Knocking Combustion of Methane-Based and Highly Knock Resistant Liquid Fuels (Wissenschaftliche Reihe Fahrzeugtechnik Universität Stuttgart) 3658351772, 9783658351779

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Wissenschaftliche Reihe Fahrzeugtechnik Universität Stuttgart

Marcel Eberbach

Knocking Combustion of Methane-Based and Highly Knock Resistant Liquid Fuels

Wissenschaftliche Reihe Fahrzeugtechnik Universita¨ t Stuttgart Reihe herausgegeben von Michael Bargende, Stuttgart, Deutschland Hans-Christian Reuss, Stuttgart, Deutschland Jochen Wiedemann, Stuttgart, Deutschland

Das Institut für Fahrzeugtechnik Stuttgart (IFS) an der Universität Stuttgart erforscht, entwickelt, appliziert und erprobt, in enger Zusammenarbeit mit der Industrie, Elemente bzw. Technologien aus dem Bereich moderner Fahrzeugkonzepte. Das Institut gliedert sich in die drei Bereiche Kraftfahrwesen, Fahrzeugantriebe und Kraftfahrzeug-Mechatronik. Aufgabe dieser Bereiche ist die Ausarbeitung des Themengebietes im Prüfstandsbetrieb, in Theorie und Simulation. Schwerpunkte des Kraftfahrwesens sind hierbei die Aerodynamik, Akustik (NVH), Fahrdynamik und Fahrermodellierung, Leichtbau, Sicherheit, Kraftübertragung sowie Energie und Thermomanagement – auch in Verbindung mit hybriden und batterieelektrischen Fahrzeugkonzepten. Der Bereich Fahrzeugantriebe widmet sich den Themen Brennverfahrensentwicklung einschließlich Regelungs- und Steuerungskonzeptionen bei zugleich minimierten Emissionen, komplexe Abgasnachbehandlung, Aufladesysteme und -strategien, Hybridsysteme und Betriebsstrategien sowie mechanisch-akustischen Fragestellungen. Themen der Kraftfahrzeug-Mechatronik sind die Antriebsstrangregelung/Hybride, Elektromobilität, Bordnetz und Energiemanagement, Funktions- und Softwareentwicklung sowie Test und Diagnose. Die Erfüllung dieser Aufgaben wird prüfstandsseitig neben vielem anderen unterstützt durch 19 Motorenprüfstände, zwei Rollenprüfstände, einen 1:1-Fahrsimulator, einen Antriebsstrangprüfstand, einen Thermowindkanal sowie einen 1:1-Aeroakustikwindkanal. Die wissenschaftliche Reihe „Fahrzeugtechnik Universität Stuttgart“ präsentiert über die am Institut entstandenen Promotionen die hervorragenden Arbeitsergebnisse der Forschungstätigkeiten am IFS. Reihe herausgegeben von Prof. Dr.-Ing. Michael Bargende Lehrstuhl Fahrzeugantriebe Institut für Fahrzeugtechnik Stuttgart Universität Stuttgart Stuttgart, Deutschland

Prof. Dr.-Ing. Hans-Christian Reuss Lehrstuhl Kraftfahrzeugmechatronik Institut für Fahrzeugtechnik Stuttgart Universität Stuttgart Stuttgart, Deutschland

Prof. Dr.-Ing. Jochen Wiedemann Lehrstuhl Kraftfahrwesen Institut für Fahrzeugtechnik Stuttgart Universität Stuttgart Stuttgart, Deutschland

Weitere Bände in der Reihe http://www.springer.com/series/13535

Marcel Eberbach

Knocking Combustion of Methane-Based and Highly Knock Resistant Liquid Fuels

Marcel Eberbach Institute of Automotive Engineering (IFS), Chair of Vehicle Drives University of Stuttgart Stuttgart, Germany Zugl.: Dissertation Universität Stuttgart, 2020 D93

ISSN 2567-0042 ISSN 2567-0352 (electronic) Wissenschaftliche Reihe Fahrzeugtechnik Universität Stuttgart ISBN 978-3-658-35177-9 ISBN 978-3-658-35178-6 (eBook) https://doi.org/10.1007/978-3-658-35178-6 © The Editor(s) (if applicable) and The Author(s), under exclusive license to Springer Fachmedien Wiesbaden GmbH, part of Springer Nature 2021 This work is subject to copyright. All rights are solely and exclusively licensed by the Publisher, whether the whole or part of the material is concerned, specifically the rights of translation, reprinting, reuse of illustrations, recitation, broadcasting, reproduction on microfilms or in any other physical way, and transmission or information storage and retrieval, electronic adaptation, computer software, or by similar or dissimilar methodology now known or hereafter developed. The use of general descriptive names, registered names, trademarks, service marks, etc. in this publication does not imply, even in the absence of a specific statement, that such names are exempt from the relevant protective laws and regulations and therefore free for general use. The publisher, the authors and the editors are safe to assume that the advice and information in this book are believed to be true and accurate at the date of publication. Neither the publisher nor the authors or the editors give a warranty, expressed or implied, with respect to the material contained herein or for any errors or omissions that may have been made. The publisher remains neutral with regard to jurisdictional claims in published maps and institutional affiliations. This Springer Vieweg imprint is published by the registered company Springer Fachmedien Wiesbaden GmbH part of Springer Nature. The registered company address is: Abraham-Lincoln-Str. 46, 65189 Wiesbaden, Germany

Preface This work was realized during my tenure as a research associate at the Institute of Automotive Engineering (IFS) at the University of Stuttgart under the supervision of Prof. Dr.-Ing. M. Bargende. My sincere gratitude goes to Prof. Dr.-Ing. M. Bargende for his excellent guidance and scientific support during this process. I would like to thank Prof. Dr.-Ing. F. Atzler for his interest and being the co-referee of this work. This work would not have been the same without the valuable support from various people at the IFS and the Research Institute of Automotive Engineering and Vehicle Engines Stuttgart (FKFS). In particular, I would like to thank HansJürgen Berner for guiding my work and for his patient support. Furthermore, I am grateful to Dr. Sebastian Scharlipp, Dr. Philipp Skarke, Dr. Christian Auerbach, Simon Hummel, and all other colleagues for the cooperation and assistance during time together at the institute. Finally, I would like to thank my family and friends who supported me with understanding during the interesting even though challenging time. Stuttgart

Marcel Eberbach

Contents Preface......................................................................................... V Figures........................................................................................ IX Tables ........................................................................................ XV Abbreviations ............................................................................XVII Symbols ....................................................................................XXI Abstract ................................................................................. XXIII Kurzfassung ........................................................................... XXVII 1

Introduction ............................................................................. 1

2

Fundamentals of Self-Ignition Processes in SI Engines ................... 5 2.1 2.2 2.3 2.4 2.5 2.6

3

Experimental Preparations....................................................... 45 3.1 3.2 3.3 3.4

4

Spark Ignition Engines ......................................................... 5 Abnormal Combustion Phenomena ......................................... 6 Reaction Kinetics of Irregular Combustion ............................. 13 Cylinder Pressure-Based Knock Detection.............................. 24 Influencing Parameters of Knocking Combustion ..................... 31 Knock Resistance of Fuels .................................................. 35

Investigated Fuels ............................................................. 45 Fuel Injection ................................................................... 52 Gas-Mixing System ........................................................... 54 Engine Test Bench ............................................................ 56

Engine Measurements ............................................................. 63 4.1 4.2 4.3

Definition of Engine Operating Points ................................... 63 Knock Detection Algorithm ................................................ 64 Test Bench Measuring Procedure.......................................... 68

VIII 5

Experimental Results and Evaluation ........................................ 71 5.1

5.2

5.3 6

Contents

Evaluation of the Knock Rate .............................................. 71 5.1.1 Methane-Ethane Mixtures ........................................ 73 5.1.2 Liquid Fuels .......................................................... 77 Investigation of Atypical Knocking Combustion Phenomena ...... 81 5.2.1 Onset of Knocking Combustion ................................. 82 5.2.2 Comparison of Knocking Behavior of Different Fuels at Similar MFB505 % Position.................................... 87 5.2.3 Distinct Second Heat Release Peak............................. 89 5.2.4 Non-Oscillating Knocking Working Cycles .................. 96 5.2.5 Influencing Parameters on the Atypical Knocking Combustion Phenomena..........................................100 Motor Methane and Octane Number Correlation.....................106

Conclusion and Outlook..........................................................115

Bibliography ...............................................................................119 Appendix....................................................................................127 A.1 Appendix 1.....................................................................127 A.2 Appendix 2.....................................................................132 A2.1 Knock Rate versus MFB50 ......................................132 A.3 Appendix 3.....................................................................136 A.4 Appendix 4.....................................................................140 A4.1 Methane-Ethane MMN – MN Correlation...................143

Figures P.1

Heat release rate, standardized of a gaseous knocking working cycle .............................................................................. XXIII P.2 Knock rate vs. 50 % mass fraction burned of a gaseous and a liquid fuel within a spark angel sweep............................................. XXV P.3 Filtered heat release rate (3 kHz-filtered) of one liquid fuel and two gaseous fuels: EOP1: n = 1500 rpm ...................................... XXVI 1.1 Heat release rate, standardized of a gaseous and a liquid knocking working cycle ......................................................................... 2 2.1 Shadowgraph photographs of knocking combustion: 33 μs between frames(left). Photodigitized picture (right) showing additional details of the self-ignition location [49]............................................ 8 2.2 Unfiltered cylinder pressure traces of individual engine cycles with and lacking the typical oscillations............................................. 12 2.3 Schlieren photographs of (a) normal flame propagation through the end gas and (b) knocking combustion (autoignition occurs in photograph 4), with corresponding cylinder pressure [29] ............... 14 2.4 Simplified temporal progression of the temperature during thermal and chain-branching explosion in an adiabatic system [72].............. 15 2.5 Calculated (line) and measured (points) ignition-delay times in hydrocarbon mixtures [72] .......................................................... 16 2.6 Explosion diagram for hydrocarbons (schematically) [72]............... 18 2.7 Qualitative progressions: a) two-stage ignition, b) temperature dependence of the ignition delay for a n-heptane-air-mixture [63]........ 19 2.8 Comparison of ignition delays for different stoichiometric air / fuel mixtures at p = 13 bar [20]....................................................... 21 2.9 Comparison of ignition delays for different stoichoimetric air / fuel mixtures at p = 13 bar [20]....................................................... 22 2.10 Ignition diagram. Two-stage ignition occurs in the low-temperature region; first stage may be a cool flame. Single-stage ignition occurs in the high-temperature region [29]...................................... 23

X

Figures

2.11 Cylinder pressure trace (left) and according oscillations of the bandpass filtered pressure trace [16] ................................................. 28 2.12 HRR traces of a binary gaseous mixture, non-, slightly-, strongly knocking combustion (3 kHz-filtered) ........................................ 30 2.13 Laminar burning velocity for different fuels as function of equivalence ratio; ambient pressure; 300 K. [29] ................................ 34 2.14 Critical compression ratio versus the number of carbon atoms in a molecule for different hydrocarbons [29] .................................... 36 2.15 Dependency of RON according to the molecular structure changes of hydrocarbons [72] .............................................................. 37 2.16 Measured MON of natural-gasoline blends as a function of the mass percentage natural gas [22] ............................................... 41 2.17 Research octane number as a function of the admixed amount of the anti-knock additive TEL [2] ................................................ 42 3.1 Functional structure of the gas mixing system (GMS) .................... 55 3.2 The single cylinder research engine [45] ..................................... 56 3.3 Piston and cylinder head disassembled and cleaned; installed spark plug and pressure transducer; squeeze crevice between cylinder head and piston (left) .............................................................. 57 3.4 Engine test bench, inlet air manifold side .................................... 58 3.5 Overview test bench setup, schematically .................................... 60 4.1 Pressure trace evaluation method of the VDO knock detection algorithm [55] ......................................................................... 65 4.2 Knocking working cycles with the typical oscillation (top) and lacking the typical oscillations (bottom) [55]................................ 67 4.3 Evaluation procedure of the knock intensity by means of the amplitude ratio [55] ..................................................................... 68 5.1 Evaluation procedure to determine the MFB505 % position .............. 72 5.2 Methane-ethane mixtures EOP1: n = 1500 rpm, MN = 48 – 100, IMEP = 18 bar to 19 bar .......................................................... 74 5.3 Methane-ethane mixtures: n = 2000 rpm (top) and n = 3000 rpm (bottom), MN = 48 – 100, IMEP = 18 bar to 19 bar (top) ................ 75 5.4 Individual working cycles showing abnormal combustion; HRR is already increasing before the spark ignites the mixtures EOP1: n = 3000 rpm, MN = 48 ........................................................... 77 5.5 Isooctane-TEL: n = 1500 rpm, p2 = 1000 mbar, T2 = 45 ◦C ............. 78

Figures 5.6 5.7 5.8

5.9

5.10

5.11 5.12

5.13

5.14

5.15

5.16

5.17

XI

Selected liquid test fuels EOP1: n = 1500 rpm (top) n = 2000 rpm (bottom), RON compare table 3.4 on page 51 ............................... 80 Selected liquid test fuels EOP1: n = 3000 rpm, RON compare table 3.4 on page 51....................................................................... 81 HRR originally (black), HRR low pass filtered (grey) and unfiltered cylinder pressure trace of a knocking working cycle at EOP1: n = 1500 rpm; compressed natural gas (CNG); MN = 87 ....... 83 Onset of knocking combustion vs. MFB50 position as a function of the amplitude ratio (knock intensity) according to the extended algorithm for all examined gaseous fuels at EOP1; cycle-based evaluation............................................................................. 84 Onset of knocking combustion vs. MFB50 position as a function of the amplitude ratio (knock intensity) according to the extended algorithm for all examined liquid fuels at EOP1; cycle-based evaluation .................................................................................. 85 Progression of the knock rate for a CH4 / C2 H6 mixture and toluene EOP1: n = 1500 rpm, similar MFB505 % position .......................... 87 Methane-ethane mixture and toluene for EOP1: n = 1500 rpm, MFB505 % = 9.5 ◦ CA a. TDC, unfiltered pcyl trace and filtered HRR (3 kHz-filtered) ...................................................................... 88 Knock intensity of gaseous fuels (standardized, cycle-based) according to the VDO-algorithm; identified knocking working cycles are colored in grey ................................................................. 90 Knock intensity of gaseous fuels (standardized, cycle-based) according to the extended algorithm; identified knocking working cycles are colored dark grey ..................................................... 91 Knock intensity of liquid fuels (standardized, cycle-based) according to the VDO-algorithm; identified knocking working cycles are colored in grey ...................................................................... 92 Knock intensity of liquid fuels (standardized, cycle-based) according to the extended algorithm; identified knocking working cycles are colored dark grey .............................................................. 93 Exclusively as knock identified IWC assigned to the respective detection algorithm for gaseous fuels (top) liquid fuels (bottom): Intensity according to the amplitude ratio vs. MFB50 .................... 94

XII

Figures

5.18 Exclusively as knock identified IWC assigned to the respective detection algorithm for liquid fuels: Intensity according to VDO evaluation method vs. MFB50 .................................................. 95 5.19 Proportion of low-oscillating knocking working cycles for liquid and gaseous fuels versus the engine speeds .................................. 96 5.20 Proportion of low-oscillating knocking working cycles for liquid and gaseous fuels versus the MN respectively RON ....................... 97 5.21 Low-oscillating cylinder pressure trace and Schlieren photographic sequence showing of deflagration similar combustion of a hot spot auto-ignited working cycle (left to right, 1.08 ◦ CA) [37] ................. 99 5.22 Characteristics of an end gas pocket according to [36] ..................101 5.23 Relative position of the onset of knock to the pressure reversal point as a function of the MFB50 position for all gaseous fuels at EOP1, cycle-based ................................................................103 5.24 Relative position of the onset of knock to the pressure reversal point as a function of the MFB50 position for all liquid fuels at EOP1, cycle-based ................................................................104 5.25 Relative position of the 2nd HR-peak to the pressure reversal point as a function of the MFB50 position for all gaseous fuels at EOP1, cycle-based..........................................................................105 5.26 Relative position of the 2nd HR-peak to the pressure reversal point as a function of the MFB50 position for all liquid fuels at EOP1, cycle-based..........................................................................106 5.27 Methane-hydrogen primary reference fuels EOP1: n = 2000 rpm, MN = 50 – 100 .....................................................................108 5.28 Regression-straight: MFB505 % positions as a function of the H2 content in the mixture for methane-hydrogen fuel matrix at EOP1: n = 2000 rpm, MN = 50 – 100..................................................109 5.29 Regression-straight: MFB505 % positions as a function of the H2 content in the mixture for methane-hydrogen fuel matrix at EOP2: n = 2000 rpm, MN = 50–100 ...................................................110 5.30 Motor methane number to the corresponding liquid fuel at EOP1: n = 1500 rpm, 2000 rpm and 3000 rpm .......................................111 5.31 Deviation of the MMN relative to the calculated AVL MN for methane-ethane mixtures at EOP1: n = 1500 rpm, 2000 rpm and 3000 rpm.............................................................................112

