NOx formation. The ability of using PFS with gasoline fuel has been a recent subject of study, and at naturally aspirated conditions PFS is ineffective for reducing ringing since gasoline is not highly 4-sensitive at these operating conditions [79]. Interestingly, as the intake pressure is increased in gasoline HCCI engines, gasoline becomes highly 4-sensitive since more intermediate temperature heat release (ITHR) reactions occur [72,79,81,87,176]. Thus, at high load conditions which utilize high boost pressure, PFS can be used with gasoline fuel to reduce ringing and allow higher efficiency. The potential for using PFS to lower heat release rates has also been studied with other fuels. PFS with ethanol shows no performance improvements because ethanol is not 4-sensitive, however other fuels like hydrobate and iso-pentanol demonstrate significant improvements with PFS [78,82,84,176]. However, the 4-sensitivity behavior of these fuels with changing intake pressures is different from gasoline, therefore the engine control strategy when using PFS will have to adapt to the fuel properties. 5.3. LTC with double fuel injections and distinct heat release stages In LTC one of the most important goals is to separate the end of the injection from the start of combustion to avoid a high equivalence ratio stratification. For diesel-like fuels the main disadvantage is that the early combustion phasing because of the reduced ignition delay leads to lowered load [139,142]. On the other hand, gasoline-like fuels exhibit higher auto-ignition resistance and extend the ignition delay period. More prolonged mixing times lead to more uniform charge distribution and leaner equivalence ratios [235]; however, misfire operation becomes an important constraint with high octane rating fuels in LTC [236]. The use of fuels with auto-ignition properties between diesel and gasoline-like fuels has allowed simultaneous reductions of NOx and soot emissions; however, high pressure rise rates still appear as an important constraint at high load [237]. A novel strategy to achieve simultaneous reductions in NOx and soot emissions, as well as in pressure rise rates at LTC is based on using low octane fuels along with double fuel injection [238e240]. Around 40% of the fuel is injected early in the compression stroke and the rest of the fuel is injected close to TDC and after the first stage of heat release to avoid excessive diffusive combustion.6 Fig. 37 shows the heat release rates and injection profiles for

6 For this operating strategy timing can be a critical consideration, thus it is important to account for the effects of injector response time over the desired range of engine speed operation.

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Fig. 36. Operating principle behind an engine using PFS Note: Premixing can be accomplished either with PFI as indicated in the figure, or with early DI.

PRF84 and gasoline at high load. The maximum pressure rise rate is 4.9 bar/degree, and the noise level is 96 db [238]. NOx emissions are reduced to 1 g/kWh, however this is still too high for most transportation and stationary applications. 5.4. Spark-assisted HCCI/spark-assisted compression ignition Through an extensive review of literature, the authors of this article have found that the terms spark-assisted HCCI (SA-HCCI) and spark-assisted compression ignition (SACI) are used to describe very similar engine combustion strategies. In this section, we refer to this branch of LTC engine technology as SA-HCCI simply for continuity with the rest of this article. 5.4.1. How SA-HCCI works, and fundamental phenomena affecting SA-HCCI The combustion process in SA-HCCI can be divided into four distinct phases [241], which are depicted in Fig. 38 and listed below: Phase 1 e Spark discharge: In studies by Reuss et al. [241], this process is defined based on measurements of spark current. Phase 1 begins at the desired spark timing and ends at the timing when spark current returns to zero. Phase 2 e Early kernel growth (EKG): EKG is the period in which a small reaction kernel exists in the vicinity of the spark discharge. This phase typically has weak chemiluminescence and a slow flame front growth area. Given that heat release is weak in this phase, it is difficult to detect with pressure diagnostics and must therefore be studied using optical techniques. EKG begins when spark current returns to zero and ends when the flame area grows most rapidly, defined as the maximum of d2Aflm/dCA2, where Aflm is the flame area measured through optical observations and CA is the crank angle timing [241]. Phase 3 e Flame propagation: Reuss et al. [241] propose that the term ‘flame’ should not be taken literally, as the actual source of heat release in this phase may be a combination of both compression ignition and true flame propagation, however it is difficult to distinguish between the two from measurements. In other studies [242,243], this phase has been labeled as the initial slow heat release (ISHR), and numerous studies have proven through premixed laminar flame simulations that the highly pre-heated and ultra-diluted in-cylinder mixture existing in SAHCCI is capable of supporting laminar reaction fronts [244e 246]. Phase 3 begins at the maximum of d2Aflm/dCA2 (i.e.

