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- Herausgeber: Wiley-VCH
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- Sprache: Englisch
Written by experts in combustion technology, this is a unique and refreshing perspective on the current biofuel discussion, presenting the latest research in this important field.
The emphasis throughout this reference is on applications, industrial perspectives and economics, focusing on new classes of biofuels such as butanols, levulinates, benzenoids and others. Clearly structured, each chapter presents a new class of biofuel and discusses such topics as production pathways, fuel properties and its impact on engines.
The result is a fascinating, user-oriented overview of new classes of biofuels beyond bioethanol.
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Seitenzahl: 392
Veröffentlichungsjahr: 2016
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Table of Contents
Cover
Title Page
Copyright
List of Contributors
Preface
Acknowledgments
Chapter 1: Fuels and Combustion
1.1 Introduction
1.2 The Options
1.3 Spark Ignition
1.4 Compression Ignition
1.5 Highly Diluted Autoignition, HCCI
1.6 Other Combustion Concepts
1.7 Summary of Combustion Processes
References
Chapter 2: Fuel Class Higher Alcohols
2.1 Introduction and Fuel Properties
2.2 Performance in Spark-Ignition Engines
2.3 Performance in Compression-Ignition Engines
2.4 Production Pathways
2.5 Outlook
2.6 Conclusions
References
Chapter 3: Fuel Class Valerates
3.1 Introduction and Fuel Properties
3.2 Performance in Spark-Ignition Engines
3.3 Performance in Compression-Ignition Engines
3.4 Production Pathways
3.5 Outlook
3.6 Conclusions
Acknowledgments
References
Chapter 4: Butyl Ethers and Levulinates
4.1 Introduction and Fuel Properties
4.2 Performance in Compression-Ignition Engines
4.3 Production Pathways
4.4 Outlook
4.5 Conclusions
References
Chapter 5: A Comprehensive Review of 2,5-Dimethylfuran as a Biofuel Candidate
5.1 Introduction to DMF
5.2 Production Pathways
5.3 Performance in Spark-Ignition Engines
5.4 Performance in Compression-Ignition Engines
5.5 Outlook
5.6 Conclusions
References
Chapter 6: Furanoids
6.1 Introduction and Fuel Properties
6.2 Performance in Spark-Ignition Engines
6.3 Performance in Compression-Ignition Engines
6.4 Production Pathways
6.5 Outlook
6.6 Conclusions
References
Chapter 7: Benzenoids
7.1 Introduction
7.2 Overview of Neat Fuel properties
7.3 Performance in Compression-Ignition Engines
7.4 Performance in Spark-Ignition Engines
7.5 Production Pathways
7.6 Outlook and Conclusions
References
Chapter 8: Biomass Pyrolysis Oils
8.1 Introduction and Fuel Properties
8.2 Performance Spark-Ignition Engines
8.3 Performance in Compression-Ignition Engines
8.4 Production Pathways from Pyrolysis Oil
8.5 Outlook
8.6 Conclusions
References
Index
End User License Agreement
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Guide
Cover
Table of Contents
Preface
Begin Reading
List of Illustrations
Chapter 1: Fuels and Combustion
Figure 1.1 The three fundamental combustion concepts, SI, CI, and HCCI.
Figure 1.2 Normal flame propagation [1].
Figure 1.3 The required compression ratio for knock to be a function of RON and MON.
Figure 1.4 Required compression ratio as a function of octane number of RON [1].
Figure 1.5 Conceptual model of diesel combustion by John Dec.
Figure 1.6 HCCI combustion at 7, 9, and 11 CAD after top dead centre (ATDC) in a single cycle.
Figure 1.7 Combinations of fuel octane number inlet temperature and compression ratio that achieve combustion at correct combustion phasing.
Figure 1.8 Combinations of inlet temperature and compression ratio that achieve combustion at correct combustion phasing with mixtures of diesel fuel and gasoline.
Figure 1.9 Spark-assisted compression ignition in between SI and HCCI.
Figure 1.10 Initial flame propagation and subsequent autoignition in SACI in a single cycle.
Figure 1.11 Partially premixed combustion, PPC.
Figure 1.12 Gradual change of unburned fuel (hydrocarbon, HC) and NO
x
with later start of injection (SOI) and indication of combustion mode.
Figure 1.13 Reactivity controlled compression ignition, RCCI.
Figure 1.14 Dual-fuel combustion.
Figure 1.15 Dual fuel with compressed natural gas in the inlet port and diesel injected in the cylinder.
Figure 1.16 Prechamber SI combustion.
Figure 1.17 Typical prechamber for large natural gas engine.
Figure 1.18 Diesel pilot ignition.
Figure 1.19 The three major combustion modes and five intermediate processes. Blue indicates single fuel and red, dual fuels.
Chapter 2: Fuel Class Higher Alcohols
Figure 2.1 (a) Laminar flame speeds for various alcohols and hydrocarbons in air at 1 atm, 353 K, and equivalence ratio of 1.1. (b) Correlation of research octane number (RON) with ignition delay time for stoichiometric fuel/air mixtures near 835 K and 20 atm. The line represents PRF mixtures of
n
-heptane and
iso
-octane.
Figure 2.2 Contribution of aldehyde emissions to organic matter hydrocarbon equivalent (OMHCE) emissions.
Figure 2.3 Impact of engine load on emissions of a CI engine fueled with pure
n
-butanol.