Figures

XIII

5.32 Methane number according to the RON values found in the literature. Evaluated as a function of MMN......................................113 A1.1 Influence of inlet air pressure and temperature variation on the knocking rate as a function of MFB50 position: EOP1; methane .....128 A1.2 Fuel temperature and rail pressure variation for toluene: EOP1, n = 1500 rpm .......................................................................129 A1.3 Knocking rate vs. the MFB50 position as a function of engine speeds: EOP1, methane ..........................................................130 A1.4 Knock rate vs. the MFB50 position as a function of λ = 0.8 to 1.4: EOP1, methane (top) and toluene (bottum) .................................131 A2.1 Methane-ethane mixtures: n = 2000 rpm, IMEP = 14 bar to 15 bar ..132 A2.2 Isooctane-TEL mixture: n = 2000 rpm (top) and 3000 rpm (bottom), p2 = 1000 mbar, T2 = 45 ◦C RON = 100 – 125 .....................133 A2.3 Liquid (acetone) and gaseous (methane) fuel with similar MFB505 % position; n = 1500 rpm, heat release (3 kHz-filtered) .....................134 A2.4 Comparison of a liquid fuel and gaseous fuels: EOP1, n = 1500 rpm, filtered HRR (3 kHz-filtered) ...................................................135 A3.1 Methane-hydrogen primary reference fuels: EOP1, n = 2000 rpm (top) and 3000 rpm (bottom), MN = 40 – 100 ..............................137 A3.2 Knock rate vs. MFB50 position as a function of content H2 in the methane-hydrogen mixture 1 % to 8 % for EOP2, n = 1500 rpm (top), 2000 rpm (bottom) ........................................................138 A3.3 Knock rate vs. MFB50 position as a function of content H2 in the methane-hydrogen mixture 1 % to 8 % for EOP2, n = 3000 rpm ......139 A4.1 Regression-straight: MFB505 % positions vs. the H2 content of methane-based mixture at EOP1: n = 1500 rpm (top) and 3000 rpm (bottom), MN = 50 – 100........................................................141 A4.2 Regression-straight: MFB505 % positions vs. the H2 content of methane-based mixture at EOP2: n = 1500 rpm (top) and 3000 rpm (bottom), MN = 60 – 100........................................................142 A4.3 Deviation of the MMN relative to the calculated AVL MN for methane-ethane mixtures at EOP2: n = 1500 rpm, 2000 rpm and 3000 rpm.............................................................................143

Tables 2.1 3.1

3.2 3.3 3.4 3.5 3.6 4.1

Testing methods to determine the RON, MON and MN [2] ............. 40 Deviations resulting for different combinations of possible manufacturing accuracy and analytical fault tolerance exemplary for the methane-ethane mixture with a MN grade of 74.8 ......................... 47 Methane-ethane mixtures and production tolerances ...................... 48 Liquid fuel matrix: Tetraethyl lead admixed in isooctane ................ 50 Specifications of the selected liquid test fuels [29] [26] [19] [70] [19] [73] [7] [1] [57] [51] [30][48] [32] ...................................... 51 Fuel matrix: Methane-hydrogen mixtures in mol % with the resulted MN deviation ................................................................... 52 Specifications of engine and test bench ....................................... 61 Engine operating points........................................................... 64

Abbreviations a. TDC AG ASTM AT AVL

after top dead center stock corporation American Society for Testing and Materials analytical tolerance Institute for Internal Combustion Engines List

b. TDC BasicPg

before top dead center basic noise level

CA CFR CNG CR

crank angle Cooperative Fuel Research compressed natural gas compression ratio

DCPD

dicyclopentadiene

EOP1 EOP2

engine operating point 1 engine operating point 2

FKFS FVV

Research Institute of Automotive Engineering and Vehicle Engines Stuttgart Research Association for Combustion Engines

GMS

gas mixing system

HDEV HRR

high pressure injection valve heat release rate

IAV IFS IMAT

Engineering Company Car and Traffic Institute of Automotive Engineering intake manifold air temperature

XVIII

Abbreviations

IMEP IT ITS IWC

indicated mean effective pressure ignition timing ignition time sweep individual working cycle

KI KI20 KO KR KRAT KSI

knock intensity knock intensity factor knock onset knock rate knock ratio knock serverity index

MA MFB MFB50 MFB505 % MFC MMN MN MON MTBE

manufactuing accuracy mass fraction burnt 50 % mass fraction burnt 50 % mass fraction burnt at 5 % knock rate mass flow controller motor methane number methane number motor octane number methyl tert-butyl ether

NTC

negative temperature coefficient

OI ON

octane index octane number

PFI PRF PRP PSD PTA

port fuel injection primary reference fuel pressure reversal point power spectral density pressure trace analysis

RON

research octane number

SI

spark ignition

Abbreviations

XIX

TDC TEL

top dead center tetraethyl lead

VDO

United DEUTA-OTA

WOT

wide open throttle

Symbols Latin Letters A A a b B cp cv Ea ΔHR Δvap H ◦ HG Hl Hu k Lst mcyl n n P p p2 p3 pi pmean ΔQrel,max Qb Qh Qw R T

pre-exponential factor Arrhenius eq. 2.7 fitted fuel parameter eq. 2.1 pre-exponetial factor fitting curve parameter of fitting curve fitted fuel parameter eq. 2.1 heat capacity at constant pressure heat capacity at constant volume activation energy for the reaction reaction enthalpy evaporation enthalpy at normally conditions lower heating value of mixture leakage heat flux lower heating value of fuels rate constant (frequency of collisions) stoichiometric combustion air demand total cylinder mass fitted fuel parameter eq. 2.1 engine speed power pressure inlet air pressure exhaust pressure pressure amplitude of oscillation i eq. 2.16 mean pressure knock intensity eq. 2.16 relativ maximum HRR amplitude ratio total heat release fuel total heat release wall heat flux ideal gas constant temperature

ms barn K J/kg/K J/kg/K J/kmol kJ kJ/mol MJ/m3 kJ MJ/kg kg/kgfuel kg rpm W Pa bar bar bar bar kJ/◦ CA kJ kJ kJ J/kmolK K

XXII t T2 T3 U V

Symbols time inlet air Temperature exhaust Temperature total internal energy volume

s ◦C ◦C kJ m3

Greek Letters ε η λ φ ϕ ρ τ

compression ratio efficiency air / fuel ratio fuel / air equivalence ratio crank angle density ignition delay Indices

cyl m max mean mech norm o p R Sample

cylinder meta maximum arithmetic mean mechanical normalized ortho para reaction sample

°CA kg/m3 s

Abstract During the investigations of the knocking behavior of methane-based fuels in the predecessor project, atypical combustion phenomena in terms of anomalies on the specific heat release trace have been observed by thermodynamic analysis which could not have been explained with today’s experience and understanding regarding the knocking behavior of gasoline fuels.

300

Methane-based fuel

250 200 150 100 50 0 -30

-15 0 15 30 45 Crank angle position in °CA a. TDC [ ]

60

Figure P.1: Heat release rate, standardized of a gaseous knocking working cycle

In particular, three distinct heat release phenomena during the knocking operation are relevant with regards to the extended investigations in the present project: The first phenomenon is an unusually early onset of knock within the

XXIV

Abstract

working cycles. Compare the 50 % mass fraction burnt (MFB50) position relative to the onset of knock in figure P.1. Such an early onset of knock means that a sizeable volume of unburned fuel mixture is still present in the end gas at the time of self-ignition which, in the common understanding of knocking, should lead to severe engine damages which, however, have not been observed in the predecessor project. The second knocking phenomenon can be described as the development of a distinct second heat release amplitude which suddenly develops with the onset of knocking combustion while the “regular” combustion is already decreasing. The third combustion phenomenon is the sporadic occurrence of self ignited working cycles without significant oscillation amplitudes on the pressure trace. Due to the fact that conventional knock detection algorithms are based on the evaluation of these oscillation amplitudes, individual self-ignited working cycles will not be detected. Therefore, an additional knock detection algorithm is used for the investigations, which takes into account the amount of fuel converted by knocking combustion and therefore also reliably detects self-ignited working cycles without any significant oscillation amplitudes. In order to further analyze the fuel-specific influence on the knocking combustion behavior of methane-based and liquid fuels, extensive measurements were carried out on a modern single-cylinder spark ignition (SI) engine with a high compression ratio (CR) for different types of fuel. Therefore, a fuel-specific high-precision admixing system has been designed and installed at the test bench accordingly. Additionally, a test fuel matrix has been created, including liquid fuels of a particular high knock resistance in order to be able to execute the knocking measurements at comparable engine operating points for both, the liquid fuels as well as the per se highly knock-resistant methane-based fuels. The selection of fuels included compositions of different mixture concentrations as well as components of different chemical molecular structures. For the purpose of comparing a gaseous fuel with a liquid fuel operating at the knock limit, more knock-prone mixtures were composed by admixing gaseous components with a lower methane number using methane as a reference and significantly more knock-resistant liquid fuels were selected using isooctane as a reference. As a result, a wide range of methane and octane numbers could be realized for the gaseous and liquid fuel matrix. Hence, an area of overlap regarding the MFB50 position at the knock limit was facilitated. The measure-

Abstract

XXV

ments of the knock resistance for each test fuel were carried out as an ignition time sweep (ITS), beginning at non-knocking engine operation with gradual adjustments of the ignition timing (IT) in advance until a previously defined knock rate threshold was exceeded. Figure P.2 shows results of ignition angle variation of a gaseous and a liquid fuel with similar MFB50 position at the knock limit.

10 MN = 74.8 Toluene

5

0

5

6

7 8 9 10 11 12 13 14 15 16 50 % Mass fraction burned in ° CA a. TDC [MFB50]

17

Figure P.2: Knock rate vs. 50 % mass fraction burned of a gaseous and a liquid fuel within a spark angel sweep

During the measurements, the knocking working cycles were identified by a conventional knock detection algorithm according to the VDO method, which assesses the knocking intensity by pressure trace oscillations as well as an extended knock detection algorithm, which additionally identifies the selfigniting working cycles that show no or only little-oscillations on the pressure trace. The results in chapter 5 show the fuel and parameter-specific influence on the knock rate (KR) as a function of the MFB50 position. The individu-

XXVI

Abstract

ally observed phenomena are examined by means of a working cycle-resolved qualitative and quantitative evaluation. Figure P.3 shows the heat release traces of knocking combustion for one liquid and two gaseous fuels having similar knock resistance. These initial run investigations show that the atypical knocking combustion phenomena could also be observed for liquid fuels.

250 200

Liquid fuel (Toluene) Methane based fuel (MN = 54) Methane based fuel (MN = 59)

150 100 50 0 -40

-20 0 20 40 60 Crank angle position in °CA a. TDC [ ]

80

Figure P.3: Filtered heat release rate (3 kHz-filtered) of one liquid fuel and two gaseous fuels: EOP1: n = 1500 rpm

Additionally, based on the extensive mixture variation with varying fuels regarding the methane number (MN) and research octane number (RON), an engine specific relation between the calculated fuel index numbers and actual knock resistance of the fuels by means of the motor methane number (MMN) was established and applied to the gaseous and the liquid test fuels.

Kurzfassung Bei den Untersuchungen des Klopfverhaltens methanbasierter Kraftstoffe im Vorgängerprojekt wurden bei der thermodynamischen Auswertung atypische Verbrennungsphänomene hinsichtlich der Auffälligkeiten bei der Wärmefreisetzung auf dem Heizverlauf beobachtet, die mit den heutigen Erfahrungen und Erkenntnissen über das Klopfverhalten von Ottokraftstoffen nicht erklärbar sind. Insbesondere drei ausgeprägte Wärmefreisetzungsphänomene während der klopfenden Verbrennung sind im Hinblick auf die erweiterten Untersuchungen im vorliegenden Projekt relevant: Das erste Phänomen ist ein ungewöhnlich frühes Einsetzen des Klopfbeginns innerhalb der Arbeitszyklen. Abbildung P.1 im Abstract zeigt exemplarisch den Heizverlauf eines klopfenden Arbeitsspiels mit eingezeichneter MFB50 Position unmittelbar vor dem zweiten steilen Anstieg aufgrund der klopfenden Verbrennung. Ein solch frühes Klopfen bedeutet, dass zum Zeitpunkt der Selbstzündung noch ein beträchtliches Volumen unverbrannten Kraftstoffgemisches im Endgas vorhanden ist, das nach allgemeinem Verständnis des Klopfens zu schweren Motorschäden führen sollte, die jedoch im Vorgängerprojekt nicht beobachtet wurden. Das zweite Klopfphänomen kann als die Entstehung einer ausgeprägten zweiten Wärmefreisetzungsamplitude beschrieben werden, die sich mit Beginn der klopfenden Verbrennung sprunghaft entwickelt, während die “reguläre” Verbrennung bereits abnimmt. Das dritte Verbrennungsphänomen ist das sporadische Auftreten von selbstzündungsbehaftete Arbeitszyklen ohne signifikante Schwingungsamplituden auf dem Druckverlauf. Da herkömmliche Klopferkennungsalgorithmen auf der Auswertung dieser Schwingungsamplituden basieren, werden einzelne selbstzündungsbehafteten Arbeitszyklen nicht detektiert. Aus diesem Grund wird für die Untersuchungen ein zusätzlicher Klopferkennungsalgorithmus verwendet, der die klopfend umgesetzte Kraftstoffmenge berücksichtigt und daher auch selbstzündungsbehaftete Arbeitszyklen ohne signifikante Schwingungsamplituden zuverlässig erfasst.

XXVIII

Kurzfassung

Um den kraftstoffspezifischen Einfluss auf das Kopfverhalten methanbasierter Kraftstoffe hinsichtlich der drei Phänomene weiter zu untersuchen, wurden umfangreiche Messungen an einem modernen Einzylinder-Ottomotor mit hohem Verdichtungsverhältnis für zusätzliche Gasgemische unterschiedlicher Klopffestigkeit durchgeführt. Hierfür wurde ein kraftstoffspezifisches hochpräzises Zumischsystem konzipiert und entsprechend am Prüfstand installiert um aus erdgasrelevanten einzelnen Gaskomponenten ein Gemisch zu erstellen, das bei laufendem Motor auf die gewünschte Klopffestigkeit eingestellt werden kann. Zusätzlich wurden ausgewählte flüssige Kraftstoffe im selben Aggregat untersucht um einen Vergleich mit den Gasgemischen anstellen zu können. Zu diesem Zweck wurde eine Testkraftstoffmatrix erstellt, die flüssige Kraftstoffe mit einer besonders hohen Klopffestigkeit enthält, um die Klopfuntersuchungen bei vergleichbaren Motorbetriebspunkten sowohl für die flüssigen Kraftstoffe als auch für die an sich hoch klopffesten methanbasierten Kraftstoffe durchführen zu können. Bei der Auswahl der Kraftstoffe wurden sowohl Zusammensetzungen unterschiedlicher Mischungskonzentrationen als auch Komponenten unterschiedlicher chemischer Molekularstrukturen berücksichtigt. Für den Vergleich eines gasförmigen mit einem flüssigen Kraftstoff an der Klopfgrenze wurden ausgehend von Methan klopffreudigere Gemische zusammengestellt und ausgehend von Isooktan deutlich klopffestere Flüssigkraftstoffe ausgewählt. Dadurch konnte ein größerer Methan- bzw. Oktanzahlbereich für die gasförmige und flüssige Kraftstoffmatrix abgedeckt werden sodass ein Überschneidungsbereich bezüglich der MFB50-Position an der Klopfgrenze erwartet wird. Die Bestimmung der Klopffestigkeit für jeden Testkraftstoff wurde als Zündwinkelvariation durchgeführt, beginnend aus dem nicht klopfendem Motorbetrieb mit schrittweisen Frühverstellung des Zündzeitpunkts, bis eine zuvor definierte Klopfhäufigkeitsschwelle überschritten wurde. Während der Messungen wurden die Klopfarbeitsspiele durch einen konventionellen Klopferkennungsalgorithmus nach der VDO-Methode identifiziert, der die Klopfintensität anhand der oszillierenden Schwingungen auf dem Druckverlauf bewertet, sowie durch einen erweiterten Klopferkennungsalgorithmus, der zusätzlich die selbstentzündungsbehafteten Arbeitsspiele identifiziert, die keine oder nur geringe Schwingungen auf dem Druckverlauf aufweisen. Die Ergebnisse im Kapitel 5 zeigen den kraftstoff- und parameterspezi-

Kurzfassung

XXIX

fischen Einfluss auf die Klopfhäufigkeit in Abhängigkeit von der MFB50Position. Die einzelnen beobachteten Phänomene werden anhand einer arbeitszyklenaufgelösten qualitativen und quantitativen Auswertung untersucht. Zusätzlich wurde aufgrund der umfangreichen Gemischvariationen bei verschiedenen Kraftstoffen hinsichtlich der Methanzahl und der Oktanzahl eine motorspezifische Beziehung zwischen den berechneten Kraftstoffindexzahlen und der tatsächlichen Klopffestigkeit der Kraftstoffe unter Verwendung der Motormethanzahl hergestellt und auf die gasförmigen und flüssigen Prüfkraftstoffe angewendet.