Fig. 37. Heat release rate and injection profiles at LTC with multiple injections for 12 bar BMEP [238].

when flame area grows most rapidly) and ends when the heat release rate increases most rapidly, defined as the maximum of d2HRR/dCA2 [241]. Phase 4 e Compression ignition: This phase most resembles HCCI combustion or SI engine knocking, as the unburned end gases are ignited through heating and compression from the advancing flame front. However, it is important to note that this phase is distinct from SI engine knocking because the heat release occurs in a more controlled way [241], likely allowing the end-gas autoignition to be tolerable for long term engine operation. This phase begins at the maximum of d2HRR/dCA2 (i.e. when heat release rate increases most rapidly) and includes all remaining heat release. A comprehensive modeling methodology for SA-HCCI was introduced in a prior study by Dahms et al. [247]. The proposed methodology simultaneously captures the many interacting phenomena of the mixed-mode combustion process, and the authors of the Dahms et al. [247] study invite collection of further experimental data over a wider range of SA-HCCI engine operating points to fully validate the model. Through detailed studies, many researchers have confirmed the four fundamental processes listed above, which can be summarized as an SI phase followed by a more rapid HCCI phase [241e 243,245,248e252]. An important observation from several studies is that the flame propagation speed of the flame front (in phase 3) increases as the chemical reactions in the end gas proceed; or in other words, the flame propagation speed depends on the degree of pre-reaction upstream of the flame front [244,247,253e255]. The increased flame speed occurs because the temperature gradient across the flame front decreases due to the end gas temperature rising as early chemical reactions in the end gas proceed [244]. A study by Martz et al. [244] has introduced another metric for quantifying the flame front’s transition from deflagrative to chemically controlled combustion (i.e. analogous to the transition between phase 3 and phase 4 listed above), thereby allowing quantification of the degree of heat release from flame propagation vs. from autoignition. Studies have shown that the relative contribution to heat release of flame propagation vs. autoignition can be continuously varied by controlling residual gas content (for instance, using variations in valve timing) [242,256e258]. However, changes in the residual gas content also cause changes in the flame propagation speed in phase 3. Prior studies show that the

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Fig. 38. Sample heat release rate from pressure diagnostics (solid line) and flame area growth from optical diagnostics (solid line with square data points) for a SA-HCCI engine [241].

burning velocity of a mixture diluted with residual gases is lower than that of a mixture diluted with air [242,252]. The change in laminar flame speed as the oxygen concentration is changed7 is due to changes in the burned gas temperature [259] and a chemical effect between a competing chain branching (O2 þ H 4 OH þ O) and chain termination reaction (O2 þ H þ M 4 HO2 þ M) occurring at the high temperature region of the reaction front [252]. A prior study [254] has mapped the operating regions where flame propagation reaction fronts can be maintained and this study identifies both a burned gas and pressure dependence in the viable range of operating conditions where reaction fronts can be maintained. A convenient multi-mode combustion diagram, shown in Fig. 39, illustrates the operating limits for SA-HCCI. The diagram was presented in a study by Lavoie et al. [257]. The lines encompassing the SACI region on the diagram illustrate the operating limits constraining this operating regime, including limits for ignition, flame propagation, excessive knocking or ringing, bulk quenching, and excessive NOx emissions, while the data points illustrate experimental limit points for SI (circles) and HCCI (triangles) [257]. 5.4.2. Why SA-HCCI increases load limits and constraints on high power SA-HCCI Section 5.4.1 summarized the fundamental processes occurring in SA-HCCI, and this section will identify how these fundamental processes allow higher load for SA-HCCI engines compared with HCCI engines at the same levels of intake pressure. Simply stated, the slower heat release from the flame propagation process combined with the lower overall rapid heat release from autoignition is the underlying reason why SA-HCCI can achieve higher load levels than HCCI for the same intake pressure [242,250,257,260]. Referring back to the ringing intensity correlation in Eq. (4.1), it can be seen that lower peak pressure rise rates are the principal means to avoid excessive ringing. As SA-HCCI allows more temporally distributed heat release and lower overall peak heat release rate, there is also a lower peak pressure rise rate (PPRR). The lower PPRR

7 It is important to recall that with air dilution high O2 concentration is maintained, while with EGR dilution the O2 concentration is reduced.