Figure 2.4 Heat release rates for various
n
-pentanol blends and engine load conditions. DP10 represents a blend of 10 vol%
n
-pentanol with diesel fuel.
Figure 2.5 Butanol production from wheat straw using
C. beijerinckii
P260.
Figure 2.6 (a) Scheme for synthesis of
n
-octanol for lignocellulosic feedstock. (b) Multistep synthesis of
n
-octanol from biogenic platform chemicals furfural and acetone.
Chapter 3: Fuel Class Valerates
Figure 3.1 Roadmap for the conversion of lignocellulosic biomass (rectangles) to chemicals (ovals) and fuels (pentagons), via the intermediate formation of C5 and C6 sugars (hexagons), suggested by Azadi
et al
.
Figure 3.2 Four-step process from lignocellulose to valerate proposed by Lange
et al.
[9].
Figure 3.3 Optimum ignition conditions for valerates as fuel compared with PRF.
Figure 3.4 Experimental values of laminar burning velocities of ethyl and methyl pentanoate/air mixtures at
T
= 423 K (a) as a function of the equivalence ratio at 1 bar (filled symbols) and 3 bar (open symbols). Comparison with
iso
-octane. (b) Laminar burning velocity for EV as a function of the pressure at
ϕ
= 1 (filled symbols: data; open symbols and dashed line: computations).
Figure 3.5 Experimental (symbols) and computed (lines) mole fraction profiles obtained from the oxidation of ethyl pentanoate in a JSR at
ϕ
= 1,
p
= 10 atm,
τ
= 0.7 s (EPE, ethyl pentanoate).
Figure 3.6 Single-cylinder SI engine bench.
Figure 3.7 Efficiency and CO
2
emissions as a function of IMEP for EV, MV, PRF as fuel, and 20% EV and MV blends. The open symbol is for 1500 rpm and the filled one for 2500 rpm.
Figure 3.8 NO and UHC emissions for pure fuels as a function of IMEP.
Figure 3.9 Example of cylinder pressure and HRR evolutions for MV, EV, and PRF95 (intake pressure = 0.6 bar, IMEP = 4.5 bar).
Figure 3.10 Estimated ignition delay for MV and EV at different temperatures.
Figure 3.11 Specific consumption as a function of the low heating value for the three different fuels for both conditions.
Figure 3.12 Unburned hydrocarbons (UHC) emissions for all fuels.
Figure 3.13 Ethylene, methane, and benzene emissions for all fuels and under both engine conditions.
Figure 3.14 Compression-ignition single-cylinder bench.
Figure 3.15 Indicated efficiency as a function of IMEP for diesel and blends with 20% of BV and PenV.
Figure 3.16 Ignition delay determined by CA5-SI
1
for BV and PenV in comparison with diesel.
Figure 3.17 CO, unburnt HC, and NO emissions as a function of IMEP.
Figure 3.18 FSN values as a function of IMEP for both blends.
Chapter 4: Butyl Ethers and Levulinates
Figure 4.1 Modeled ignition delay times of di-
n
-butyl ether (DBE),
n
-octane, and
n
-butanol.
Figure 4.2 Spray investigations of DNBE and blends with diesel as defined by the Coordinating European Council (CEC), penetration length, and probability of ignition.
Figure 4.3 Emissions and efficiency for different blends of 2-MTHF and DNBE, mid-part-load operation.
Figure 4.4 Emissions and efficiency for different blends of 2-MTHF and DNBE, higher-part-load operation.
Figure 4.5 Effect of EL/diesel blending on cetane number [16].
Figure 4.6 Part-load operation engine results for diesel blends including EL.
Figure 4.7 Part-load operation engine results for diesel blends with biodiesel and BL [19].
Figure 4.8 Emission results comparison for diesel, 1-decanol, and BLT [13].
Figure 4.9 Relative soot luminosity of diesel and butyl levulinate (BTL) with and without port deactivation (PD), load point 1500 min
−1
, 6.8 bar IMEP.
Figure 4.10 From lignocellulose to levulinic acid.
Figure 4.11 From levulinic acid to levulinate esters.
Chapter 5: A Comprehensive Review of 2,5-Dimethylfuran as a Biofuel Candidate
Figure 5.1 Rationale for converting carbohydrates to DMF.
Figure 5.2 Schematic diagram of the process for conversion of fructose to DMF (In aqueous phase, there are 1.5 g of 0.25 M HCl, with 30 wt% salt-free fructose. The volume ratio of the aqueous phase and extracting phase is 3 : 2. In R2, CuRu catalyst is used with the ratio of Cu and Ru fixed at 3 : 2).
Figure 5.3 High-speed Schlieren images of DMF, ethanol, and gasoline laminar flame propagation.
Figure 5.4 Mass fraction burn profiles for DMF, ethanol, and gasoline when using fuel-specific MBT/KLSA, (a) 3.5 and (b) 8.5 bar IMEP.
Figure 5.5 In-cylinder pressure profiles for DMF, ethanol, and gasoline when using fuel-specific MBT/KLSA, (a) 3.5 and (b) 8.5 bar IMEP.
Figure 5.6 Formaldehyde and acetaldehyde engine-out emission concentrations using ethanol, DMF, and gasoline.
Figure 5.7 Ignition delay at different EGR rates of the blends.
Figure 5.8 Cylinder pressure and heat release rate for diesel, D20, D40, and G40 at a 30% EGR rate (engine speed = 1400 rpm, equivalent of 60 mg diesel single injection and the timing of 6° BTDC).