1 Introduction Constantly stricter CO2 -emission regulations and the increasing global demand for energy, require well-thought-out concepts for the future, especially for automotive applications. The automotive industry needs sustainable drive concepts that are supported by political decisions and accepted by the consumers and – at least as necessary – that also makes sense from an actual technical perspective. To ensure the internal combustion engines can meet current economic and technical requirements, steady optimization of the thermodynamic combustion process development is essential. In order to reduce fuel consumption and exhaust emissions, the trend towards higher compression ratios remains an issue. Particularly in the context of hybridization, engine map points of optimum efficiency can be preferred where knocking cycle detection and combustion control of the MFB50 position should then be even more accurately applied. For supercharged engines, the thermodynamically optimal ignition timing is hardly achieved without passing the knocking combustion limit. Knock is the auto-ignition of the unburned mixture in the end gas ahead of the propagating flame front after the actual ignition timing and during the normal combustion. Thus, when knocking combustion occurs, an adjustment of the ignition timing is necessary to avoid engine damage. Ideally, the engine should be operated as close as possible to the optimum combustion position in order to achieve a high thermodynamic efficiency, without being exposed to excessive knocking combustion. The development and application of a reliable knock detection device for high efficient monovalent vehicle engines, requires profound information about the knocking behavior of fuels. Therefore, a fundamental objective is to investigate the abnormal combustion of various fuels towards the knock limit.

© The Author(s), under exclusive license to Springer Fachmedien Wiesbaden GmbH, part of Springer Nature 2021 M. Eberbach, Knocking Combustion of Methane-Based and Highly Knock Resistant Liquid Fuels, Wissenschaftliche Reihe Fahrzeugtechnik Universität Stuttgart, https://doi.org/10.1007/978-3-658-35178-6_1

2

1 Introduction

Methane-based fuel Gasoline

-30

-15

0 15 30 45 Crank angle position in °CA a. TDC [ ]

60

Figure 1.1: Heat release rate, standardized of a gaseous and a liquid knocking working cycle

Atypical combustion phenomena were observed during the investigations of the knocking behavior of methane-based fuels. Figure 1.1 shows a representative heat release trace of a knocking working cycle for a gaseous mixture of mainly methane and the standardized (same energy content) qualitative heat release trace of a knocking working cycle measured with regular gasoline. Both knocking working cycles were measured at the knock limit of the respective engine operating point with wide open throttle (WOT). Besides the fact that it is impossible to investigate the knocking behavior of fuels with strongly varying knock resistance in a supercharged engine with high compression ratio at the same operating point, this example is only intended to explain the atypical knocking combustion phenomena by the differences in the qualitative comparison of the two heat release rates shown in the figure above on this page. 1. Early onset of knock combustion (MFB) 2. Distinct amplitude of 2nd heat release peak 3. Non-oscillating knocking working cycles

1 Introduction

3

The high knock resistance of methane-based fuels and the comparably high octane-rated liquid fuels enable further potentials, which can be exploited through efficiency-optimized MFB50 positions even at higher compression ratios and WOT. Especially in combination with supercharging, the potential to increase the indicated efficiency can be fully exploited. Natural gas, particularly methane produced of renewable energy sources, can contribute to achieve the upcoming emission targets in passenger and freight traffic, especially in the forthcoming transition period on the way to sustainable drive concepts. Knowledge about the knocking behavior of gaseous mixtures and highly knockresistant liquid fuels can provide useful information for the development of synthetic fuels and sustainable admixing agents to increase the knock resistance of conventional fuels and fulfill the requirements of upcoming engine concepts. To obtain a more detailed understanding of the root cause, comprehensive liquid and gaseous fuel mixture variations of different knock resistance and fuel-specific properties were selected and examined on the single-cylinder test bench.

2 Fundamentals of Self-Ignition Processes in SI Engines The following chapter discusses the combustion-specific mechanisms underlying the self-ignition processes in SI engines. Based on this fundamental knowledge, the experimental preparations are carried out considering the three observed atypical combustion phenomena described before. The individual sections refer to abnormal combustion, the reaction kinetic of hydrocarbons, and the pressure trace based combustion evaluation.

2.1 Spark Ignition Engines The history of the spark-ignition engine dates back more than a hundred years. Continuous development has resulted in today’s well thought-out engine concepts, which have little in common with the early days. However, the functional principle of a conventional spark-ignition engine has remained the same. Fuel, air, and residual gas are mixed, compressed in the cylinder, and under normal operating conditions, the combustion is initiated solely by a well-timed spark towards the end of the compression stroke. The spark plug supplies the activation energy by an electrical discharge to the vaporized mixture between the electrode gap. Once the flashover took place, the running pre-reactions initiate the combustion. Following inflammation beginning from the spark plug, an irregularly shaped turbulent flame front develops in a volumetric expansion and propagates through the conditioned mixture across the combustion chamber toward the walls and finally extinguishes. As a result of the combustion process, the cylinder temperature and hence pressure increase, usually reach their maximum after top dead center (a. TDC) but before the unburned mixture is completely consumed. Since the inflammation and hence the early stage of flame development depends on spark flashover, subsequent combustion cycles vary considerably. The charge motion and mix© The Author(s), under exclusive license to Springer Fachmedien Wiesbaden GmbH, part of Springer Nature 2021 M. Eberbach, Knocking Combustion of Methane-Based and Highly Knock Resistant Liquid Fuels, Wissenschaftliche Reihe Fahrzeugtechnik Universität Stuttgart, https://doi.org/10.1007/978-3-658-35178-6_2

6

2 Fundamentals of Self-Ignition Processes in SI Engines

ture homogeneity in the vicinity of the spark plug makes it even more crucial. The stochastic cycle-by-cycle variation is typical for spark-ignition engines but limits the engine operating range by the crucial cycles [52]. The burn duration is the time in which the flame develops and propagates through the cylinder volume until extinguishing near the combustion chamber wall or in front of narrow crevices and varies typically between 30 ◦ CA and 90 ◦ CA [29]. Combustion begins before top dead center (b. TDC), continues into the expansion stroke, and diminishes after the peak cylinder pressure is exceeded. If the spark timing is adjusted early, the compression work transferred between the piston and the cylinder charge increases. If the ignition timing is set late, the end of the combustion process is also shifted in the same direction, and the peak cylinder pressure decreases towards the expansion stroke. As a consequence, the expansion stroke work transfer from the cylinder charge to the piston is reduced. Between these two opposing trends, the optimum spark timing concerning the combustion efficiency is located. However, there are several factors such as engine design, fuel composition, and operating parameters that may disturb the normal combustion process and can lead to abnormal combustion phenomena discussed in the following chapter.

2.2 Abnormal Combustion Phenomena The phenomenological characteristics of an abnormal combustion are the audible metallic noise and the risk of damages in the combustion chamber. With conventional recording techniques, a reliable detection of abnormal combustion can be implemented. There are several modes of abnormal combustion which trace back to autoignition related processes. Only some of them are relevant in practice of the engine development. The focus here is set on the severely knocking working cycles, which can cause significant engine damage and those, even if not severe, which induce a significant source of noise. Abnormal combustion is characterized by extremely rapid combustion, which is not triggered by an external ignition source and releases the chemically

2.2 Abnormal Combustion Phenomena

7

bound energy in the shortest time. It can be differentiated more precisely according to its trigger: • Reaction kinetic pre-ignition • Surface ignition • Knocking combustion Reaction kinetic pre-ignition takes place before the actual ignition timing when reaction kinetic oxidation processes provoke a rapidly increasing radical concentration [52]. As soon as the activation energy is exceeded, spontaneous combustion begins immediately and the chemically bound fuel energy is converted instantaneously. A volumetric ignition process is initiated in the mixture and captures the volume in a short time without a deflagrative flame front propagation. The chemical process ends with the runoff terminating reaction, but not until the chain increasing radicals are consumed [35]. Surface ignition leads into uncontrolled combustion, which is caused by an energy transferring hot-spot in the combustion chamber. Typical spots could be an overheated valve edge, a glowing combustion chamber deposit, or an overheated spark plug electrode itself. If self-ignition occurs before the actual spark ignites the mixture, it is classified as “pre-ignition”, otherwise, it is classified as “post-ignition”. Overheated surfaces can have the unpleasant effect that although the ignition device is switched off, an unintentional ignition continues firing the engine and could lead into hazardous engine operation (runaway surface ignition). Among the three abnormal combustion mechanisms outlined above, knocking combustion is the most common, assuming that the engine is operated under conventional conditions typical for road traffic. Engine knock occurs after the regular ignition in the unburned mixture, the so-called “end gas”, triggered by a spontaneous self-ignition. If the flame front has already consumed the end gas before the reaction kinetic processes have induced self-ignition, of course, no abnormal combustion can occur anymore. The focus of the here implemented investigations is set on the atypical knocking combustion phenomena observed for methane-based fuels.

8

2 Fundamentals of Self-Ignition Processes in SI Engines

The extensive experimental and numerical investigations of the last decades have resulted in a comprehensive knowledge of the processes associated with knocking combustion, even though the exact physical and chemical processes have not yet been fully clarified, there is a commonly accepted assumption that knocking combustion is preceded by auto-ignition of exothermic centers in the unburnt mixture [50]. The development of knocking combustion is determined by the initiated chemical chain reactions and secondary flame front as well as the predominant condition in the local end gas pocket and its volume. The process is influenced on the one hand by local mixture inhomogeneities and on the other hand by unequal temperature distribution. The self-ignition exothermic centers are mainly located near the combustion chamber wall, as shown in Figure 2.1 and are not directly affected by the flame due to their spatial separation. Compression pressure and heat input by the spark-induced regular flame front can provoke self-ignition[38] [76] [60].

 

 

Figure 2.1: Shadowgraph photographs of knocking combustion: 33 μs between frames(left). Photodigitized picture (right) showing additional details of the self-ignition location [49]

Consequently, spontaneous combustion can occur in several self-ignition centers within a short period and can lead to a spontaneous and fast release of chemical energy [72]. Even if the mean unburned mixture temperatures are lower than those required for an ideal homogeneous thermal explosion, it can lead to severe knocking combustion. In general, a distinction in the flame expansion is made between a deflagration and a detonation. The diffusion and heat conduction processes determine

2.2 Abnormal Combustion Phenomena

9

the propagation of a deflagrative flame front. The flame propagation speed is composed of a laminar and a turbulent component. In contrast, a detonation is associated with a shock wave that can be caused by the heat released during the chemical reaction. The chemical reaction itself is ignited by the high temperature and high pressure at the edge of the shock wave and is several times faster than the speed of sound. Usually, the prevalent mixture compositions enable both kinds of flame propagation, deflagration and detonation, whereby a transition from deflagration to detonation is also possible [8][75]. In a sizeable end gas pocket, the energy conversion is faster and, depending on the distance, a detonation wave can develop. The initiated reaction kinetic in small pockets is lower and spreads to the surrounding at moderate speed. The pressure wave excitation in the combustion chamber and the risk of severe component damage are comparatively low. A pre-differentiation of the three flame propagation forms can be made based on the local temperature gradient between the compression igniting source (hot spot) and its surrounding unburned mixture [77]. Deflagration This type of combustion is characterized by a relatively low mean end gas temperature and a steep temperature gradient (order of magnitude 100 K mm−1 [37]), which leads to light knock and moderate flame propagation speed. Deflagration is the flame propagation, which is similar to the flame velocity of a regular combustion, caused by chemical reaction and molecular or turbulent transport processes into the end gas. The heat transfer processes determine the propagation speed of the reaction zone and thus the flame speed and its energy conversion rate. The energy conversion reaction rate is within the same magnitude as for regular combustion. The effect on the heat release rate is moderate and depends on the number of exothermic centers. Without a distinct pressure gradient, only slight pressure waves originate, and thus, the oscillation on the pressure trace signal is not particularly pronounced. The propagation velocity of a deflagration is relatively small [33].

10

2 Fundamentals of Self-Ignition Processes in SI Engines

Thermal Explosion (Homogeneous Self-Ignition) The thermal explosion shows relatively high mean end gas temperature and a small temperature gradient (order of magnitude 1 K mm−1 [37]) resulting in an abrupt energy conversion of the entire end gas due to the almost isothermal end gas zone. This type of knocking combustion is comparable to volumetric combustion triggered by a self-ignition source. The flame propagation speeds are extremely fast for this case and medium pressure oscillations are formed instantaneously. Such a uniform temperature distribution is improbable in a gasoline engine like combustion processes. Developing Detonation Detonation arises from exothermic centers of intermediate size with mean end gas temperatures and mean temperature gradient (order of magnitude 10 K mm−1 [37]). The combination promotes the development of strong pressure waves resulting in severe knock. Several self-ignition centers may arise as moderate knock and interact into developing detonation. The transition into developing detonation requires the pressure gradients merging into a shock wave, if sufficient time and unburned mixture is available. The shock wave runs ahead of the reaction zone, whereby the propagation speed is determined by the propagation speed of the shock wave. Therefore, the actual state of a not completely distinctive detonation wave is also described as a developing detonation. The developing detonation is the recurrent regime to cause knock damage in spark-ignition engines [37]. It is commonly assumed that knocking combustion appears in several spots in the unburned mixture almost simultaneously, see the two spots, highlighted by arrows in figure 2.1. Which of the three above described regimes predominates can hardly be distinguished absolutely, as they usually merge seamlessly [77]. In a conventional gasoline engine with port fuel injection (PFI) assuming a homogeneous mixture quality, temperature is a predominant factor influencing the chemical reaction at a developing detonation, as discussed in section 2.3.

2.2 Abnormal Combustion Phenomena

11

As a consequence, the temperature-time history determines the location where self-ignition starts within an end gas pocket. [72] [16]. When abnormal combustion occurs, the end gas burns uncontrollably. In addition to the regular flame front, a secondary reaction front forms, originating from the exothermic center in the end gas where self-ignition occurs. Depending on the type of flame propagation, velocities of up to 2000 m s−1 are possible, considering that a deflagratively propagating turbulent flame front hardly exceeds a propagation velocity above 20 m s−1 for conventional turbulence levels. [36] [59]. There is a tendency of high-pressure deflection as a function of the propagation speed on the pressure trace. As a consequence, high local pressure gradients occur, and the arising pressure waves propagate across the cylinder. The high-frequency pressure oscillation on the cylinder pressure trace is a result of the waves impinging against the combustion chamber walls as well as the deflagratively moving regular flame front. Reflected waves cause interferences and superposition effects. Standing waves can develop in the combustion chamber. This oscillating system is characterized by the thermodynamic propositions of the gas state, and the geometrical shape of the cylinder head and piston surface [61]. Despite the piston movement, it can be assumed that a steady-state resonance exists for a sufficiently short period of time in the combustion chamber, although the piston movement is continuous, but comparatively slow near the top dead center. The hollow and metallicsounding noise, which is transmitted through the engine structure when a portion of the end-gas ignites spontaneously, is known as “knock”. Figure 2.2 shows the pressure trace of three individual working cycles (IWC) with different knock intensities at the same engine operating point. For low knock intensities (KI), the onset of knock is usually later in the working cycle, and the amplitude of the pressure fluctuations is less distinctive. Severe knock usually occurs closer to top-center and thereby relatively early in the combustion process when significant percentages of the mixture are still unburned.