and the resulting lower ringing intensity allow more fuel to be added resulting in a higher load before ringing limits become constraining. A prior study [248] has shown that as higher power output conditions are achieved, a larger portion of the overall heat release must come from the flame propagation phase to avoid high PPRR. However, given that the ringing limits are being avoided, an important constraint for high power output SA-HCCI becomes the levels of NOx emissions [248,261]. 5.4.3. Power output levels demonstrated in SA-HCCI Several studies have demonstrated improved power output from naturally aspirated SA-HCCI in comparison with naturally aspirated HCCI. Urushihara et al. [262] demonstrated power output of 6.5 bar IMEP for naturally aspirated SA-HCCI. Szybist et al. [248] demonstrated 7.5 bar NMEP with SA-HCCI, and demonstrated the ability to maintain high load (between 7.0 and 7.5 bar NMEP) for a range of engine speeds from 1000 to 3000 RPM. Manofsky et al. [261] demonstrated that the load limits could be extended from roughly 3.7 bar NMEP with naturally aspirated HCCI up to roughly 7.5 bar NMEP with naturally aspirated SA-HCCI while maintaining good efficiency and low emissions. To the authors’ knowledge, the highest load for naturally aspirated SA-HCCI was demonstrated in a study by Yun et al. [258], where 10 bar IMEPg was achieved while maintaining good efficiency and satisfying all constraints in terms of noise (including ringing), stability (i.e. cyclic variability and misfire) and emissions. This same study explored two separate SA-HCCI operating strategies, one using negative valve overlap (NVO) and another using positive valve overlap (PVO), with the 10 bar IMEPg result being achieved for the NVO case. For the NVO case it was confirmed that as load increases combustion timing must be delayed to avoid high ringing (consistent with the discussion in Section 5.1) and for the NVO strategy this can be achieved with changes in the amount of NVO or from delayed spark timing. During the NVO portion of the study it was found that regardless of engine load, 30e40% of the fuel was burned during flame propagation and the rest during compression ignition. The study concluded that there was still room for improvement in terms of achieving even higher load, beyond 10 bar IMEPg, using the NVO strategy. The PVO strategy was shown to offset some of the drawbacks of the NVO strategy (such as

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Fig. 39. Multi-mode combustion diagram for SI, SA-HCCI/SACI, and HCCI regions in terms of unburned gas (x-axis) and burned gas (y-axis) temperature for iso-octane in air [257].

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at low load regions [192,250,251,257,262], and prior studies have developed detailed techniques for quantifying the source of cyclic variability for SA-HCCI and show that much of the variability in autoignition timing occurs during the early kernel growth period [241]. Low load limits for SA-HCCI have been extended while maintaining acceptable levels of operating efficiency [248,257,264] and particular strategies have been identified that allow successful low load operation of SA-HCCI, namely the multiple injection multiple ignition (MIMI) strategy [146,266], and the enhanced swirl strategy [242]. SA-HCCI experiments have shown lower levels of CO and HC emissions compared with normal SI engine operation. Additionally, SA-HCCI with residual gas dilution can still maintain stoichiometric operating compatibility with three-way catalysts while potentially achieving the higher levels of exhaust temperature required for oxidizing CO and HC through aftertreatment [248]. 5.5. HCCI and LTC for power generation in advanced powertrains