Figure 5.9 Soot and NO
x
emissions for diesel, D20, D40, and G40 at various EGR rates (engine speed = 1400 rpm, equivalent of 60 mg diesel single injection and the timing of 6° BTDC).
Chapter 6: Furanoids
Figure 6.1 Pressure histories of biofuels (CR = 14,
p
0
= 0.076 MPa,
p
1
= 3.0–3.1 MPa,
T
1
= 840–850 K).
Figure 6.2 Pressure histories of PRF90 and PRF90 + 20% v/v.
Figure 6.3 Correlation between blending fraction and improvement of octane number (CR = 14,
η
v
= 76%).
Figure 6.4 Results of the load variation at 2000 rpm with CR = 12 for all fuels and CR = 8.5 for RON 95.
Figure 6.5 (a) MBT/KLSA spark timings at various engine loads for 2-MF, DMF, ethanol, and gasoline and (b) ratio of heat of vaporization (HV) and lower heating value (LHV).
Figure 6.6 Theoretical assessment of the mixture formation potential for 2-MF and ethanol during a cold start at 30 °C ambient temperature and 100 kPa ambient pressure.
Figure 6.7 Adiabatic flame temperature for 2-MF, DMF, and
iso
-octane at 120 °C with varying equivalence ratios Φ.
Figure 6.8 Laminar burning velocities of 2-MF, DMF, and
iso
-octane at different equivalence ratios Φ.
Figure 6.9 (a) Mass fraction burned at 3.5 bar IMEP, (b) mass fraction burned at 8.5 bar IMEP, (c) initial combustion duration from 3.5 to 8.5 bar IMEP, and (d) combustion duration from 3.5 to 8.5 bar IMEP for 2-MF, DMF, ethanol, and gasoline at an engine speed of 1500 rpm and
λ
= 1.
Figure 6.10 Gaseous emissions of the load variation at 2000 rpm with CR = 12 for all fuels and CR = 8.5 for RON 95.
Figure 6.11 Total PM size distribution.
Figure 6.12 Particulate matter emissions of the load variation at 2000 rpm with CR = 12 for all fuels and CR = 8.5 for RON 95.
Figure 6.13 Results of the load variation at
n
= 2000 rpm and CR = 12.
Figure 6.14 Engine-out emissions of diesel, biodiesel, GTL, and 2-MTHF under part-load operating conditions.
Figure 6.15 Model-based analysis of the influence of oxygen content, aromatic content, cetane number, and 50% boiling point on particulate emissions.
Figure 6.16 Emission results comparison of EN590 diesel and 2-MTHF, blended with 30% DNBE.
Figure 6.17 Thermogravimetric analysis of DPF samples for diesel, biodiesel, and 2-MTHF/DNBE blend.
Figure 6.18 Possible routes to transform furfural into furanic fuel compounds.
Figure 6.19 From furfural to 2-MF, lower pathway.
Figure 6.20 Production of FFA from 5-HMF.
Figure 6.21 From levulinic acid to 2-MTHF via γ-valerolactone.
Figure 6.22 From levulinic acid to 2-MTHF via γ-valerolactone.
Figure 6.23 Conceptual process design for a possible continuous 2-MTHF production process.
Chapter 7: Benzenoids
Figure 7.1 FLoL plotted against ID under conditions with and without EGR for diesel. (1) CHxnO (5%), (2) DBM, (3) CHxnO (9%), and (4) anisole. Note that scaling refers to a correction imposed onto the data to account for differing in-cylinder temperature and pressure histories.
Figure 7.2 Luminosity versus time for different fuels, normal (e.g., no EGR) inlet air conditions.
Figure 7.3 Soot plotted against NO
x
emissions at close to stoichiometric combustion for diesel and various blends with oxygenates with and without EGR [8].
Figure 7.4 Indicated efficiency plotted against lambda for diesel and various blends with oxygenates with and without EGR [8].
Figure 7.5 Volumetric fuel consumption plotted against lambda for diesel and various blends with oxygenates with and without EGR [8].
Figure 7.6 Soot plotted against NO
x
emissions at close to stoichiometric combustion for diesel and various blends with oxygenates with and without EGR [12].
Figure 7.7 Indicated efficiency plotted against lambda for diesel and various blends with oxygenates with and without EGR [12].
Figure 7.8 Van Krevelen diagram showing the H/C versus O/C ratios of different biomass and fossil materials [31].
Figure 7.9 Chemical structure of cellulose and lignin [32].
Figure 7.10 Overview of lignin conversion approaches in temperature/pressure space [34].
Figure 7.11 Influence of ethanol concentration on monomer yields for hydrothermal lignin conversion [40].
Figure 7.12 Influence of residence time on monomer yields for hydrothermal lignin conversion in THF.
Figure 7.13 Solvolysis of lignin in ethanol without (a) and with (b) a CuMgAlO
x
catalyst, reacting for 4 h at 573 K.
Figure 7.14 (a) Product distribution for various lignin conversion technologies at 673 K and 4 h reaction time. (Reproduced with permission Huang
et al.
[46] of Elsevier.). (b) Major compounds classes found in lignin oil produced via catalytic solvolysis [52].
Figure 7.15 Market volume and pricing for lignin and lignin-derived products. OS, organosolv lignin; BTX, benzene, toluene, xylene; *, black liquor and nonfermentables.