12

2 Fundamentals of Self-Ignition Processes in SI Engines

180 Severe oscillation Light oscillation Without oscillation

160 140 120 100 80 60 40 20 0 -60

-40

-20 0 20 40 60 Crank angle position in °CA a. TDC [ ]

80

Figure 2.2: Unfiltered cylinder pressure traces of individual engine cycles with and lacking the typical oscillations

The high-frequency pressure fluctuation is superimposing the uniform pressure trace of the regular propagating flame front immediately with the onset of knock and decays continuously during the further combustion cycle. The oscillation on the pressure trace is the result of the pressure wave arriving at the piezoelectric transducer combined with the structure-related vibration of the components. A further distinction can be made between the primary and secondary knocking effect. The primary knocking effect describes the self-ignition in the end gas itself, the secondary knocking effect describes the subsequently developing typical but not necessarily always occurring pressure oscillations in the cylinder [3] which in turn are divided into low and substantial fluctuation on the pressure trace as they were observed in these investigations. The results can

2.3 Reaction Kinetics of Irregular Combustion

13

be found in the chapter 5.2.4 on the evaluation of the third atypical observed knocking combustion phenomena. In Figure 2.3 on page 14, a knocking and non-knocking propagation flame front in a sequence of Schlieren photographs is compared. Each of these images is referenced to the crank angle (CA) position and can be assigned to the exact time in the recorded pressure trace of the working cycle. In the upper set of photographs, the turbulent flame front moves steadily through the end-gas while the normal combustion process is completed. In the bottom sequence of photographs (b), the initial flame front (photographs 1 to 3) behaves comparable to the normal combustion, and suddenly the entire end gas region ahead is captured. The cylinder pressure rises rapidly and oscillations at 6 to 8 kHz were detected [29]. There are extensive investigations on abnormal combustion in literature available, some individual publications concerning knocking give a good overview and also a deeper insight into the topic [71] [13] [24] [72] [74].

2.3 Reaction Kinetics of Irregular Combustion The self-ignition of a gaseous hydrocarbon-air mixture takes place when the energy released is larger than the heat lost to the surroundings. The self-ignition is a series of primary chain sequences developing into a radical chain explosion. The Ignition processes can be divided into the thermal explosion and the chemical chain explosion. In contrast to the latter, the thermal explosion takes place if the acceleration of the reaction is not produced by radical propagation, but in an exothermic reaction where the heat generation exceeds the heat dissipation to the surrounding. A characteristic of the thermal explosion is the immediate rise in temperature, whereas the chain explosion starts with the SI engine characteristic ignition delay before the explosion, compare to figure 2.4 on page 15. The ignition delay τ is given by the equation eq. 2.1.

14

2 Fundamentals of Self-Ignition Processes in SI Engines

Figure 2.3: Schlieren photographs of (a) normal flame propagation through the end gas and (b) knocking combustion (autoignition occurs in photograph 4), with corresponding cylinder pressure [29]

2.3 Reaction Kinetics of Irregular Combustion

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Figure 2.4: Simplified temporal progression of the temperature during thermal and chain-branching explosion in an adiabatic system [72]

During the ignition delay, the radical concentration is increasing at an exponential rate. The degree of energy liberated is small and can hardly be detected. The chemical reactions (compare equation eq. 2.4) take place during the induction time, whereas the temperature remains nearly constant. As soon as the amount of radicals becomes large enough to consume a significant fraction of the fuel and immediate ignition will take place. Due to the temperature dependence of the underlying partial reactions, the ignitions delay time depends strongly on the temperature. This is shown in figure 2.5 for several hydrocarbon-air mixtures. The ignition delay τ is given by the equation eq. 2.1 and describes the ignitiondelay time considering the fuel-specific parameters. It can be seen that it depends exponentially on the reciprocal temperature. τ = A · p−n · e(B/T ) with τ p T A, n, B

ignition delay current pressure current temperature fitted parameters depending on the fuel

eq. 2.1

16

2 Fundamentals of Self-Ignition Processes in SI Engines

Figure 2.5: Calculated (line) and measured (points) ignition-delay times in hydrocarbon mixtures [72]

At the beginning of a radical reaction is the chain initiation, which is necessary to initiate a combustion reaction and, therefore, usually endothermic. The chain propagation mechanism sustains the running reaction. The energy released during these first reactions is used either to initiate further chain initiation reactions or for chain propagation. A branching of the chain reaction occurs when the number of free radicals increases. However, the start reactions as well as the chain propagation and chain branching are overlaid by reactions that inhibit the sequences, the terminating reactions. As long as the free-radicals reaction absorbs the released reaction energy, the temperature of the system does not rise noticeably. Only after a specific ignition delay time, sufficient exothermic reactions take place, which lead to an increase in temperature and a subsequent explosive energy conversion, this happens when more chain branching and chain propagating reactions are running

2.3 Reaction Kinetics of Irregular Combustion

17

than chain-terminating. The rate and ratio of chain initiation, chain propagation, chain branching, and chain termination reactions are strongly dependent on temperature and pressure [72]. Exemplary partial reactions between methane and elementary oxygen are described in eq. 2.2 to eq. 2.5 [46]. Chain initiation

Chain propagation

CH4 + O2 → CH3 • + HO•2

eq. 2.2

CH4 + OH • → CH3• + H2 O

eq. 2.3

CH4 + O• → CH3• + OH •

eq. 2.4

CH3• +CH3• → C2 H6

eq. 2.5

Degenerate branching

Chain termination

The perfect exothermic oxidation of a hydrocarbon compound CxHy can be described in general by the gross reaction equation eq. 2.6 y y Cx Hy + (x + )O2 ⇒ x ·CO2 + H2 O + ΔHR 4 2

eq. 2.6

Figure 2.6 shows schematically different explosion limits of hydrocarbons as a function of pressure and temperature. The region of the explosive combustion is separated by the curved line. On the left, the delimiting line in the diagram is the regime of slow reactions where the pressure and temperature combination does not enable a sufficient concentration of radicals to ignite the mixture and enable an explosion. When the pressure is increased above a certain value (first explosion limit), spontaneous ignition is observed. Due to the higher pressure, the diffusion rate of the radicals to the wall where they are destroyed is reduced, since the diffusivity in gases is inversely proportional to the density. At the same time, the relative concentration of radicals in the gas phase increases (chain termination eq. 2.5 at the surface and chain branching eq. 2.4 in the gas

18

2 Fundamentals of Self-Ignition Processes in SI Engines

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Figure 2.6: Explosion diagram for hydrocarbons (schematically) [72] phase). Due to the higher combustion chamber pressure in engines at knockrelevant operating points, the first and second explosion limits are of minor importance for autoignition in the endgas. The third explosion limit is the thermal ignition limit. Its position in the diagram results from the competing processes of heat production by chemical reactions and wall heat losses. With increasing pressure the heat production increases at constant volume, such that at sufficiently high pressures the transition to explosion occurs. The explosion limits are determined by highly non-linear processes. The explosion limits of hydrocarbon-air mixtures are much more complex due to additional chemical processes, e.g. the formation of peroxide, especially around the third explosion limit. For temperatures T > 1200 K, the single-stage high-temperature ignition is predominant. For a certain pressure and temperature range (T > 900 K) relevant to the end gas, the area of the so-called cool flames is passed through (low-temperature ignition). Frequently, a negative temperature coefficient in the explosion curve is also found in the temperature range of these cool flames.

2.3 Reaction Kinetics of Irregular Combustion

19

Thus, in these pressure areas, as the temperature increases, the conversion of the reactive species into inert products may outweigh the chain formation and chain propagation reactions resulting from the increase in pressure. For more detailed information, see [72]. B ººººººº.4# RANGE

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Figure 2.7: Qualitative progressions: a) two-stage ignition, b) temperature dependence of the ignition delay for a n-heptane-air-mixture [63]

Usually the end gas heat losses in vehicle combustion engines are higher than in the research engine from which the results above are derived. Thus, autoignition occurs at comparatively lower temperatures. As long as the H2 O2 decomposition is quite slow at these temperatures, other chain branching mechanisms govern the ignition process which explains the observation of the so-called twostage ignition and a negative temperature coefficient (NTC) of the ignitiondelay time. The two-stage ignition and the NTC is shown in figure 2.7. In the low-temperature range between 800 K to 900 K, chain branching initially takes place, which leads to a slight temperature increase. At higher temperatures, the precursors formed in the chain branching mechanisms decompose back into their original substances (degenerated chain branching) due to the oxygen addition, before a second ignition takes place after a subsequent induction period. This is characterized by the almost fuel-independent reactions of the high-temperature oxidation. The range of a NTC is characterized by the fact that an increase in temperature results in a slow-down of the ignition and thus an extension of the ignition-delay time, see figure 2.7 b.

20

2 Fundamentals of Self-Ignition Processes in SI Engines

According to Arrhenius (1889), the correlation between temperature and the chemical reaction rate can be described using an exponential approach and an additional pressure factor, see equation eq. 2.7. The mathematical relation is applicable to all chemical elementary reactions. The coefficients k describes the reaction rate and is strongly dependent on the temperature: Ea k = A · e(− RT )

eq. 2.7

Ea represents the activation energy, R is the universal gas constant and A the pre-exponential factor which is comparatively low dependent on the temperature T. Similar to the Arrhenius approach for the reaction rate of the thermal explosion, an exponential approach also exists for the rate of radical formation [52]. The temperature dependence of the ignition delay can be determined experimentally based on stationary ignition-delay investigations, proceeding from the elementary reaction up to complex reaction schemes of different hydrocarbon compounds. Figure 2.8 shows measured ignition-delay times versus temperature variation for different hydrocarbon-air mixtures [61]. The logarithmically plotted ordinate shows that ignition delays above 900 K depends almost exponentially on the reciprocal of the temperature. This relationship reflects the temperature dependence (Arrenhius’ law) of the underlying elementary reactions. However, considering the low-temperature oxidation range at temperatures below 900 K regarding possible NTC gradients, the typical temperature dependence of the ignition delay no longer follows the s-shaped curve within a specified temperature interval, see 2.8 for n-heptane. The differences in the ignition-delay behavior of various fuels in the enginerelevant range of temperature above 1000 K indicate that the low-temperature kinetics of the transition temperature range is responsible for the differences in knocking behavior and octane rating of the fuels investigated [20].

2.3 Reaction Kinetics of Irregular Combustion

21

Figure 2.8: Comparison of ignition delays for different stoichiometric air / fuel mixtures at p = 13 bar [20] Figure 2.9 shows the comparison of measured and calculated ignition-delay times of n-heptane / air mixtures for different pressures. In the low temperature range below 700 K the curves show a low pressure dependence. At higher temperatures, the negative temperature gradient shows a pressure-dependent trend of the individual curves, which show a narrower NTC range as the pressure increases and are additionally shifted towards higher temperatures. In the high temperature range, the pressure dependence remains [20]. The oxidation of hydrocarbon-air mixtures in an autoignition event can exhibit different types of behavior, or a sequential combination of them, depending on the pressure and temperature of the mixture [29]: • slow reaction or multiple cool flames (slightly exothermic reactions) • two-stage ignition (a cool flame followed by a hot flame) • a single-stage ignition (hot flame)

22

2 Fundamentals of Self-Ignition Processes in SI Engines

Figure 2.9: Comparison of ignition delays for different stoichoimetric air / fuel mixtures at p = 13 bar [20] At temperatures between 570 K to 670 K, single damped pressure waves can occur, whose accompanying phenomenon, the faint blue light emission, can be detected. Only a small amount of reactants are reacting and the temperature increases only a few tens of degrees. These so-called cool flames depend on the conditions in the combustion chamber and the fuel and may well trigger a hot flame or thermal explosion, where the reaction accelerates rapidly after ignition. This sequence is called tow-stage ignition. If the mixture temperature continues to rise, a transition to single-stage ignition takes place. While all hydrocarbons have ignition delay times followed by a very fast reaction rate, some hydrocarbon compounds do not exhibit the behavior of a cool flame or two-stage ignition [29].

2.3 Reaction Kinetics of Irregular Combustion

23

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'

 
' ' Figure 2.10: Ignition diagram. Two-stage ignition occurs in the lowtemperature region; first stage may be a cool flame. Single-stage ignition occurs in the high-temperature region [29]

Figure 2.10 shows these ignition limits for three different fuels. For isooctane, a two-stage ignition process in low-temperature regions is observed. There is a first ignition-delay time interval before the cool flame appears and then a second time interval till the hot flame combustion process begins. Ignition in the high-temperature region is by a continuous one-stage process. The cool flame phenomenon depends strongly on the hydrocarbon structure. Normal paraffins show pronounced cool flames, branched-chain paraffins are more resistant. Olefins give even lower luminosity cool flames with longer ignitiondelay periods. Methane and benzene show only a high-temperature ignition limit without the cool flame phenomena. Aromatics hardly emit detectable luminosity. This indicates, that some compounds knock by a low-temperature

24

2 Fundamentals of Self-Ignition Processes in SI Engines

tow-stage ignition mechanism, some via a high-temperature single-stage ignition mechanism, and for some fuels both mechanisms may play a role. For stoichiometric hydrocarbon-air mixtures, no knock relevant reaction occurs below 470 K unless the mixture is ignited by an external source such as a spark [29]. In the case of real knocking phenomena, it is not always possible to clearly distinguish the mechanisms and characteristics of the knock. Self-ignition generally takes place at several locations in the combustion chamber, and the energy conversion mechanisms (deflagration, detonation and explosion) do not always occur separately, but can rather merge into one another. In contrast to the spark induced ignition process, self-ignition can also occur in the air-fuel mixture if a sufficient amount of energy is transferred to the unburned mixture by the compression processes. In the case of spark ignited gasoline engine combustion, self-ignition processes are undesirable. Self-ignition before the actual ignition timing in the compression phase is referred to as pre-ignition. However, the much more frequent self-ignition occurs after the conventional ignition in the end gas. Compression and heat transfer by the spark induced flame can cause critical pre-reactions in the unburned mixture, which finally lead to selfignition. Self-ignition occurs primarily at so-called exothermic centers in the end gas. If present, these can be highly ignitable oil droplets or hot particles, but are more often the result of temperature and mixture inhomogeneities.