with increased pumping losses at higher load), while altering the amount of PVO could be used to vary the amount of trapped residual gas. The PVO strategy achieved lower ringing compared with the NVO strategy, and was able to achieve earlier combustion phasing while maintaining similar levels of combustion stability for all loads because the PVO strategy allowed a higher portion of the fuel to be burned through flame propagation. The PVO strategy allowed more favorable in-cylinder charge conditions to avoid ringing and allow higher efficiency through lower pumping losses; however the load levels achieved with PVO were not reported. In the opinion of the authors of this review article, there is little published research on SA-HCCI under boosted conditions. The higher load potential of boosted SA-HCCI may be an interesting area for future research; however there are many complex interacting effects which must be studied at high pressure conditions that influence the four fundamental processes in SA-HCCI identified in Section 5.4.1. 5.4.4. Other benefits of SA-HCCI Section 5.5 briefly discusses the use of dual-mode engines that use HCCI at intermediate load conditions and SI at higher load conditions. SA-HCCI has been shown to be an effective method for bridging the gap between HCCI and SI as it allows successful operation above the HCCI ringing limit and below the SI misfire limit [192,261,263]. Aside from the higher load that can be achieved with SA-HCCI compared to HCCI at equivalent levels of engine intake pressure, and the ability to bridge between HCCI and SI, SAHCCI presents many other benefits that can overcome the challenges related to pure HCCI. These are: 1) improved combustion timing control ability, 2) extension of the low load limits and improved cyclic stability at the low load points, and 3) potential improvements in emissions. Each of these is discussed in the three paragraphs below. This review article has briefly discussed other strategies that can allow effective cycle-resolved control of combustion timing in HCCI (see Section 2.3.2) and the use of SA-HCCI can also enable cycleresolved control [192,241,243,257,264]. Roughly speaking, the autoignition timing is proportional to spark timing, however the detailed relationships between spark timing and autoignition timing have been studied [242]. Additionally, the ability to control autoignition timing with SA-HCCI during mode transitions from HCCI to SI has also been demonstrated [265]. Cycle resolved control of autoignition timing has also been shown to be possible even down to engine speeds as low as 1000 RPM [247,248]. HCCI in its purest form has difficulty in operating at low load operating points due to unacceptable levels of cyclic variability or due to misfire. SA-HCCI has demonstrated improved cyclic stability

In a personal transportation context, two emerging technologies can allow wide market penetration of HCCI engines. The first is dual-mode engines which operate as HCCI engines for lower and medium load regions, and then as traditional Diesel or SI engines for high load regions [245,267e269]. Simulations of a SI-HCCI engine in the EPA urban and highway drive cycles have shown that this dual-mode approach can allow a 17% fuel economy improvement over a conventional SI engine [270]. The second approach is using HCCI engines within a hybrid-electric powertrain [270e272]. In this second approach, the hybrid powertrain can accommodate many of the transient and high load requirements upon an engine, making HCCI engines a good fit for this approach. Recent research has shown that combining the above strategies of using SI-HCCI engines within a mild hybrid powertrain can enable a 35% fuel efficiency improvement over a traditional SI engine [270], and a 6e 12% fuel efficiency improvement over a mild hybrid powertrain using a high efficiency Atkinson SI cycle [271]. In a locomotive application, diesel-electric powertrains are already being used [273] to combine the high efficiency aspect of diesel engines with the high torque (at low speeds) aspect of electric motors. Although diesel engines provide high efficiency similar to HCCI engines, the HCCI engine applied to locomotive applications can significantly reduce the need for expensive aftertreatment technologies [138,274e279]. An HCCI engine in a locomotive application can use simple oxidation catalysts to remove harmful unburned HCs and CO, while a Diesel engine also requires expensive aftertreatment technologies to remove nitrogen oxides (such as selective catalytic reduction, or lean NOx traps) [280,281]. 6. Conclusions 6.1. Summary of important concepts A detailed review of the fundamental phenomena in HCCI engines and their interactions was presented in this article, particularly in relation to the high load operating limits. First, a review of hydrocarbon fuel breakdown was presented, including the chemical pathways for low and intermediate temperature chemistry and hot ignition. The characteristics of different fuels were discussed, with a focus on single- and two-stage ignition, the influence of molecular structure on fuel vaporization, and the property 4-sensitivity. Next, the importance of in-cylinder charge conditions was discussed. This included a review of intake charge conditions, the different types of EGR and its effects, and the importance of thermal and mixture stratifications.