Chapter 8: Biomass Pyrolysis Oils
Figure 8.1 Impact of oxygen level on hydrogen consumption.
Figure 8.2 Value chain cost reduction versus residual oxygen in upgraded bio-oil.
List of Tables
Chapter 2: Fuel Class Higher Alcohols
Table 2.1 Physical and chemical properties of alcohol, gasoline, and diesel fuels
Table 2.2 Blending recommendations for
n
-butanol and
n
-pentanol with diesel fuel
Table 2.3 Summary of alcohol performance in SI engines
Table 2.4 Summary of alcohol performance in CI engines
Chapter 3: Fuel Class Valerates
Table 3.1 Properties of valerates as fuels
Table 3.2 Operating conditions for unregulated pollutant emission study
Table 3.3 Summary of the studies focused on the global performance of pure valerates and blend in IC engines
Table 3.4 Key performance indicators for the Lange
et al
. [28] three-step process
Table 3.5 Different processes to obtain valeric esters
Chapter 4: Butyl Ethers and Levulinates
Table 4.1 Fuel properties
Chapter 5: A Comprehensive Review of 2,5-Dimethylfuran as a Biofuel Candidate
Table 5.1 Properties of DMF, gasoline, and ethanol
Table 5.2 Catalytic reaction for HMF production using various catalysts
Table 5.3 Catalytic reaction for converting HMF to DMF
Table 5.4 Summary of the application of DMF in spark-ignition engines
Table 5.5 Summary of the application of DMF blends in compression-ignition engines
Chapter 7: Benzenoids
Table 7.1 Aromatics related fuel quality standards in Europe
Table 7.2 Neat benzenoid properties.
a
Table 7.3 Fuel oxygen content and CN of the tested fuels in Ref. [7]
Table 7.4 Fuel oxygen content and CN of the tested fuels in Ref. [12]
Table 7.5 Fuel oxygen content and CN of the tested fuels in Ref. [16]
Table 7.6 Summary of CI engine studies
Table 7.7 Impact of MAE on antiknock quality [21]
Table 7.8 Octane index and fuel economy of the tested benzenoid blends [21]
Table 7.9 Sensitivity, octane index and heating values for various benzenoids
Table 7.10 Summary of SI engine studies
Table 7.11 Cellulose/lignin content of selected biomass [30]
Table 7.12 Lignin building blocks [33]
Table 7.13 Hydrothermal processing of lignin to guaiacol
Table 7.15 Hydrothermal processing of lignin to 4-ethyl guaiacol
Table 7.16 Solvolysis of lignin to guaiacol
Table 7.17 Solvolysis of lignin to 4-methyl guaiacol
Table 7.18 Solvolysis of lignin to 4-ethyl guaiacol
Table 7.19 Catalytic solvolysis of lignin to guaiacol
Table 7.20 Catalytic solvolysis of lignin to methyl-guaiacol
Table 7.21 Catalytic solvolysis of lignin to ethyl-guaiacol
Table 7.22 Catalytic solvolysis of lignin benzyl alcohol
Chapter 8: Biomass Pyrolysis Oils
Table 8.1 Composition and physical property ranges for raw pyrolysis oils
Table 8.2 Composition of distillation fractions from a biomass pyrolysis oil hydrotreated to different oxygen levels [20]
Edited by Michael Boot
Biofuels from Lignocellulosic Biomass
Innovations beyond Bioethanol
Editor
Dr. Michael Boot
Eindhoven University of Technology
Department of Mechanical Engineering
Multiphase & Reactive Flows Group
Den Dolech 2 (Room GEM-N 1.23)
5612AZ Eindhoven
The Netherlands
All books published by Wiley-VCH are carefully produced. Nevertheless, authors, editors, and publisher do not warrant the information contained in these books, including this book, to be free of errors. Readers are advised to keep in mind that statements, data, illustrations, procedural details or other items may inadvertently be inaccurate.
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British Library Cataloguing-in-Publication Data
A catalogue record for this book is available from the British Library.
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The Deutsche Nationalbibliothek lists this publication in the Deutsche Nationalbibliografie; detailed bibliographic data are available on the Internet at <http://dnb.d-nb.de>.