2.4 Cylinder Pressure-Based Knock Detection In a production vehicle, knocking combustion is detected by evaluating the indirect knocking effect using a structure-related sound sensor that detects the vibrations transmitted by the engine components, which, in turn, are stimulated by the acoustic pressure- and shock waves. If necessary, the knock control system can temporarily defuse the situation by retarding the ignition angle into the non-knocking range. This is followed by an incremental approach towards the efficiency-optimal MFB50 position until the knock limit is approached again. Before a reliable control system can be implemented in passenger cars, com-

2.4 Cylinder Pressure-Based Knock Detection

25

prehensive application work had to be done in a preliminary stage on the test bench. The pressure in a combustion chamber is recorded and processed by a so-called indicating system. The evaluation is of particular importance for the assessment of the ongoing combustion process and, besides optical access, the most reliable method to evaluate knock [62]. Especially for the knock investigations, a water-cooled piezoelectric transducer is capable of enduring extreme loads over a short period of time. The piezoelectric effect is a unique feature of specific crystals that react to an externally applied load with a proportional charge shift in the crystal stack. Piezoelectric pressure transducers detect relative pressures; thus it is necessary to reference it to absolute pressure between the combustion working cycle, i.g. by referencing it to the inlet air pressure. The indicating measuring system references the current cylinder pressure versus the corresponding crankshaft position. A charge amplifier processes the small charge quantities emitted by the transducer and passes them on as a proportional voltage signal. Based on the measured pressure signal, further combustion parameters are calculated, which can be divided into the so-called “direct” and “indirect” indicating parameters. Directly indicated parameters such as maximum cylinder pressure and its maximum gradient are derived from the cylinder pressure trace and enable quick rating of the component mechanical stress. The evaluation of the indirectly indicated parameters must be preceded by intermediate calculations such as the heat release trace. With the calculation of the heat release rate, statements about ignition delay, the combustion position relative to the top dead center (TDC), and the onset of knock can be made. The heat release trace is a fast approximation. Simplifications are used for its calculation, which allows a real-time calculation of the specific heat release trace without time-consuming calculations of the wall heat losses, which are required for the calculation of the burn rate. The derivation of the calculation formula for the differential heat release rate starts with the transformation of the first law of thermodynamics. dV dHl dQb dU dQw = − + p· − dϕ dϕ dϕ dϕ dϕ

eq. 2.8

26

2 Fundamentals of Self-Ignition Processes in SI Engines

For the derivation, the crevice leakage rate dH l is neglected. The differential heat release rate dQh is defined as the sum of the (differential) burn rate dQb and dQw which is the convective heat-transfer rate to the combustion chamber walls. dQh dQb dQw dU dV dHl = + = + p· − dϕ dϕ dϕ dϕ dϕ dϕ 

eq. 2.9

≈0

With the simplified calculation of the change in internal energy assuming the cylinder mass mcyl to be constant and the isochoric heat capacity cv corresponding to dU dT = mcyl · cv · dϕ dϕ

eq. 2.10

follows for the differential heat release rate dV dT dQh = mcyl · cv · + p· dϕ dϕ dϕ

eq. 2.11

differentiating the ideal gas law p·

dmcyl dp dT dV +V · = mcyl · R · +R·T · dϕ dϕ dϕ dϕ

eq. 2.12

follows again neglecting the change of state of the cylinder mass dp p · dV dT dϕ +V · dϕ = mcyl · dϕ R

eq. 2.13

inserted in eq. 2.11, taking into account the relationship with the polytropic exponents κ 1 cv = R κ −1

eq. 2.14

2.4 Cylinder Pressure-Based Knock Detection

27

after transition from differentials to finite differences finally follows ΔQh = Δϕ

κ κ−1

1 · p · ΔV + κ−1 ·V · Δp Δϕ

eq. 2.15

Using eq. 2.15, the heat release rate can finally be calculated based on the pressure trace and crank angle signal as well as the volume function with the crank mechanism without further model assumptions. [6] The knock evaluating criteria can be based on either the combustion pressure trace itself or the calculated heat release rate [72] [61]. Pressure Trace Related Parameters The knocking combustion can be detected by means of the pressure trace signal. In the case of regular combustion, the signal shows an even trace according to the pressure increase or decrease in the combustion chamber. In the case of knocking combustion, additional energy is converted rapidly from the location of spontaneous self-ignition. This results in additional pressure waves, which show up as oscillations on the pressure curve. These high-frequency pressure fluctuations superimpose the “normal” pressure trace of the regular flame front in a working cycle. The oscillating signal can be extracted by a frequency filter, as shown in figure 2.11. The so-called band-pass filters only allow frequencies within a specific range to pass [15]. The upper and lower cut-off frequency is selected according to the frequencies occurring during knocking combustion. Pressure transducer and engine combustion natural frequencies can be analyzed and excluded from the knock detection analyses [70].

28

2 Fundamentals of Self-Ignition Processes in SI Engines

Figure 2.11: Cylinder pressure trace (left) and according oscillations of the band-pass filtered pressure trace [16]

This evaluation method additionally enables to assess the knock intensity based on the band-pass filtered pressure trace. The absolute value of the peak amplitude, can be referred to the knock intensity [16] [41] [53]. Alternative approaches have been developed to evaluate the knock intensity. The “knock intensity factor” as described with equation eq. 2.16 evaluates the oscillation associated with knock by analyzing for a finite number of pressure oscillation amplitudes. More precisely, for the arithmetic mean of the squared difference between pressure amplitude pi and mean pressure pmean in the defined crank angle window of 20 ◦ CA following form the onset of knock [38]. nsamp

KI20 =

1

∑ (p(i) − pmean )2 nsamp

eq. 2.16

i=1

There are also different approaches to the definition of the onset of knock combustion. However, most of them use the first abruptly rising pressure pulse in the knock-relevant crank angle range, which deviates significantly from the expected pressure profile due to regular combustion, as the beginning of knock. The possibility of cutting off specific frequencies and the natural frequency of the recorded pressure trace extract the primary information of the simplified signal to apply knock evaluation. Non-knocking working cycles have by defi-

2.4 Cylinder Pressure-Based Knock Detection

29

nition a knock intensity of zero [12] [53]. The shape of the band-pass-filtered pressure trace is influenced by the combustion chamber design, the natural frequencies of the components, standing waves, and the signal sampling rate as well as the position of the transducer [11]. By using an incorrect band-pass filter range or crank angle window splitting, there is a risk of losing information from knock-relevant frequencies. As well as with complex combustion chamber shapes and / or inappropriately chosen sensor positions, there is also the risk that the pressure waves will be insufficiently detected. [12] [38]. Another method evaluates the first derivation of the pressure trace, which shows the cylinder pressure increase within 1 ◦ CA. Evaluation by means of a bandpass filter within a selected crank angle range makes it possible to evaluate the intensity and identify the beginning of knock. The knock intensity is referred to as a predefined threshold value [69]. This knock evaluation criterion offers the advantage to compare the values of different engine types and fuels, but large pressure gradients are not necessarily the characteristic of knocking working cycles. This disadvantage can be partially compensated by a wellthought-out definition of the applied frequency filter. However, it should also be noted that the exact ranges of knocking frequencies are engine-specific, and therefore one band-pass filter is not generally transferable to other applications. Generally, the intention should be to capture all necessary frequencies with knock-specific information, but without including too many low-frequency ranges, particular engine resonance frequencies. Usually, the natural frequency of the pressure transducer itself is higher than the knocking frequencies and, therefore, not included in the relevant range of knock evaluation [12] [10]. Another evaluation method is based on the energy content of the cylinder pressure signal. The power spectral density (PSD) of the cylinder pressure trace assesses the knock intensity. The evaluation corresponds to the integrated Fourier spectrum of the knocking pressure trace fluctuation applied for a certain frequency range. Knock specific frequency ranges used in publication are e.g. 0-20 kHz [25] or 5-12 kHz [40]. Discrete Fourier transform or a wavelet transform is suitable for frequency analyses. For the evaluation of the Fourier spectrum the natural frequency of the pressure transducer and the combustion chamber natural frequencies are to be excluded. The onset of knock is defined as

30

2 Fundamentals of Self-Ignition Processes in SI Engines

the first noticeable increase of the amplitude and thus of the frequency-specific knocking frequency. It should be noted, however, that the exact ranges of typical knocking frequencies are engine specific and thus can only be applied to a limited extent. When selecting the band-pass filter, all knock-relevant frequencies should be recorded, but without including too low frequency ranges or component natural frequencies [11] [53]. Heat Rate-Related Parameters Like the pressure trace related parameters for knock detection, the heat release trace in differential form and band pass filtered can be used as a basis for knock detection. These methods evaluate the increase resulting from the self-ignition or the following rapid decrease of the heat release rate in the combustion chamber. Figure 2.12 shows three heat release rate traces with various severe intensity of knocking combustion. 350

Strong knock Light knock No knock

300 250 200 150 100 50 0 -50 -30

-20

-10 0 10 20 30 40 50 Crank angle position in °CA a. TDC [ ]

60

Figure 2.12: HRR traces of a binary gaseous mixture, non-, slightly-, strongly knocking combustion (3 kHz-filtered)

2.5 Influencing Parameters of Knocking Combustion

31

The stronger the knock intensity, the more distinctive the second heat release peak [55]. The onset of knock is thereby identified by exceeding a predefined threshold value starting from the position where the maximum amplitude of the heat release oscillations occurs [75].

2.5 Influencing Parameters of Knocking Combustion Glucksstein et al. [23] have already investigated in the middle of the last century the relationships of knock as a function of the end gas temperaturepressure history for primary reference fuel (PRF) of varying RON and have shown the exponential increase of the reaction kinetic processes in the end gas. The influence of the fuel properties, the engine boundary conditions and the compression ratio on the knocking characteristics should be taken into account. Besides the knock performance of the fuel, there are several engine related factors affecting the self-ignition processes. Derived from the fundamental physical processes mentioned in chapter 2.2, it can be summarized that for the initiation of auto-ignition in the end gas, a certain pressure and temperature level must be typically exceeded. Additionally, there has to be sufficient time for the temperature and pressure level to initialize the pre-reactions in end gas and trigger self-ignition. Consequently, a further classification in engine operating and engine design parameters can be made. Ignition Timing The modification of the ignition timing has an immediate effect on the MFB50 position and is thus particularly useful for knock control. Adjusting the ignition timing is a useful approach to shift the combustion gradually toward the knock limit. The cylinder pressure trace composes of the compression pressure and the combustion pressure. The former would be the pressure trace as it would appear in motored engine operation, i.e. without combustion. If the ignition timing is adjusted towards early, the main combustion shifts towards the TDC and thus the MFB50 position. Early combustion positions lead to significantly higher peak pressures and increases the temperature level in the

32

2 Fundamentals of Self-Ignition Processes in SI Engines

end gas, which in turn promotes the probability of knock and its intensity. At the same time, the onset of knock is shifted to early.[16][28][25]. Shifting the spark timing to early reduces the burn duration and thereby counteracts the tendency to knock. The faster combustion process close to the TDC increases the indicated mean effective pressure (IMEP) and the thermodynamic efficiency. There is a parabolic relationship between the MFB50 position and the thermodynamic efficiency. The parabolic peak point is dependent on the ignition-delay time, at roughly 8 ◦ CA a.TDC ± 2 ◦ CA. The parabolic relationship results due to the increasing wall heat losses for advanced ignition timing and the increase in exhaust gas heat losses towards late [3]. While an efficiency-optimized ignition timing can be set at partial load (due to the lower pressure and temperature level), the increased tendency to knock at WOT usually requires an adjustment of the ignition timing. The experimental results in section 5.1 show the influence of the ignition timing on the fuel’s knock tendency in more detail. The mean MFB50 position within the same engine operating point was shifted approximately by 4 ◦ CA ± 1 ◦ CA from the first detected knocking working cycle until a knock rate of 5 % was exceeded. For the knock investigations here, a knock limit of 5 % knocking work cycles was defined. Engine Design The knock tendency depends on the combustion chamber design. The compression ratio is of particular interest as increasing the compression ratio enhances the thermodynamic efficiency. Thereby the motored pressure trace is shifted to higher compression end pressures. The geometrical shape of the reduced cylinder dead volume has more distinctive squish crevices. Higher end gas temperatures and changed charge motion influence the knock probability. [61]. In contrast to the ignition timing, the combustion chamber design cannot be easily adopted in engine operation. Hence, the knock rate within one measuring point was increased by adjusting the ignition timing. Figure 3.3 shows the

2.5 Influencing Parameters of Knocking Combustion

33

combustion chamber shape, the pent-roof of the cylinder head and the piston surface. The design of the intake manifold influences the cylinder charge motion during the gas exchange. As the piston moves upwards in the compression stroke, the charge motion decompose into turbulence. The thereby increased flame surface structure accelerates the energy conversion process as the burn duration decreases. At the end of the stroke, the squish crevices can either lead to a faster burnout by additional generated turbulence or the flame can not completely consume the end gas in the tight crevices [68][63]. On the one hand, a shortening of the burning time reduces the time available for self-ignition. On the other hand, temperature and pressure increases in the end gas accelerate the pre-reactions which may trigger self-ignition. The position of the spark plug defines the maximum flame propagation distance to the cylinder wall and thus, the time during which the end gas could initiate autoignition before the regular flame front consumes it. Moreover, it is favorable to start the ignition of the cylinder charge in the vicinity of high component temperatures – usually the exhaust valves, so that the unburnt mixture has less time to absorb heat and thereby increase the probability of selfignition. With increasing engine speed, the absolute time available for combustion is reduced (disproportionately). This means that there is less time for possible self-ignition in the end gas. On the other hand, the pressure and temperature level increases again because, firstly, the energy is converted faster per cycle and, secondly, the number of working cycles per time increases following the engine speed. The waste heat is dissipated more slowly and the charge exchange efficiency decreases, so that critical temperature levels in the end gas can be achieved earlier and self-ignition is more likely to occur. The influence of the air / fuel ratio on the burning velocity depends on the chemical factors of the fuels. For conventional hydrocarbon fuel, the highest laminar burning velocity is close to the stoichiometric ratio approximating toward the peak point from the rich area side (≤ λ ), see figure 2.13. In consequence, the knock tendency decreases for under- and over stoichiometric mixtures [21]. The higher the evaporation enthalpy of the fuel, the stronger the cooling effect on the mixture temperature. The methane-based fuels are already present

34

2 Fundamentals of Self-Ignition Processes in SI Engines

in the gaseous state when they are injected, so that the heat absorbed by the mixture during homogenization is relatively low.

Figure 2.13: Laminar burning velocity for different fuels as function of equivalence ratio; ambient pressure; 300 K. [29] If the inlet temperature is increased, the final compression temperature rises, which favors the chemical pre-reactions and accelerates the inflammation of the mixture, as a result, the exhaust gas temperature increases. Hence, the tendency to knock increases, although the captured cylinder charge becomes smaller as the former effects outweigh the reduced filling [25] [66].

2.6 Knock Resistance of Fuels

35

If the proportion of the internal residual gas increases at a constant cylinder mass, the mixture temperature is higher at the time of inlet valves close. The knock promoting effect is comparable to the above-discussed change of the inlet temperature, whereas the dilution and decelerating laminar burning velocity counteract this effect.

2.6 Knock Resistance of Fuels Besides the engine specific design parameters, the knock tendency is influenced by the fuel-related properties or the knock resistance respectively. In common gasoline-like fuels, the hydrocarbon compounds vary enormously in their molecular size and structure. The molecular structure is the geometric, spatial relative arrangement of the individual H and C atoms within a molecule, which substantially influences its anti-knock behavior. Figure 2.14 and 2.15 give an overview how the knock tendency is related to the fuels molecular structure and number of atoms. The individual hydrocarbon compounds are divided into different chemical families. The engine fuels are mainly composed of the following series and classes [29]: • Alkanes (paraffins) • Cyclanes (napthenes) • Alkenes (Olefins) • Aromatics Figure 2.14 shows the critical compression ratio of different chemical families plotted versus their number of carbon atoms. It shows on one hand the tendency that the higher the number of carbon atoms, the higher the sensitivity of the fuel to knock. Aromatics are classified at a compression ratio above 14 although it binds relatively many C atoms in the molecule. Their molecular structure consists of the firm ring formation with double bonds, which requires high activation energy to crack and therefore makes it relatively knock resistant. However, it is disadvantageous that its harmful effect on health can not be completely excluded.

36

2 Fundamentals of Self-Ignition Processes in SI Engines

Depending on the individual hydrocarbon compounds, the NTC behavior is different, compare figure 2.8. Several olefins and aromatics show no NTC, in contrast to paraffins [42]. The maximum autoignition heat release rates for aromatics are lower than those for the paraffinic fuel [9].