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The factors determining combustion timing were then reviewed, followed by a brief review of common strategies for controlling combustion timing. Combustion timing was identified as one of the key variables influencing HCCI operating and emissions characteristics, and thus an in-depth review of the impact of combustion timing on other variables was presented. The role of heat transfer and the factors that determine the amount of heat loss were then discussed. This included a review of how different engine operating parameters influence the amount of heat loss from the in-cylinder charge, and the relative importance of heat loss compared with other efficiency losses. A brief review of HCCI emissions characteristics was presented, with a discussion of unburned hydrocarbons, carbon monoxide, nitrogen oxides, soot and particulate matter. The emissions characteristics of HCCI engines are identified as one of its principle benefits. HCCI allows high efficiency to minimize emissions of the global pollutant CO2, while simultaneously allowing low emissions of urban pollutants. HCCI generally has very low emissions of nitrogen oxides, soot and particulate matter thereby avoiding the need for expensive aftertreatment technologies like particulate filters, SCR or lean NOx traps. A review of engine operating strategies and less expensive aftertreatment technologies was presented to minimize emissions of unburned hydrocarbons and carbon monoxide. With regards to high load operation of HCCI, the limits constraining maximum load were discussed in detail. One of the principal constraints is the ringing limit, where excessive heat release rates cause high amplitude pressure oscillations, analogous to knocking in spark-ignited engines. Misfire and cyclic variability are also important constraints that dictate the extent to which delayed combustion timing can be used for avoiding ringing. Under high load conditions, where high intake pressures are used, peak in-cylinder pressure limits also become a constraint. For certain fuels, particularly those with low or intermediate temperature heat release, excessive reactivity limits also become a constraint as intake temperatures below ambient conditions would be required to allow sufficiently delayed combustion timing. To circumvent these excessive reactivity limits, EGR is often used, however EGR can only be used up until the oxygen availability limits where further increases in fuel injection are ineffective because of the unavailability of oxygen. Finally, for any operating point, the efficiency and emissions criteria must be considered. For instance, although HCCI generally produces low NOx, at certain high load conditions NOx begins to become a concern. Furthermore, with excessive combustion timing delays, unburned hydrocarbons or carbon monoxide can also become a concern. Finally, some promising strategies for extending the high load limits of HCCI were discussed. The first strategy, the use of delayed combustion timing, is particularly promising because it effectively avoids ringing, maintains low NOx, and for fuels with low or intermediate temperature heat release, substantial combustion timing delays can be used while maintaining stable combustion. While operating at high equivalence ratios, where high in-cylinder temperatures are used, low levels of unburned hydrocarbons and CO are produced, while the use of delayed combustion timing also allows high exhaust temperatures such that oxidation catalysts can be used more effectively. Prior research has demonstrated the ability to achieve high power output, up to IMEPg ¼ 16.34 bar with gasoline and IMEPg ¼ 18.1 bar with E10, using delayed combustion timing. The second strategy for extending high load limits is the use of partial fuel stratification. This strategy is only effective for fuels that exhibit 4-sensitivity in their ignition timing. For these fuels, the majority of fuel can be premixed using PFI or early DI, while DI is used to create a small region of higher equivalence ratio. The

regions with higher equivalence ratios ignite earlier and produce a sequential ignition which propagates outwards to the lower equivalence ratio regions, thereby producing a smoother heat release rate compared with pure HCCI. This strategy allows for low ringing without the need for highly delayed combustion timing, thereby allowing higher power output or higher thermal efficiency than the delayed combustion timing strategy. The third strategy for enabling higher load is the use of double fuel injections, however this strategy is best for low octane fuels. Roughly 40% of the fuel is injected early to allow a relatively dilute and homogeneous charge which ignites to create the first stage of heat release. The second stage of heat release is created by directly injecting the fuel during or after the first stage of heat release. This strategy has shown promise for enabling lower pressure rise rates and noise levels, however NOx emissions may be a challenge. The fourth strategy allowing higher load is with spark-assisted HCCI. SA-HCCI takes advantage of the slower heat release of flame propagation to avoid excessive pressure-rise rates that lead to ringing. SA-HCCI allows substantially improved load over HCCI without requiring intake pressure boosting, and provides many other benefits such as improved low load operation and improved cycle-resolved control of combustion timing. Finally, the use of HCCI or LTC engines in advanced hybridelectric powertrains was discussed. Combining these advanced engines with electric powertrains relaxes the high load requirements of the engine and allows the engine to remain in its optimal efficiency region. Prior studies have shown that HCCI combined with electric powertrains over the EPA urban and highway drive cycles enables significant fuel savings compared with SI engines, and even SI-electric hybrids. 6.2. Promising research directions This article has systematically identified the fundamental phenomena influencing the operation of LTC engines (specifically HCCI), emissions considerations, limiting regimes and strategies for operating at high load. Despite the substantial improvements to LTC engines in recent years, additional research challenges still remain. In the United States, the National Science Foundation (NSF) and Department of Energy (DOE) have formed a partnership for a program targeting a 25e40% improvement in fuel economy in light duty vehicles and the attainment of 55% brake thermal efficiency in heavy-duty engine systems. Developing engine technologies to achieve these targets will require substantial research, even beyond the prior research achievements covered in this article. Readers are directed to the NSF/DOE program solicitation [282] for more information on “big picture” challenges that may be the focus of research in the next few years. The remainder of this section focuses on areas that the authors assess as gaps in the literature, or further fundamental understanding that is required to improve LTC engines. Further expansion of the high load limits is required for LTC. The authors believe that more research is required to translate the promising high load operating strategies discussed in Section 5 into engine operating strategies that cover the full engine load and speed requirements while simultaneously addressing the control complexities that occur on a multi-cylinder engine. Effective cycleresolved control for each engine cylinder will be critical to maintain engine stability while operating near the various operating limits discussed in Section 4. Facilitating further high load operation will require better understanding of the fundamental phenomena occurring at high load conditions, particularly when using engine boosting. For instance, better understanding is required of the chemical kinetic processes leading to intermediate temperature heat release in certain fuels and better explanations are required to understand the reported