© ,2016 Wiley-VCH Verlag GmbH & Co. KGaA, Boschstr. 12, 69469 Weinheim, Germany
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Print ISBN: 978-3-527-33813-9
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Cover Design Formgeber, Mannheim, Germany
List of Contributors
Stephen Arbogast
AOTA Energy Consultants LLC
Houston
TX 77204
USA
Robert M. Baldwin
National Renewable Energy Laboratory
15013 Denver West Parkway
Golden
CO 80401
USA
Don Bellman
AOTA Energy Consultants LLC
Houston
TX 77204
USA
Michael Boot
Eindhoven University of Technology
Department of Mechanical Engineering
Multiphase & Reactive Flows Group,
Den Dolech 2 (Room GEM-N 1.23)
5612AZ Eindhoven
the Netherlands
Francesco Contino
Vrije Universiteit Brussel
Faculty of Applied Sciences
Department of Mechanical Engineering Pleinlaan
2B-1050 Brussels
Belgium
Philippe Dagaut
CNRS - ICARE
1 C Avenue de la Recherche Scientifique 45081 Orléans cedex
France
Guillaume Dayma
CNRS - ICARE
1 C Avenue de la Recherche Scientifique 45081 Orléans cedex
France
Fabrice FOUCHER
Université d'Orléans - Laboratoire PRISME
8 rue Léonard de Vinci
45071 Orléans cedex
France
Fabien Halter
CNRS - ICARE
1 C Avenue de la Recherche Scientifique 45081 Orléans cedex
France
Benedikt Heuser
RWTH Aachen University
Institute for Combustion Engines (VKA),
Lehrstuhl für Verbrennungskraftmaschinen
Forckenbeckstr. 4
52074 Aachen
Germany
Bengt Johansson
King Abdullah University of Science and Technology (KAUST)
Clean Combustion Research Center, Building 5
Room 4335
Thuwal 23955-6900
Saudi Arabia
Florian Kremer
RWTH Aachen University
Institute for Combustion Engines (VKA)
Lehrstuhl für Verbrennungskraftmaschinen
Forckenbeckstr. 4
52074 Aachen
Germany
S. Mani Sarathy
King Abdullah University of Science and Technology (KAUST),
Clean Combustion Research Center
Chemical Engineering
Thuwal 23955-6900
Saudi Arabia
Robert L. McCormick
National Renewable Energy Laboratory
15013 Denver West Parkway
Golden
CO 80401
USA
Christine Mounaïm-Rousselle
Université d'Orléans - Laboratoire PRISME
8 rue Léonard de Vinci
45071 Orléans cedex
France
Dave Paynter
AOTA Energy Consultants LLC
Houston
TX 77204
USA
Stefan Pischinger
RWTH Aachen University
Institute for Combustion Engines (VKA),
Lehrstuhl für Verbrennungskraftmaschinen
Forckenbeckstr. 4
52074 Aachen
Germany
Chongming Wang
University of Birmingham
School of Engineering
Department of Mechanical Engineering,
Edgbaston Park Road,
Edgbaston
Birmingham B15 2TT
UK
Jim Wykowski
AOTA Energy Consultants LLC
Houston
TX 77204
USA
Hongming Xu
University of Birmingham
School of Engineering
Department of Mechanical Engineering,
Edgbaston Park Road, Edgbaston
Birmingham B15 2TT
UK
Preface
Soot, NOx, and low efficiency could be regarded as hallmarks of a fundamental mismatch between fossil fuels and internal combustion engines. This is not unsurprising considering the former were produced millions of years before the advent of the latter. In this respect, the emergence of ever-increasing biofuel mandates comes as a blessing in disguise; for now, we have the opportunity to blend in biofuels that might compensate for intrinsic fossil fuel deficiencies.
Governmentally imposed mandates, such as is currently the case for biofuels, are typically a testament to otherwise failed product–market combinations, suggesting that once again there might be a mismatch in what the engine would like to drink and what is on the menu. Consider, for example, ethanol. This biofuel, admittedly, has a high octane number, but the price for this benefit is high (e.g., 50% less energy per unit of volume, high heat of vaporization, hydrophilic nature). Again, a mismatch is to be expected given that fermentation of sugar to ethanol dates back as far as Neolithic times. In order to better address market demands, it might therefore prove worthwhile to first figure out what molecules the engine would actually like to drink and only then think about appropriate production processes from biomass.
Historically, however, the communication between producers and consumers of fuels has been poor to nonexistent. In fact, this book was first and foremost motivated by my astonishment that at the various biofuel conferences I attended, people who understood how an engine works were scarce to say the least. Tellingly, at engine technology conferences, few could explain how their tested biofuels were produced from biomass. This brings us to the goal of this book, which is to bridge the knowledge gap between biofuel production and biofuel combustion.
Following an introductory chapter on where future engine technology is headed, each subsequent chapter deals with a specific class of biofuels. The buildup is always the same, whereby first the choice for the fuels in question is motivated by engine experiments. The second part then goes on to discuss how the requisite molecules could be produced from biomass. Each chapter ends with a discussion on the future prospects of the fuel and a summary of the key conclusions.
NetherlandsFebruary 2016
Michael D. BootEindhoven University of Technology
Acknowledgments
The editor would first like to acknowledge the excellent contributions of all coauthors, without whom no book would have been possible. The financial support from the Dutch Technology Foundation (STW) and the Province of Noord-Brabant is also acknowledged.
Chapter 1Fuels and Combustion
Bengt Johansson
1.1 Introduction
All internal combustion engines use fuel as the source for heat driving the thermodynamic process that will eventually yield mechanical power. The fuel properties are crucial for the combustion process. Some combustion processes require a fuel that is very prone to ignition, and some have just the opposite requirement. Often, there is a discussion on what is the optimum. This optimum can be based on the fuel or the combustion process. We can formulate two questions:
What is the best possible fuel for combustion process x?
What is the best possible combustion process for fuel y?
Both questions are relevant and deserve some discussion, but it is very seldom that the fuel can be selected without any considerations, and similarly, there is only a limited selection of combustion processes to choose from. This brief introduction discusses the combustion processes and the link to the fuel properties that are suitable for them. Thus, it is more in the line of the first question of the aforementioned two.
1.2 The Options
For internal combustion engines, there are three major combustion processes:
Spark ignition (SI) with premixed flame propagation
Compression ignition (CI) with nonpremixed (diffusion) flame
Homogeneous charge compression ignition, HCCI with bulk autoignition of a premixed charge.