1

Figure 2.14: Critical compression ratio versus the number of carbon atoms in a molecule for different hydrocarbons [29]

2.6 Knock Resistance of Fuels

37

Figure 2.15 additionally shows the RON dependence on the position of the C atom in the molecule. These initial findings can be used to narrow down the chemical families from which suitable reference for the comparison to methane-based fuels could potentially originate. 965



BZb



  

  

Bbb 

-b

         

  



`;

  ,X6

2b


( (W=

,>XQ;8a,W=@;,%[;,

Ub

  



Zb ,>

 

3X6 ,>W %6



8>

b

7Zb

 b

Z

2

U

-

U

Figure 2.15: Dependency of RON according to the molecular structure changes of hydrocarbons [72]

The research octane number (RON) index is a measure of the knock resistance property for gasoline-like fuels. The higher the RON, the more knock-resistant

38

2 Fundamentals of Self-Ignition Processes in SI Engines

the fuel behaves during engine combustion in the Cooperative Fuel Research (CFR) engine. However, the octane number (ON) can vary considerably, because as it rates the knock resistance, it consequently depends as well on the engine design parameters of the test engine. The range of the octane number scale is derived from the underlying determination process, which uses two monovalent hydrocarbon fuels, normal heptane and isooctane. By definition, normally n-heptane (n-C7 H16 ) has the RON value 0 and isooctane (C8 H18 or 2.2.4-trimethylpentane) has an RON of 100. These two representative fuels were selected because their resistance to knock differs considerably. Isooctane was more knock-resistant than all relevant gasoline-like fuels available at the time. Blends of the two fuels define the knock resistance of intermediate octane numbers by the share of isooctane. The CFR engine is a robust fourstroke overhead-valve engine with an adjustable compression ratio from 3 to 30, and while the engine is running, a mechanism that elevates or lowers the cylinder and cylinder head assembly relative to the crankcase, thus enables to adjust the compression ratio. Another test method for determining the knock tendency of a fuel is the motor octane number (MON) which has been additionally established beside the RON by American Society for Testing and Materials (ASTM) [2]. This one is performed under stricter boundary conditions concerning the knock influencing parameters like the inlet mixture temperature and the advanced ignition timing. This is the reason why the MON is always lower than the RON. But still, the testing procedures’ boundary conditions only partially represent the operating conditions of thermally highly stressed boosted engines of today. Nevertheless, the octane number classifies the knock resistance of different fuels measured under similar conditions. That makes a direct comparison of their fuel-specific anti-knock properties possible. Although the test procedure is technically outdated and can only predict the knock resistance to a limited extent for modern, highly efficient combustion engines, it is the most common method to evaluate the knock performance of liquid fuels. The numerical difference between the RON and MON is the fuel sensitivity, see eq. 2.17. For supercharged SI engines, the fuel sensitivity can describe the knock-resistant better [64]. Fuel sensitivity = RON − MON

eq. 2.17

2.6 Knock Resistance of Fuels

39

The octane index (OI) is typically determined by the two standard test conditions of the MON and RON and is dependent on the engine operation conditions, see eq. 2.18 where K is the weighing factor an S is the fuel sensitivity [47]. OI = K · MON + (1 − K)RON = RON − KS

eq. 2.18

Any blend ratio of the PRF, a mixture of isooctane and n-heptane, has by definition the same octane rating according to both test procedure determination methods (RON and MON). Since the PRF consists of paraffins, it may be expected that other fuels, consisting mainly of products of the same chemical family to have little fuel sensitivity, compare eq. 2.17. In contrast, olefins and aromatics of different chemical molecular structure have high fuel sensitivity. In general, gasoline consists mainly of natural components of crude oil, which contains a high proportion of saturated hydrocarbons having a low sensitivity. Whereas cracked or reformed gasoline, which contains mainly unsaturated hydrocarbons, have a high sensitivity [29]. For gaseous fuels, the corresponding classification number to describe the knock performance of gaseous fuels is the MN. The determination is comparable to the procedure for determining the research octane and motor octane number of liquid fuels. The MN was developed in an Research Association for Combustion Engines (FVV) research project in cooperation with the Institute for Internal Combustion Engines List (AVL) and introduced at the end of the 1960s [14]. The investigation included binary and ternary gaseous mixtures consisting of gaseous components, which are predominantly present in natural gas. The mixtures were as well examined in a CFR engine within the same procedure as for the liquid fuels. The methane number is defined as the proportion of methane of the reference mixture H2 and CH4 that knocks at the same compression ratio and boundary conditions. The knock limit is determined with the use of electronic knock sensors. Among gaseous fuels, pure methane is characterized by its very high anti-knock properties. Thus, the methane number of 100 corresponds to a RON of approximately 120 to 140 [4] [29] [54]. The setting parameters of the mentioned test methods are recapitulated and shown in Table 2.1

40

2 Fundamentals of Self-Ignition Processes in SI Engines

Table 2.1: Testing methods to determine the RON, MON and MN [2]

PRF n in rpm ε IT in ◦ CA T2 in ◦C

Research Octane Number (RON) DIN EN ISO 5164 n-heptane iso-octane 600 4 – 16 13 51.7 ± 5 (air)

Motor Octane Number (MON) DIN EN ISO 5163 n-heptane iso-octane 900 f(ε) 14 – 26 149 (mixture)

Methane Number (MN) AVL methane hydrogen 900 5.5 – 21 15 25 (mixture)

The fundamental ASTM standard methods to rank the anti-knock performance of a fuel can not unconditionally predict how it will behave under knocking conditions in different engines. In the development history of engines, upcoming requirements and new challenges have continuously advanced further engine optimizations so that several knock influencing factors have changed over time. In the work of [18], [17] and [39] a relation between the calculated AVL methane number of a gaseous mixture and the knocking behavior in a CFR engine was established. Therefore, the RON values were examined for different gases. Genchi investigated in [22] the knock resistance of gasoline / natural gas mixtures. Figure 2.16 shows the MON as a function of the percentage of natural gas in a gasoline / natural gas mixture, measured in a CFR Waukesha engine. A calculated natural gas MN of 80 was stated. A similar qualitative course was shown in the investigations of [54] for the octane number as a function of the MN.

2.6 Knock Resistance of Fuels

41

Figure 2.16: Measured MON of natural-gasoline blends as a function of the mass percentage natural gas [22]

Antiknock Agent Tetraethyl Lead Since the anti-knock properties of a fuel are indirectly related to the efficiency of the engine, highly knock-resistant fuels are of particular interest. The octane number of hydrocarbon fuels can be increased by additives like an anti-knock agent, which is less expensive than modifying the fuel’s hydrocarbon composition by refinery processing. The discovery of the strong anti-knock effect of tetraethyl lead (TEL) was a milestone in the development of gasoline engines. Tetraethyl lead Pb(C2 H5 )4 is one of the most effective additives among the well-known anti-knock agents. The lead atoms are incorporated by the tetrahedral molecular structure and inhibit the reactivity of hydrocarbons. Due to the arrangement of the four ethyl groups in this molecule, it has lipophilic and hydrophobic properties, resulting in good solubility in oily solutions such as fuels. The alkyl decomposes and lead oxide inhibits the pre-flame chain branching reaction and therefore decelerating the reaction rate, as a consequence, autoignition in the end-gas is delayed [29]. Lead has little effect on

42

2 Fundamentals of Self-Ignition Processes in SI Engines

the two-stage ignition respectively on the cool flame, compare figure 2.7. The ASTM determination method of RON values above 100 uses isooctane, which has, by definition, a RON of 100, is intermixed with TEL. Comprehensive measurements and mixing tables were at that time created to give the exact ratios to receive a certain RON between 100 and 120. To produce mixtures with a RON above 120, the mixing ratio can be determined using equation eq. 2.19 [2]. The knock resistance cannot be increased arbitrary with the TEL effect, but follows the natural logarithm and converges, as shown in figure 2.17.

135 130

ASTM Table

Equation 2.19

Measurements Calculated eq.

125 120 115 110 105 100 0.0

0.5

1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 Proportion tetraethyl lead in isooctane in ml / l

5.0

Figure 2.17: Research octane number as a function of the admixed amount of the anti-knock additive TEL [2]

28.28(T ) RON  = 100 + 1 + 0.736(T ) + [1 + 1.472(T ) − 0.035216(T EL)2 ]0.5 >100

with T =

eq. 2.19 Tetraethyl lead (TEL)

2.6 Knock Resistance of Fuels

43

Because of health and environmental reasons, admixing TEL into gasoline is not any more permitted for passenger cars and vehicle of the transportation sector. Low leaded fuel is only occasionally used in aircraft piston engines. The accompanying disadvantages concerning the toxicological aspects and the use of catalytic devices for the exhaust after-treatment has led to alternative solutions. Oxygenates (oxygen-containing organic compounds) – alcohols and ethers – have excellent anti-knock quality because of their high evaporation enthalpy. Fuel evaporation starts with the injection and continues in the manifold as fuel droplets move with the airflow. For direct-injected engines, the mixture cooling is more effective. The higher the remaining non-evaporated portion of the evaporation enthalpy at inlet closure, the more heat is extracted from the mixture in the combustion chamber. This influences the prevailing temperature conditions at the beginning of the compression stroke.

3 Experimental Preparations In order to eventually obtain measurement results that help to understand the described combustion phenomena better, it is first necessary to create a test fuel matrix, in particular the composition of the gas mixtures and the selection of the appropriate liquid fuels. Technical gases of specific mixture composition require a corresponding preparation in advance. For the comparison of their knocking behavior, gaseous and liquid fuels are to be investigated under identical conditions. This results in necessary adjustments for the engine test bench and the fuel supply infrastructure as well as the two individual injection systems. The knocking heat release traces of the examined methane-based fuels differ in the qualitative comparison from the typical trace of a gasoline-like fuel as described in chapter 1 and this can not be thoroughly explained by the generally accepted comprehensive knowledge about knocking combustion. To find the root cause and to gain a better understanding of the knocking phenomena of gaseous fuels, a selection of knock-resistant liquid fuels is made, to examine them at the same engine map point and similar knock limit as the actually very knock-resistant methane-based fuels. Finding mutualities in the behavior of knocking liquid fuels at similar early combustion positions may help facilitate the evaluation. The challenge is to reduce the methane number (MN) for the gaseous fuel mixtures on the one hand and to increase the research octane number (RON) for the liquid fuels on the other hand to such an extent that the fuel’s knock limits converge. In the investigations carried out here, both approaches were pursued.

3.1 Investigated Fuels The selection criteria for the liquid reference fuels are primarily based on their anti-knock performance. Since there is no direct correlation between the meth© The Author(s), under exclusive license to Springer Fachmedien Wiesbaden GmbH, part of Springer Nature 2021 M. Eberbach, Knocking Combustion of Methane-Based and Highly Knock Resistant Liquid Fuels, Wissenschaftliche Reihe Fahrzeugtechnik Universität Stuttgart, https://doi.org/10.1007/978-3-658-35178-6_3

46

3 Experimental Preparations

ane number and the octane number which can predict the knock resistance of fuels in modern combustion engines, the selection of the test fuels should cover a wide RON / MN range. On the one hand, very knock-resistant liquid fuels and, on the other hand very knock-likely gaseous fuels are to be investigated, so that their knock limits overlap for the same engine operating point at similar MFB50 positions. Methane-based Fuels The fuel matrix of the gaseous fuels was defined by selecting individual components for the resulting mixtures representing a wide range in MN grades. The composition specifications are calculated with the benchmark tool “GasCalc” from SmartSim GmbH. This software tool calculates the proportion of the admixing component to achieve the specified methane number grade based according to the AVL calculation method. All necessary methane number grades could be composed through binary gaseous mixtures which are based on pure methane with admixed additional individual components. Thus, the methane number is only defined by the percentage of the admixed gas components. Depending on the components’ saturation vapor pressure and the predominating partial pressure, the maximum permissible total filling of the individual gas cylinders is restricted. Still, the total filling of one gas cylinder should be sufficient to finish at least one measurement series to ensure comparability and keep deviations low. In order to ensure that the experiments are close to realistic applications, all admixing components that could occur in natural gas have been taken into account for the fuel matrix. The first batch of mixtures was composed and certified externally. The supplierdependent manufacturer- and analysis tolerances must be taken into account to fulfill the mixture target of the predefined parameters concerning the MN grade. Therefore, the worst-case deviation for each composed mixtures was calculated in preceding error analysis. For the investigations, the knock resistance of the mixtures are decisive to carry out an accurate comparison. The tolerance is set at plus / minus one point for the methane number. Assuming a mixing accuracy of 5 % and additionally the maximum analytical tolerance of 2 % for the methane-ethane mixture MN = 74.8, the maximum

3.1 Investigated Fuels

47

allowed MN deviation of one point would be exceeded by 0.12, see table 3.1. The specifications regarding the maximum allowed deviation could be met for all methane-ethane mixtures. Table 3.1: Deviations resulting for different combinations of possible manufacturing accuracy and analytical fault tolerance exemplary for the methane-ethane mixture with a MN grade of 74.8 MA = manufacturing accuracy / AT = analytical tolerance [ %] +5 % / +2 % +2 % / +1 % +1 % / +0,5 % Target value -1 % / +0,5 % -2 % / +1 % -5 % / +2 %

Deviation MN

Methane number

[-] 1.12 0.59 0.2

[-]

74.8 0.1 0.18 0.55

Ideally, the gradation of the methane number should decrease in even steps until a mixture composition is reached, which can still be transferred into a controlled knocking combustion. The compositions of all mixtures are listed in table 3.2 including the mixing accuracies and the actual deviations of each mixture. With this procedure, the calculated deviations in the methane number grade for each mixture could be set accurately. The maximum deviation could even be kept well below half of the predefined threshold of 1 point in the MN grade. At the same time, the planning and design of an in-house gas mixing system (GMS) for the test bench was initiated, which provides gas mixtures of very high precision in real-time during engine operation. This makes it possible to create different mixtures with a composition similar to natural gas in varying concentrations of the admixed components. The GMS and its specifications are described later in the chapter, see section 3.3.

48

3 Experimental Preparations

Table 3.2: Methane-ethane mixtures and production tolerances Tolerance MA = manuf. accuracy AT = analytical tol. deviation is (MA 5 %) + 2 % AT (+ 0.033) - 2 % AT (- 0.033) deviation is (MA 5 %) + 2 % AT (+ 0.081) - 2 % AT (- 0.081) deviation is (MA 2 %) + 1 % AT (+ 0.060) - 1 % AT (- 0.060) deviation is (MA 2 %) + 1 % AT (+ 0.089) - 1 % AT (- 0.089) deviation is (MA 2 %) + 1 % AT (+ 0.13) - 1 % AT (- 0.13) deviation is (MA 2 %) + 1 % AT (+ 0.18) - 1 % AT (- 0.18) deviation is (MA 2 %) + 1 % AT (+ 0.25) - 1 % AT (- 0.25) deviation is (MA 2 %) + 1 % AT (+ 0.34) - 1 % AT (- 0.34) deviation is (MA 2 %) + 1 % AT (+ 0.46) - 1 % AT (- 0.46) deviation is (MA 2 %) + 1 % AT (+ 0.64) - 1 % AT (- 0.64)

dev. MN

0.12 0.11 0.07 0.07 0.12 0.13 0.19 0.29 0.23 0.07 0.12 0.16 0.18 0.19 0.16 0.17 0.17 0.30 0.13 0.14

MN

Hu

CH4

C2 H6

[-] 94.71 94.59 94.82 90.07 90 90.14 84.95 84.83 85.08 79.47 79.66 79.76 74.8 74.57 74.73 69.89 69.77 70.05 64.54 64.36 64.73 59.14 58.98 59.31 54.11 53.94 54.41 48.68 48.55 48.82

[MJ/kg] 49.95 49.95 49.95 49.84 49.84 49.85 49.76 49.76 49.76 49.64 49.63 49.64 49.48 49.47 49.48 49.3 49.3 49.31 49.07 49.06 49.08 48.8 48.79 48.8 48.47 48.46 48.48 48.09 48.08 48.1

[mol %] 98.35 98.32 98.38 95.92 95.84 96.01 94.03 93.97 94.09 91.09 91.00 91.18 87.05 86.92 87.18 82.21 82.03 82.39 75.35 75.1 75.6 66.14 65.8 66.48 53.69 53.23 54.15 35.98 35.34 36.62

[mol %] 1.64 1.68 1.61 4.07 4.15 3.99 5.97 6.03 5.91 8.90 8.99 8.82 12.95 13.08 12.82 17.79 17.97 17.61 24.65 24.9 24.4 33.86 34.2 33.52 46.31 46.77 45.85 64.02 64.66 63.38

3.1 Investigated Fuels

49

Liquid Fuels As mentioned above, the liquid fuels for the planned experiments must be significantly more knock-resistant than conventional regular gasoline. The knock resistance could be increased by adding anti-knock additives. Another possibility is to select hydrocarbons from other chemical families with higher knock resistance. In section 2.6, figure 2.14 and 2.15 show how the molecular size and structure of different chemical families as well as their number of carbon atoms influence the knocking behavior. However, there is no generally applicable correlation between MN and RON which can predict the knock resistance in modern engines with highly developed combustion processes. In literature, the RON data for the same hydrocarbon can differ considerably. Rahmouni et al. [54] have summarized the results of investigations among different authors regarding the measured RON of pure gaseous components and mixtures. For the investigations carried out here both approaches were pursued; first, a series of liquid mixtures with increasing knock resistance was composed by adding the anti-knock agents TEL, and then representatives from other chemical families of hydrocarbons were selected. Isooctane Enriched with TEL The ASTM has determined the mixing ratios of isooctane and TEL for octane grades above 100 on the basis of experimental investigations. Based on this data, a liquid fuel matrix with an evenly graduated octane number range was composed. The association provides accurately graded mixture tables for isooctane and TEL mixtures between RON 100 and 120. To investigate highly knock-resistant fuels above a RON values of 120, the mixture ratio can be extrapolated by eq. 2.19 as discussed in section 2.6. In cooperation with a laboratory specialized in fuel analysis, it was possible to produce appropriate mixtures from the probably last TEL deposits in Germany. However, the maximum mixture volume for each octane number stage was two liters, as there are only residual amounts of pure TEL in Germany. The compositions with the resulting RON are shown in the table 3.3.