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qualitative observations that fuels with LTHR or ITHR tend to exhibit more 4-sensitivity. For spark-assisted HCCI, a clear gap in the literature exists in understanding the complex interacting phenomena for boosted SA-HCCI operation. As ringing is one of the principal constraints that dictate the high load limits, better fundamental understanding is required of this phenomena. Research efforts to improve the ringing intensity correlation are underway, and better quantitative understanding of how ringing impacts heat loss is required. Unburned hydrocarbon and carbon monoxide emissions from certain LTC operating regimes are a problem that requires further research. Strategies for lowering HC and CO emissions from HCCI are required, or aftertreatment technologies must be developed to address this problem while considering the lower exhaust gas temperatures from LTC and HCCI engines. Beyond simply expanding the load limits, better systems integration of LTC engines in power generation or vehicle powertrain systems is required while considering the operating range and requirements of all interconnected systems. To cite a specific example, practical deployment of high load LTC or HCCI engines will require turbochargers that are designed for the unique operating conditions of these engines. Acknowledgments The authors wish to acknowledge many colleagues in the engine research community for their thoughts and insights regarding the fundamental phenomena discussed in this article. The participants at the bi-annual U.S. DOE Advanced Engine Consortium meetings have provided valuable understanding about low temperature combustion engine technologies. For the fruitful discussions that helped improve the authors’ understanding on the topics covered in this article, the authors particularly acknowledge Dr. John Dec at Sandia National Laboratory, Professor Darko Kozarac at the University of Zagreb, Professor Mani Sarathy at King Abdullah University of Science and Technology (KAUST), and Professors Robert Dibble and J-Y Chen at the University of California at Berkeley. Funding to support some of the research topics discussed in this article was provided through the U.S. Department of Energy (through the HCCI/Advanced Engine Consortium), the Natural Sciences and Engineering Research Council of Canada, laboratory directed research and development funds at Lawrence Berkeley National Laboratory, and Programa de Sostenibilidad de Grupos de Investigación 2013e2014 Vicerrectoría de Investigación (Universidad de Antioquia e Colombia). Finally, the authors thank our readers for choosing this article. We welcome any feedback or questions of the topics discussed in this article, or discussion of emerging engine or vehicle powertrain technologies. In particular, we welcome discussions on potential research collaborations to develop engine and powertrain technologies for vehicles of the future. References [1] Flowers DL, Aceves SM, Martinez-Frias J, Dibble RW. Prediction of carbon monoxide and hydrocarbon emissions in iso-octane HCCI engine combustion using multizone simulations. Proceedings of the Combustion Institute 2002;29(1):687e94. http://dx.doi.org/10.1016/S1540-7489(02)80088-8. [2] Jun D, Ishii K, Iida N. Combustion analysis of natural gas in a four stroke HCCI engine using experiment and elementary reactions calculation. SAE 200301-1089; 2003. http://dx.doi.org/10.4271/2003-01-1089. [3] Battin-Leclerc F. Detailed chemical kinetic models for the low-temperature combustion of hydrocarbons with application to gasoline and diesel fuel surrogates. Progress in Energy and Combustion Science 2008;34(4):440e98. http://dx.doi.org/10.1016/j.pecs.2007.10.002. [4] Zádor J, Taatjes CA, Fernandes RX. Kinetics of elementary reactions in lowtemperature autoignition chemistry. Progress in Energy and Combustion Science 2011;37(4):371e421. http://dx.doi.org/10.1016/j.pecs.2010.06.006.

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