These three processes can be expressed as the corner points in a triangle according to Figure 1.1. Within this triangle, all practical concepts reside. Some are a combination of SI and HCCI, some a combination of SI and CI, and others a combination of CI and HCCI. We start by describing the basic three concepts and then move on to discuss the variations and the fuel implications that we can get with combined systems using, for instance, SI and CI at the same time.
Figure 1.1 The three fundamental combustion concepts, SI, CI, and HCCI.
The combined concepts to be discussed are as follows:
Spark-assisted compression ignition (SACI)
Partially premixed combustion (PPC)
Reactivity controlled compression ignition (RCCI)
Dual fuel
Prechamber flame ignition
Pilot-assisted compression ignition (PACI) or diesel pilot ignition.
The following section aims to give an introduction to the combustion processes and highlight the fuel requirements. Both chemical and physical properties are discussed.
1.3 Spark Ignition
The SI process is in principle very simple. Fuel and air are mixed and then the charge is compressed. Close to the piston top dead center, a spark is generated between two electrodes of a spark plug. This results in a locally very hot zone that starts exothermic reactions. Those reactions heat up the vicinity of the reaction zone and thus the reactions start there. The resulting propagating reaction zone is most often called a flame. The rate at which this flame propagates depends on the reactivity of the charge and how much the flame is distorted by turbulent eddies. Figure 1.2 illustrates the principle, with a spark plug to the right and a flame some 2/3 distance from the spark plug to the back wall of the combustion chamber.
Figure 1.2 Normal flame propagation [1].
(Reproduced with permission from Heywood [1] of McGraw Hill.)
The enhancement of flame speed by turbulence is not much affected by the fuel properties but by the reactivity of the charge. This reactivity is most often expressed as a laminar flame speed. Most hydrocarbon fuels exhibit a laminar flame speed at around 0.4 m/s close to that of stoichiometric mixtures, and this then drops to very low numbers as the mixture strength approaches the lean or rich dilution limits. In fact, those limits are defined when the laminar flame speed is zero. The major outliner when it comes to laminar flame speed is hydrogen. The laminar flame speed of hydrogen is very much higher than that of hydrocarbons. This means that a hydrogen engine can be expected to undergo a much faster combustion than the one using conventional fuels. Also, the alcohols tend to have a slightly higher flame speed, but here the difference to regular hydrocarbons is much less (<10%).
1.3.1 Uncontrolled SI Combustion, Knock
The conventional flame propagation with SI is not that fuel-dependent; except for hydrogen, most fuel behaves the same. Instead, the major problem comes when the combustion process is not working according to plan. With SI combustion, the fuel and air are mixed to burnable proportions well before the combustion starts. This means that the reactive mixture will be heated to a rather high temperature for some time. The worst conditions are for the end gas. This is the gas that is furthest away from the spark plug in the combustion chamber. This charge is heated and pressurized during the compression stroke but also by the increased pressure resulting from the combustion of the fuel closer to the spark plug. If the end gas charge is heated too much for too long, it will autoignite. Then the combustion will be very fast and all the end gas fuel/air mixture is converted into combustion products infinitely fast, at least in terms of time scales relevant for the engine. The autoignition is sometimes called knock from the engine noise if operated with autoignition. The noise comes from the pressure fluctuations resulting from the pressure wave originating at the autoignition site in the cylinder. If a moderate amount of fuel is burned with autoignition, the pressure oscillation amplitude will be a few bars only. This can be heard but will not be directly damaging. But if the autoignition is earlier, more fuel is burned in the autoignition process and the pressure oscillations can directly destruct the engine architecture. The high-amplitude pressure oscillations will also increase heat transfer to the combustion chamber walls by breaking down the thermal boundary layer. This will reduce engine efficiency and also heat the walls. If the walls are heated too much by prolonged knocking, the wall material will eventually reach a temperature where the mechanical strength will no longer be sufficient. When the piston material melts, there will be a permanent damage or even a hole, losing all compression and hence engine performance.
1.3.2 Autoignition of SI Engine Fuel
Spark ignition knock is determined by engine parameters such as flame travel length and compression ratio and also by engine operating conditions such as engine speed, load, and spark timing. But it is also fundamentally linked to the autoignition tendency of the fuel/air mixture in the cylinder. This is given by the fuel properties as such but also the fuel/air mixture strength, the amount of hot residual gases in the cylinder, the amount of exhaust gas recirculation (EGR), and other aspects affecting the reactivity of the charge.
To avoid SI engine knock, a fuel with a high resistance to autoignition should be selected. The number(s) most often used to quantify the knock tendency of a fuel is called the octane number. To make it a bit more complex, there are two, Research Octane Number (RON) and Motor Octane Number (MON). Both are extracted by operating a standardized engine, called a corporate fuel research (CFR), but with two sets of conditions. The first, RON, uses an engine speed of 600 rpm and an inlet temperature of 49 °C (150 °F) and fixed spark timing. For the MON, the engine speed is increased to 900 rpm, the inlet temperature is set to 149 °C (300 °F), and spark timing is adjusted for maximum knock.
When evaluating the octane number, the engine compression ratio is increased until a predefined level of knock is detected with a standardized microphone in the cylinder. The compression ratio is then compared to mixtures of two reference fuels: n-heptane, which is prone to autoignition, and iso-octane, which is much more resistant to autoignition. The mixture of n-heptane and iso-octane is adjusted until the same level of autoignition is reached. The percentage of iso-octane in the reference mixture is then defined as the tested fuel octane number. Thus, pure n-heptane will have an octane number of 0 by definition and pure iso-octane will have 100. But it should be noted that the scale between 0 and 100 is not linear. Figure 1.3 shows the compression ratio as a function of RON (and MON).