50

3 Experimental Preparations

Table 3.3: Liquid fuel matrix: Tetraethyl lead admixed in isooctane RON 105 110 115 120 125

TEL [ ml / l ] 0.123 0.339 0.734 1.509 3.123

Highly Octane-Rated Hydrocarbons The limited information data for highly octane-rated hydrocarbons makes the selection quite difficult. Ultimately, only test bench measurements engine show how knock-resistant the fuels actually are. As mentioned above, aromatics have a favorable molecular structure, so that they can be used as additives or as knock-resistant fuel themselves. Table 3.4 shows the selection of the liquid knock-resistant test fuels. The P1 Fuel was composed by a fuel laboratory exclusively for the knock investigations and provided including a RON determination according to the ASTM 2699 method. The other liquid test fuels originate from the aromatic family. The RON ranges result from the different specifications in the literature. On the one hand, this is due to the different experimental setups, on the other hand, there are no suitable PRFs for very knock-resistant fuels, so that RON grades above 120 were usually determined mathematically. The mass-related calorific value of fuels and the volumetric calorific value of stoichiometric mixtures change according to the fuels specifications. The volumetric calorific value of the stoichiometric mixtures remains approximately at the same level. In order to keep the injected energy per working cycle constant for different fuels at the same engine operating point, the calorific value of the mixtures is adjusted by the inlet charge pressure [52].

3.1 Investigated Fuels

51

Table 3.4: Specifications of the selected liquid test fuels [29] [26] [19] [70] [19] [73] [7] [1] [57] [51] [30][48] [32] liquid test fuels

RON

Hu

P1 fuel, mixture of mesitylene C9 H12 isopentane C5 H12 Toluene C7 H8 Acetone C3 H6 O Benzene C6 H6 Xylole-m C8 H10 Xylole-o C8 H10 Xylole-p C8 H10 Methyl tert-butyl ether C5 H12 O Dicyclopentadiene C10 H12

[-] 120 119 – 171 92 120 – 124 110 99 – 124 115 – 162 101 – 129 117 – 155 110 – 135 108 – 229

[MJ/kg] 41.01 40.78 45.36 40.94 28.54 39.56 42.08 40.93 42.10 34.22 37.22

ign. temp [◦ C]

Δvap H ◦

550 420 535 527 555 540 465 540 435 500

47.1 24.7 38.1 30.1 394 43.2 42.1 41.6 30.2 42.4

[kJ/mol]

Gaseous Primary Reference Fuels Composed with H2 and CH4 The knock resistance of hydrocarbons is determined in a CFR engine, using the methods described in chapter 2.6. The methane numbers of the gas mixtures have been determined with the AVL calculation method. To minimize the engine design specific influences on the knock characteristic of the fuels, primary reference fuels and test fuels are analyzed in the same engine. For this purpose, certified binary methane-hydrogen mixtures with up to 50 % hydrogen are prepared and examined according to their knocking behavior in the vicinity of the knock limit. The methane-hydrogen mixtures are listed in table 3.5.

52

3 Experimental Preparations

Table 3.5: Fuel matrix: Methane-hydrogen mixtures in mol % with the resulted MN deviation Methane 50 60 70 80 90 100 – 91

Hydrogen 50 40 30 20 10 0–9

MN deviation 0.00 -0.14 0.05 0.08 0.02 GMS

3.2 Fuel Injection The knocking behavior of a fuel cannot be classified sufficiently by its molecular structure or the conditions prevailing in the combustion chamber. In the inlet section, there are influencing factors that affect the external mixture preparation. To ensure equal and reproducible boundary conditions, for all test fuels of both gaseous and liquid state, the fuel is preconditioned and supplied by the PFI. How far the position of the injection nozzle had to be dislocated upstream depends on the time and flow condition. The predetermined inlet temperature of the mixture should be reached at the inlet manifold at the latest. To ensure comparable conditions, the predefined mixture temperature and a very good homogeneity at the inlet valve closes i.e. at the beginning of the compression stroke have to be achieved. Therefore, it was necessary to install two independent injection systems. Besides the position, the injection timing was optimized as well. An additional low-pressure transducer was installed close to the fuel injection valves. The fuel injection was adjusted to the optimum flow motion in the intake manifold. With an additional low-pressure indication, whose pressure sensor was installed in the vicinity of the injection valve, it was possible to determine an injection characteristic with minimum wall wetting. The pressure trace was evaluated by the indicating system to ad-

3.2 Fuel Injection

53

opt the injection timing individually to engine speed and load point for both, gaseous and liquid fuels. The liquid test fuels are provided in two independent vessels which are pressurized by compressed nitrogen. One is filled with a reference fuel (e.g. toluene) to thermally condition the engine at the measuring point. The other pressure vessel is filled with the test fuel and can be switched on while the engine is running. Thus, also small amounts of expensive test fuels could be examined. Additionally, for fuels with large evaporation enthalpy, a heat exchanger was installed to precondition the inlet air temperature. The injected mass is measured by a Siemens Sitrans Coriolis mass meter in front of the injection valve. To avoid wall wetting, a single-hole nozzle high-pressure injection valve was mounted in the inlet air pipe bend to achieve a good level of atomization. The rail pressure was set to 120 bar to ensure a sufficiently high mixture homogeneity for the knock investigations before the air-fuel mixture enters the combustion chamber. Comprehensive preliminary investigations with Toluene indicated that a further rail pressure increase up to 250 bar no longer had any influence on the knock sensitivity, nor was the fuel temperature increase up to 70 ◦C found to have any effect on the knock rate of the investigated fuels. The gaseous test fuels are likewise injected into the intake air flow. With two gas valves installed facing each other in the intake manifold. The gaseous fuel supply can optionally be selected from four different cascades. One supplies the mixture from the GMS to the engine, another is connected to a bundle of cylinders filled with natural gas, and two others are coupled to individual gas cylinders which means that pure gaseous components or premixed gaseous mixtures can be investigated. The control system ensures that only one cascade can be selected at the same time and that the cascade is emptied completely before the next experimental gas can be applied. Each operating point is set and stabilized using natural gas. In contrast to the expensive experimental fuels, this is available in sufficient quantities. The gas mass is again determined with a second highly accurate Sitrans Coriolis mass meter. Several pressure stages expand the gases to the injection rail pressure of 10 bar.

54

3 Experimental Preparations

3.3 Gas-Mixing System For this project, a high-performance GMS was installed at the Institute of Automotive Engineering (IFS) of the University of Stuttgart to compose gas mixtures from individual components as they occur in natural gas. The mixture proportion of each individual component can be variably adjusted within the defined limits during engine operation. This results in two significant advantages: On the one hand, the required gas mixtures could be mixed and supplied in real-time and on the other hand, the knock resistance, i.e. the MN of the mixture can be continuously adjusted during engine operation by increasing or decreasing the mass flow rate of the appropriate component. Methane, as in natural gases, always represents the main and basis component of the composition. Binary and ternary mixtures and optional one inert gas component can be composed. The exact molar percentage of the admixing component is then being added through real gas calibrated mass flow controller (MFC) and a precise Coriolis mass meter. Different real gas calibration curves are saved on each MFC, so that different gases and flow rates can be admixed through the same MFC. Figure 3.1 shows the schematic set-up of the GMS, the maximum mixing quantities of each gas and the possible component combinations. Butane and propane are in liquid state in the gas cylinder, and after the predefined amount is set, they evaporate immediately in an electric heater, just before they are added to the methane. The following individual components are connected to the GMS. • Methane CH4 • Ethane C2 H6 • Hydrogen H2 • Propane C3 H8 • Butane C4 H10 • Inert gases: CO2 and N2 The sensitive measuring instruments are mounted on a panel which is installed in a separate cabin next to the engine test bench. The purity of the individual gas components must exceed a certain grade to ensure the accuracy of the resulting mixture.

Figure 3.1: Functional structure of the gas mixing system (GMS)

P4.1

D1 DirektVerdampfer

RSV 1.2 0,07bar

V 4.1 N.C.

V 5.3 N.C.

DHV 4.1 ca. 30barü

HPLC-Pumpe 4.1

HPLC-Pumpe 4.2

HPLC-Pumpe 4.3

KH 4.1

KH 4.2

KH 4.3

KH 4.4

Druckhalteventil

IN 4.1 Propan Butan

B1 Ca. 2l gekühlt

RSV 2.4 0,07bar

MFC 2.1 H2 0.11..5.5 g/h C2H6 2.36..118 g/h

V 2.1 N.C.

EPC 1.2

2-10bar Gasflasche Flaschenlager

V 5.4 N.C.

B2 Mischkammer 4 Liter 10 … 12 barü

MFC 3.1 N2 4.8..240 g/h CO2 7.6..380 g/h H2 0.35..17.5 g/h C2H6 5..250 g/h

RSV 7.1 0,07bar

RSV 3.1 0,07bar

P3.1

In 3.1 N2 15 barü

V 3.1 N.C.

MFC 2.2 H2 4.3..215 g/h C2H6 116...5800 g/h

V 2.2 N.C.

N2 Inertisierung

RSV 2.3 0,07bar

RSV 2.2 0,07bar

RSV 2.1 0,07bar

V 1.1 N.C.

P2.2

P2.1

IN 2.2 C2H6 15 barü

P1.1

IN 2.1 H2 15 barü

RSV 1.1 0,07bar

MFC 1.1 CH4 420..21000 g/h

Druckluftventilinsel und Vorsteuerventile

Gekühlter Pumpenkopf Wird kundenseitig überwacht

LFC 4.1 C3H8 3000 g/h C4H10 3000 g/h

RSV 4.1 0,07bar

Druckluftversorgung 4-6 bar

IN 8.1

P8.1

IN 1.1 CH4 15 barü

IN 3.2 CO2

RSV 3.3 0,07bar

P

RSV 3.2 0,07bar

P3.2

15 barü

(optional H2 oder C2H6)

V 5.5 N.C.

OUT 7.1 10 bar

MFC 3.2 N2 108..5400 g/h CO2 170..8500 g/h H2 5.5..275 g/h C2H6 118..5900 g/h

V 3.2 N.C.

RSV 5.1 0,07bar

P5.1

IN 5.1 N2 15 barü

V 6.1 N.O.

V 5.2 N.O.

V 5.1 N.O.

Düse 6.1 1,5mm

Kundenseitig

Gekühlt auf 11°C

12mm

6mm

OUT 6.1 Abzug

3.3 Gas-Mixing System 55

56

3 Experimental Preparations

3.4 Engine Test Bench To ensure a high reproducibility of the measurement results, the test bench and the engine fluids are completely conditioned except for humidity. The research engine is based on an M278 gasoline engine manufactured by the Mercedes-Benz AG. The crank drive was reinforced accordingly to adapt the engine to the high mechanical stress of knocking combustion. Through an adjusted piston crown, a geometric compression ratio of ε = 13.03 is achieved. The elevation is necessary to bring all highly knock-resistant test fuels into controlled knocking combustion without damaging the experimental engine. Figure 3.2 shows the single cylinder research engine.

Figure 3.2: The single cylinder research engine [45]

3.4 Engine Test Bench

57

Figure 3.3: Piston and cylinder head disassembled and cleaned; installed spark plug and pressure transducer; squeeze crevice between cylinder head and piston (left) Figure 3.3 shows the geometric conditions in the combustion chamber with spark plug and pressure transducer on the basis of a disassembled disused test unit. Beside the elevated top land of the peak pressure reinforced piston, the squeezing crevice can be seen on the left side in the picture. The centrally positioned injection valve seat in the cylinder head is plugged with a dummy. To achieve a high homogeneity in mixture preparation, the fuel is already supplied into the inlet air section, as mentioned in section 3.2. Thus, a comparatively low cycle to cycle variation has been achieved. Figure 3.4 show the experimental engine and the test bench.

58

3 Experimental Preparations

Figure 3.4: Engine test bench, inlet air manifold side The crank angle resolved indication measurement data is implemented with an AVL Indiset 631 advanced system in combination with a Kistler crank angle encoder type 2613B. For each operating point examined, 500 consecutive individual work cycles were recorded at a sampling rate of 0.1 ◦ CA. The Crankangle-related indication system with piezoelectric quartz pressure transducers type 6061B (Kistler AG) records the pressure trace in the combustion chamber 3.3. The natural frequency of the pressure transducer is 90 kHz. The water-cooled sensor is characterized by a very low thermal shock sensitivity and is aligned laterally in the combustion chamber roof to avoid pipe oscillation which might complicate or even falsify the correct distinction between knocking and non-knocking working cycles. The amplitudes characteristic of knocking combustion on the pressure trace depends on the location of the pressure transducer in the combustion chamber [58]. The low-pressure indication of intake and exhaust gas pressure is realized by a piezoresistive absolute pressure transducer (Kistler, type 4045A5-V39, and 4075A10-V39). Due to

3.4 Engine Test Bench

59

the high thermal stress, the exhaust gas pressure transducer is located in a cooled switchover adapter (Kistler type 7531). Table 3.6 specifies the measuring equipment and the test bench characteristics. Externally run conditioning systems provide oil, air, water and fuel for the engine in accurate controlled pressure and temperature ranges within the predefined boundaries. The gaseous fuel is injected with two solenoid-controlled Bosch NGI2 valves, just above upstream, is a Bosch high pressure injection valve (HDEV) of the first generation which injects the fuels well atomized 3.4. The FI2 re injection control unit made by Engineering Company Car and Traffic (IAV) is used to control both injection systems and the ignition. The intake manifold is equipped with an additional low-pressure transducer type 4045A5-V39 located close to the fuel injection.

60

3 Experimental Preparations

$ ($ +%(#$

$/+
$,$%



"(
(

 

,!$$

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$









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"(
(  #!
0









 

&  )  0# ,




 



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$(#

+($% +%#(& 

#$$.
 
 %#$(# *%# 






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"(
(
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%#
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#$$.
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Figure 3.5: Overview test bench setup, schematically


!# 

3.4 Engine Test Bench

61

Table 3.6: Specifications of engine and test bench TEST ENGINE: Manufacturer Number of cylinders Cylinder head Stroke Bore Displacement Compression ratio: INJECTION: GAS Injection valve Pressure LIQUID Injection valve Pressure INLET AIR CHARGE SYSTEM: Type Charge air pressure INDICATION SYSTEM: Type Sampling rate PRESSURE TRANSDUCER: Cylinder Intake Exhaust GAS-SUPPLY: CNG Max. pressure Max. flowrate GMS Pressure reduction stages

Daimler AG 1 M278 DELA 86 mm 92.9 mm 583 cm3 13.03

Bosch NGI2 10 bar Bosch HDEV 1 120 bar Screw compressor 3 bar absolute AVL Indiset 631 Adv. / Indicom 0.1 ◦ CA Kistler 6061B, water cooled Kistler 4045A5-V39 Kistler 4075A10-V39, water cooled Typ 7531 switching adapter 36 x 50 l connected 200 bar 30 Nm3 /h 50 l gas cylinder 3

4 Engine Measurements The experimental measurements can be divided into two main parts according to the objectives of the project. First, the fuel matrices with the methane-ethane mixtures and liquid fuels were investigated to generate the result data sets for the comparison of the knocking combustion and the knock limit of the gaseous and the liquid fuels. In the second part, the investigations are completed with the methane-hydrogen fuel matrix to derive the motor methane number relation for each test fuel, and based on this correlation, the reference of the MN to the RON.

4.1 Definition of Engine Operating Points The definition of the engine operating points and variation parameters is determined according to the observed atypical knocking combustion phenomena of the methane-based fuels. Ideally, one particular engine operating point is appropriate to examine the knocking behavior of all test fuels with different knock-resistant properties. However, the fuel should be at least sufficiently knock-resistant to allow at late ignition timing to be set in knock-free operation. To define the boundary conditions, the influence of the inlet temperature and the boost pressure was investigated in preliminary measurements so that for all fuels controlled knocking operation with identical boundary conditions can be provoked. However, it must also be ensured that the most knock sensitive fuels can be measured in a controlled manner without pre-ignition. The results show the influence of the inlet air manifold pressure and temperature on the knock limit position. The according figures are attached in appendix A1.1, their respective boundary conditions are summarized in Table 4.1.