Figure 1.3 The required compression ratio for knock to be a function of RON and MON.
The required compression ratio as a function of fuel carbon atoms can be seen in Figure 1.4. The iso-octane and n-heptane are shown but also fuel with an octane number higher than that of iso-octane. Both methane, CH4, and benzene with the aromatic ring require higher compression ratios.
Figure 1.4 Required compression ratio as a function of octane number of RON [1].
(Reproduced with permission from Heywood [1] of McGraw Hill.)
The RON test was first introduced in 1928 when the effect of fuel autoignition tendency on SI engine performance was first realized. Later, it was argued that the conditions for RON were less suitable and a more severe test was generated. This is called the Motor Octane Number. The same engine is used as with RON, but the inlet temperature is increased from 49 to 149 °C (300 °F) and the engine speed is increased to 900 rpm. There is also a significant difference in how the inlet temperature is measured. For the RON case, it is the air temperature before the fuel is added to the inlet, whereas for the MON case, it is the fuel/air mixture temperature. Thus, if a fuel would have significant heat of vaporization, leading to cooling of the charge as the fuel evaporates, a clear difference between RON and MON is expected.
For most practical hydrocarbons, the RON is higher than MON, often around 5–10 units. This has led to different definitions of octane number worldwide. In the European Union, the RON is used and there are no requirements of MON. In the United States, the average of RON and MON is used and is called Octane Index. As MON is most often lower, the numbers found on filling station fuel pumps are different in the European Union and the United States. A fuel rated as 87 in the United States can be equivalent to 92 in the European Union.
There has been some debate about the suitability of using both RON and MON in the United States, especially from Gautam Kalghatgi from Shell (and since a few years, Saudi Aramco) [2]. The argument is that engines of today are using a rather high boost pressure and an intercooler. This means that the condition in the cylinder after compression is different from the RON and especially MON cases. With high boost and effective intercooling, the pressure level is high, whereas the temperature is moderate, and hence the T/p ratio is low. This is in stark contrast to the MON case where the inlet temperature is high and compression ratio moderate, yielding moderate pressure after compression. For the MON case, the T/p ratio will be very high. The RON condition will be something in between with a moderate T/p ratio. It is thus argued that for a given RON number, a fuel with a lower MON will ignite easily at high temperature and low pressure (high T/p-ratio) and, as a consequence, ignite less easily at low temperature and high pressure (low T/p-ratio). In other words, for a given RON, a fuel with a low MON will be less prone to autoignition at high pressure and low temperature. This is how modern engines are operated.
1.3.3 Physical Properties of SI Engine Fuel
Apart from the chemical properties, SI engine fuel must also exhibit some suitable physical properties. The most important, for liquid fuel, is the boiling point range. Most SI engines still comprise carburetors or fuel injection systems that add the fuel in the inlet system. Thus, a significant fraction of the fuel is deposited on the inlet walls and needs to be evaporated to enter the cylinder. During cold start, the walls are cold and thus fuel enrichment is needed to get sufficient fuel amount into the cylinder. Here the light fraction of the fuel is very important, as it is this fraction that will find its way to the cylinder. Also, for a heated-up engine, the fuel needs to boil off the inlet walls. Thus, the final boiling point (T90) must be reasonably low, around 180 °C. The boiling point range for commercial gasoline is specified in EU within the EN228 fuel specification.
Another important factor for SI engine fuels is the heat of vaporization. Most gasoline exhibits significant heat of vaporization, but some fuels such as ethanol and especially methanol exhibit much more. The heat of vaporization can be used to increase engine performance. The cooling effect of the fuel evaporating can be used to get more air into the cylinder and hence burn more fuel. But it can also be used as an “internal intercooler” reducing charge temperature and hence reducing the risk of knock.
It can be noted that most SI engines are using low-pressure fuel injection and thus can also be adapted to gaseous fuels without major problems. With gaseous fuels, the fuel is mixed with air in the inlet, or it can be injected into the cylinder very early after inlet valve closing if additional engine performance is required.
1.4 Compression Ignition
The combustion process in CI engines is totally different from that in SI engines. Only air is compressed, and close to top dead centre, TDC, a high-pressure fuel is sprayed into the cylinder. After some initial mixing, the combustion starts with a rather uniform, very rich (λ = 0.2) autoignition. The mixture is so rich that the reaction rate is not extremely fast. After the initial homogeneous, rich autoignition, diffusion-controlled combustion process occurs. Here the fuel is first mixed with air up to the lift-off position where a premixed rich flame is formed. After the flame formation, additional reactions including soot formation take place inside the spray plume and then a thin zone close to stoichiometric reaction zone is formed in the outer boundaries of the plume. Here most of the NOx is formed and much oxidation of soot results. Figure 1.5 shows the conceptual model presented by John Dec in 1997 [3].
Figure 1.5 Conceptual model of diesel combustion by John Dec.
(Reproduced with permission from Dec [3] of SAE International.)
1.4.1 Autoignition of CI Engine Fuel
The fuel used in a CI engine is vastly different from that in an SI engine. In an SI engine, the fuel should sustain high pressure and temperature without autoignition. The exact opposite is true for the CI engine. Here the fuel should autoignite as fast as possible after injection has started. This will reduce the amount of fuel burned in the rich premixed region and thus also reduce the first initial spike in the rate of heat release curve. CI engine fuel is rated in the same way as the octane scale is used for SI combustion. The same CFR engine is used but modified to have a direct injection (DI)-type fuel injection system with moderate pressure. The fuel is injected at 13 CAD before TDC and then the compression ratio is adjusted to start autoignition at TDC. Hence, the engine is operated with 13 CAD of ignition delay. The reference fuels are not the same as for the octane number. For CI operation, it is cetane (now called hexadecane, C16H34) that is given a number of 100 and α-methyl naphthalene is given 0.
An ignition delay of 13° is more than that of modern CI engines. In combination with an injection pressure of only 300 bar, almost an order of magnitude less than that of modern CI engines, it is argued that fuel ratings in the CFR is less representative of modern engines. Thus, an alternative way of characterizing CI engine fuel has been introduced. It is called derived cetane number (DCN) and extracted from a device called ignition quality tester (IQT). This is a constant-volume chamber heated to 575 °C. The fuel is injected under this constant condition, and the resulting ignition delay is then used as a direct measure of the cetane number (CN). Thus, it is in principle enough to inject once to the IQT to obtain a reading and hence much less fuel is needed to obtain the DCN and the CN from the CFR engine. There are also indirect ways to estimate the fuel autoignition tendency by measuring the fuel density and boiling point range. With density and boiling point in four points, a cetane number, CI, can be extracted. This does not take any ignition improver additives into account and will not be a direct measurement of autoignition.
1.4.2 Physical Properties of CI Engine Fuel
Also, for the physical properties, the requirements of fuel for CI and SI engines are in stark contrast. The CI engine works with a very high pressure fuel injection system to get the spray fully atomized and to get fast mixing. A pressure of 2000–3000 bar induces a very high load on the mechanical components in the fuel system and the risk of cavitation is always there. Thus, the fuel must exhibit suitable properties for high-pressure compression. It is important that the fuel exhibits a minimum viscosity; otherwise, it will leak past the fuel pump elements. It should also be incompressible. This is in principle the case of all liquids, but some fuels are rather compressible. A typical case is dimethyl ether (DME). It is very hard to compress DME above 500 bar due the compressibility of the fuel [4]. As a CI engine fuel pump should be able to compress the fuel to a pressure much higher than the cylinder pressure, all gaseous fuels will result in significant parasitic losses.
In practice, the physical and chemical properties of most fuels are linked in a favorable way. Smaller lighter molecules most often also have a high octane number and are thus suitable for SI engine operation. If we focus on conventional fossil fuels, the lightest hydrocarbon would be methane. This has a high RON (130) and is easy to add to the inlet of an SI engine. Also, ethane, propane, and butane, often called liquid petroleum gas (LPG), have high octane numbers and are gaseous under ambient conditions. Crude oil fractions up to roughly eight carbon atoms per molecule have a boiling point below 180 °C and can be directly used as gasoline components. The breakpoint of 180 °C occurs to coincide with the lowest boiling point of diesel fuel for the CI engine. The heavier fractions tend to have a lower octane number (and hence higher cetane number) and are thus well suited for the CI engine.
A very oversimplified model of how to produce gasoline and diesel fuel is thus as follows:
1.
Drill a hole in the ground.
2.
Pump the crude oil out from the hole.
3.
Pour the crude oil in a distilling pot and heat the pot to 180 °C.
4.
What boils off is condensed in a separate cooled pot and is called
gasoline
.
5.
Heat the pot to 350 °C.
6.
What boils off is condensed in a separate cooled pot and is called
diesel fuel
.
7.
What stays in the pot is called
heavy fuel oil
or
possibly asphalt
.
1.5 Highly Diluted Autoignition, HCCI
Apart from the dominating SI and CI engines, it is also possible to operate with the third type of combustion: autoignition. This is often called homogeneous charge compression ignition. With HCCI, the fuel and air are fully premixed before combustion as in the SI engine, but combustion is started by the increased pressure and temperature during the compression stroke.
The combustion starts in many zones at the same time and can be called a distributed reaction. It is incorrect to name it homogeneous combustion though as the combustion process shows significant inhomogeneity with some zones burning much faster than the other zones. Figure 1.6 shows three images of the combustion process from an optical engine operated in HCCI mode. In this particular engine, a Scania 2-l cylinder truck engine, the combustion had a tendency to ignite in the outer regions, and only later, the reactions start in the center [5].
Figure 1.6 HCCI combustion at 7, 9, and 11 CAD after top dead centre (ATDC) in a single cycle.
(Reproduced with permission from Vressner et al. [5] of SAE International.)
With HCCI, the charge is very much diluted to make sure the reactivity is moderate. This is needed as the combustion rate would otherwise be extreme, ruining the engine structure possibly in a single cycle with excessive pressure rise rate and peak in-cylinder pressure. This also enables a very efficient thermodynamic cycle and hence HCCI can show much better total fuel efficiency [6] and an SI engine operated with the same fuel in the same engine.
The combustion process in HCCI is very sensitive and can be considered a balance in temperature. The fully premixed change will break down and combustion start when the temperature and pressure are high enough for a prolonged period. For convenience, it is often expressed as an autoignition temperature; once it is reached, the reaction starts. For iso