© The Author(s), under exclusive license to Springer Fachmedien Wiesbaden GmbH, part of Springer Nature 2021 M. Eberbach, Knocking Combustion of Methane-Based and Highly Knock Resistant Liquid Fuels, Wissenschaftliche Reihe Fahrzeugtechnik Universität Stuttgart, https://doi.org/10.1007/978-3-658-35178-6_4

64

4 Engine Measurements

Table 4.1: Engine operating points Parameter Air / fuel ratio Inlet air temperature EOP 1 Inlet air temperature EOP 2 Permitted max. exhaust gas temp. Boost pressure (abs.) EOP 1 Boost pressure (abs.) EOP 2 Exhaust gas back pressure (absolute) Permitted cylinder peak pressure Engine speeds

Symbols λ T2 T2 T3 (max, permitted) p2 p2 p3 ≈ p2 p(max, permitted) n

Value 1 65 45 930 2 1.6 160 1500 2000 3000

Unit [-] [◦ C] [◦ C] [◦ C] [bar] [bar] [bar] [bar]

[rpm]

To compare fuels of different chemical states at the same load and stoichiometric fuel ratio, the boost pressure of the liquid fuels was adapted according to the injected energy of the gaseous fuel. This adaptation is due to the constant amount of energy that the fuel has to supply for an individual working cycle to ensure comparability. Depending on the fuel’s calorific value, stoichiometric combustion can result in slightly deviating cylinder fillings.

4.2 Knock Detection Algorithm The implemented knock algorithm analyzes the individual working cycles based on the Siemens-VDO-Algorithm [67], whose knock detection evaluates the oscillations on the cylinder pressure trace provoked by the onset of knocking combustion. The approach of the VDO-Algorithm is visualized in Figure 4.1 and discussed below.

4.2 Knock Detection Algorithm 

window 

65



window

SBc\O>EDU@

     

SBc\OBhigh_SDVVBILOW>EDU@

  window  window            
















   Crank angle>ƒCA@





Figure 4.1: Pressure trace evaluation method of the VDO knock detection algorithm [55] 1. Separation of the signal in two windows corresponding to pmax 2. High pass filtering, rectificating and integrating the pressure trace in both windows results in KIafter or rather KIbefore 3. Definition of the dimensionless basic noise level (BasicPg) to compensate the signal noise 4. If the knock ratio KRAT = (KIafter + BasicPg) / (KIbefore + BasicPg) > KRATlimit ⇒ Individual working cycle is declared as a knocking one. KRATlimit is set at 0.815 for these investigations.

66

4 Engine Measurements

The AVL knock detection algorithm allows only for a rudimentary parametrization. In the previous project, it was noticed that some individual knocking working cycles of methane-based fuels occur without the typical oscillation on the pressure trace, but with primary knocking effects. The VDO algorithm can detect knocking working cycles which show the secondary knocking combustion effect, the onset of oscillations on the pressure trace. The non-oscillating working cycles, as also observed with methane-based fuels, cannot be detected according with this algorithm. Below on this page, there are comparable heat release rate curves of individual knocking working cycles, one with the typical oscillation as it can be identified as knocking by the original algorithm, in the upper part of figure 4.2 and another with the lack of the typical oscillation (bottom) which would not have been identified by the original algorithm. Based on these differences, Scharlipp developed and implemented a complementary criterion in the AVL software environment at the IFS of the University of Stuttgart in a former project [55]. The extended algorithm compares and evaluates the second distinct peak on the heat release trace caused by knocking combustion and relates the amplitude to the gross heat rate released in the working cycle, comparable to a vibe approximation from the onset of knock [27]. The amplitude difference is defined as the sum of the amounts of the maximum positive and maximum negative deviation of the heat release track, see figure 4.3. The relative maximum amplitude difference used to evaluate the intensity of irregular combustion is then referenced to the first, regular heat release maximum. For these investigations, the absolute amplitude ratio was set to a value of 0.7, so that the knock algorithm evaluates individual working cycles equal to or above this threshold as knocking working cycles.

dQH [kJ/°(m3)]

dQH [kJ/°(m3)]

4.2 Knock Detection Algorithm

240 220 200 180 160 140 120 100 80 60 40 20 0 -20 -90 -75 -60 -45 -30 -15 0 15 30 Crank Angle [°CA]

240 220 200 180 160 140 120 100 80 60 40 20 0 -20 -90 -75 -60 -45 -30 -15 0 15 30 Crank Angle [°CA]

67

45

60

75

90

45

60

75

90

Figure 4.2: Knocking working cycles with the typical oscillation (top) and lacking the typical oscillations (bottom) [55]

dQH [kJ/°(m3)]

68

4 Engine Measurements

240 

220 


  200   

 

180 +  
  
 
160   


 


  140 


  120 100 =     80    60  

   

  40 20 0 -20 -90 -75 -60 -45 -30 -15 0 15 30 45 60 75 90 Crank Angle [°]

Figure 4.3: Evaluation procedure of the knock intensity by means of the amplitude ratio [55] The two parallel running knock algorithms are linked with an “or” condition. If only the second distinct heat release peak occurs without the oscillations on the pressure trace, the cycle based result is still detected through the amplitude ratio method.

4.3 Test Bench Measuring Procedure Since the investigations should all be carried out at operating points close to the knock limit, the knock limit has to be defined first. Cyclical combustion fluctuations contribute to the fact that when approaching the knock limit, there is not a spontaneous change to 100 % knocking working cycles, but rather a more or less continuous increase in the knock rate [25]. Especially for fuels consisting of only one chemical compound, a sharp transition into knocking

4.3 Test Bench Measuring Procedure

69

operation is to be expected as soon as the knocking limit is approached, in contrast to fuel mixtures consisting of chemical compounds of different knocking tendencies, where a smooth transition with first individual knocking working cycles is to be expected. In this project, the knock limit is determined by a given number of detected knocking cycles per measurement. One individual working cycle is detected as knocking if a specific evaluation criteria of the above discussed knock algorithm is fulfilled. The knock measurements for each fuel were carried out as an ignition timing sweep, starting at non-knocking engine operation with subsequent successive adjustment of the ignition timing (ignition timing shifted to early) up to exceeding a previously defined knock rate limit. For each ignition timing of the spark sweep, 500 consecutive individual working cycles were recorded. On the one hand, the knock rate has to be determined with sufficient statistical significance, and on the other hand, the engine should run as short as possible at knocking operation. The measurement series is finished as soon as the first adjusted igniting timing exceeds the knock rate of 5 % identified knocking working cycles.

5 Experimental Results and Evaluation The following chapter shows the results of the knock resistance investigations in graphical representation. In the evaluation of the experimental results, emphasis was put on the knock rate along the knock limit and the characteristic trace of heat release for different methane-based fuels and liquid fuels. Subsequently, relevant characteristic values for the assessment of the atypical phenomena of knocking combustion were determined and analyzed.

5.1 Evaluation of the Knock Rate According to the procedure described in section 4.3, an ignition angle variation beginning in the non-knocking range is evaluated until the knock threshold is exceeded, so that a minimum number of working cycles of abnormal, i.e. knocking combustion are included in each batch of measurements. This range comprises at least three to approximately six measuring points of different ignition timings and will be evaluated for all types of fuel. The knock rate at each measuring point is derived from the ratio of knocking to non-knocking working cycles and is plotted versus the averaged MFB50 position. For the evaluation the MFB50 position is used instead of the ignition timing in order to exclude the influence of the varying ignition-delay times of the fuels on the one hand, and on the other hand to address the key question to which extent the combustion position could be the root cause of the observed atypical knocking phenomena. In figure 5.1, the results are plotted as an example for toluene. The sequence of measurement on the test bench starts at late combustion positions, i.e. on the right side of the diagram. Beginning from there, the knock rate increases as the ignition timing adjustment is advanced along the abscissa in the opposite direction, i.e. towards the TDC.

© The Author(s), under exclusive license to Springer Fachmedien Wiesbaden GmbH, part of Springer Nature 2021 M. Eberbach, Knocking Combustion of Methane-Based and Highly Knock Resistant Liquid Fuels, Wissenschaftliche Reihe Fahrzeugtechnik Universität Stuttgart, https://doi.org/10.1007/978-3-658-35178-6_5

72

5 Experimental Results and Evaluation

10 Toluene Varified knock resistance

Knock rate

5

Spark advance

0 -2

0

2 4 6 8 10 12 14 16 18 50 % Mass fraction burned in ° CA a. TDC [MFB50]

20

Figure 5.1: Evaluation procedure to determine the MFB505 % position Due to the ignition timing adjustment, the increasing pressure and temperature levels induce more severe conditions with regard to the self-ignition processes in the end gas and thus gradually increasing the knock rates. For each point, the averaged MFB50 position is calculated. With an exponential fit function, see eq. 5.1, the development over the entire range of the results can be described at the knock limit, which was defined at the beginning of the investigation at 5 %. f (x) = a · eb·x

eq. 5.1

In general, the following behavior can be derived from this function. The earlier the ignition timing in a series of measurements, the higher the knock rate. Once the knock limit is exceeded, it further increases disproportionately according to the exponential approach of the curve. As it was not possible to accurately predict and set the knock limit to 5 % during a measurement on the

5.1 Evaluation of the Knock Rate

73

test bench, the 50 % mass fraction burnt at 5 % knock rate (MFB505 % ) position for each fuel can be calculated as the intersection of the exponential function on the 5 % knock limit. Depending on the knock resistance of an additional fuel examined at the same operating point, the results are arranged to the right or left of this curve. In figure 5.1, additional measurement results are shown in grey and dotted lines as an example. All mixtures of one fuel matrix for the same engine operating point are shown in one diagram. The next section presents the results for the binary methane-ethane mixtures. Additionally appendix A.1 shows the progression of the knock rate as a function of the MFB50 position for different parameters and boundary conditions affecting the knocking combustion in the engine.

5.1.1 Methane-Ethane Mixtures The knock resistance of the methane-ethane mixtures varies according to the fuel matrix composed in chapter 3.1 over a wide MN range. Hence, their individual knock limit results in a wide range of crank angles positions close to TDC, where the comparison to liquid fuels with similar knock limit position can be carried out. Based on those, the atypical knocking combustion phenomena are analyzed and compared with the results of the liquid fuels from the following part. The behavior of the gaseous fuels will be explained exemplarily in figure 5.2 for the engine speed n = 1500 rpm. At the far left of the diagram, the resulting curve of methane can be seen which by definition has a calculated MN of 100. Looking at the intersections at the 5 % knock limit, all result curves for mixtures of lower calculated MNs are oriented to the right. Thus, on the one hand methane is the most knockresistant gas examined in the test engine and on the other hand the knock resistance of all subsequent mixtures qualitatively corresponds to the gradually decreasing calculated MN. This means that for low MN the tendency of the mixture to knock is higher, corresponding to the proportion of ethane added. The mixture with the highest ethane content first approaches its knock limit at MFB505 % ≈ 12.3 ◦ CA a. TDC. The mixture shows to be the most likely to knock, which in this case exceeds the knock rate already at combustion positions which are still comparably late and thus at even lower pressure

74

5 Experimental Results and Evaluation

10 MN = 100.0 MN = 94.71 MN = 90.07 MN = 84.95 MN = 79.47 MN = 74.80 MN = 69.89 MN = 64.54 MN = 59.14 MN = 54.11 MN = 48.68

5

0

5

6

7 8 9 10 11 12 13 14 15 16 50 % Mass fraction burned in ° CA a. TDC [MFB50]

17

Figure 5.2: Methane-ethane mixtures EOP1: n = 1500 rpm, MN = 48 – 100, IMEP = 18 bar to 19 bar and temperature levels. The development of the knock rate, i.e. from the first occurring knocking working cycle, within an ignition timing variation until exceeding the knock limit, here, constitutes within a crank angle degree range of 3 to 4.5 ◦ CA. In summary, it can be concluded that with a methane number variance of slightly more than 50 points on the scale of mixtures, the shift of the MFB50 position extends over a range of 6 ◦ CA. Based on the test results, it can be confirmed that the development of the knock resistance of the mixtures corresponds to the calculated determination of the AVL MN. The discussed behavior of the methane-ethane mixtures can be transferred to the results of both experimental series at higher engine speed. Figures 5.3 and 5.3 give an overview of the complete experimental results of the methane-ethane mixtures at higher engine speeds.

5.1 Evaluation of the Knock Rate

75

10 MN = 100.0 MN = 94.71 MN = 90.07 MN = 84.95 MN = 79.47 MN = 74.80 MN = 69.89 MN = 64.54 MN = 59.14 MN = 54.11 MN = 48.68

5

0

5

6

7 8 9 10 11 12 13 14 15 16 50 % Mass fraction burned in ° CA a. TDC [MFB50]

17

10

5

0

MN = 100.0 MN = 94.71 MN = 90.07 MN = 84.95 MN = 79.47 MN = 74.80 MN = 69.89 MN = 64.54 MN = 59.14 MN = 54.11 MN = 48.68

5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24 25 26 27 50 % Mass fraction burned in ° CA a. TDC [MFB50]

Figure 5.3: Methane-ethane mixtures: n = 2000 rpm (top) and n = 3000 rpm (bottom), MN = 48 – 100, IMEP = 18 bar to 19 bar (top)

76

5 Experimental Results and Evaluation

At first glance, the measurement results of the individual mixtures at higher engine speeds show comparable behavior both within the individual ignition timing sweep and in total. By comparing figures 5.2 to 5.3 a trend over engine speed can be observed. The knock limit position MFB505 % for the mixtures of the two lowest MN grades, see figure 5.3 shows an offset towards late MFB505 % positions. At 3000 rpm engine speed, the pressure and temperature levels are achieved for the mixtures with the highest knock tendency (MN = 48 and 54) at early combustion positions. The thermodynamic evaluation of the heat release rate reveals that here, individual working cycles show abnormal combustion even before the ignition spark had fully developed. Moreover, preignition occasionally occurred additionally to the knocking combustion. The pre-reaction kinetics, ignition delay time, and even ignition itself had already occurred before the actual ignition timing, as indicated by the intersection of the falling edge of the ignition signal with the rise of the heat release rate (HRR) for individual cycles, see figure 5.4. Pre-ignition and knocking combustion are two fundamentally different abnormal combustion phenomena as described in section 2.2. Consequently, the results showing a distinct pre-ignition are not considered in the further evaluation, since only the knocking combustion is relevant for the investigations carried out here. A further reduction of the calculated methane number by additional admixing of a knock-sensitive gaseous component, would make an upcoming working cycle with abnormal combustion almost uncontrollable at the same boundary conditions and engine map point. Pre-ignition would be the consequence and the associated mechanical stress could lead to the destruction of the engine. A variation of the engine speed generally affects the charge motion of the mixture, the turbulent burning flame speed and the available time per working cycle. As a result of this, the faster convertion of the energy of the mixture reduces the period remaining for auto-ignition in the end gas whereby the knock probability declines. A deviation from the usually observed tendency of decreasing knock rate with increasing engine speed as shown here is assumed as overcompensation of the described effect by the basically rising temperature level at higher speeds [23] [14].

5.1 Evaluation of the Knock Rate

30 20

77

Methane-based fuel Spark current trace

10 0 -10 -20 -30 -40

-30 -20 -10 Crank angle position in °CA a. TDC [ ]

0

Figure 5.4: Individual working cycles showing abnormal combustion; HRR is already increasing before the spark ignites the mixtures EOP1: n = 3000 rpm, MN = 48 5.1.2 Liquid Fuels According to the procedure in the first part of the experimental measurements of gaseous fuels, the second part investigates the liquid test fuels. The fuels of the isooctane TEL mixtures were examined first. However, it turned out right at the beginning that none of the mixtures could be set without knock at engine operating point 1 (EOP1), even with correspondingly late ignition timing. The knock tendency of these mixtures is much too high to be investigated at the previously defined high-load operating point. Nevertheless, in order to be able to investigate the knocking behavior near the knock limit of the fuels, a new low load point with an inlet air pressure of p2 = 1000 mbar and an inlet air temperature of T2 = 45 ◦C has been established. As a consequence, the results cannot be compared to the gaseous fuels, since for a conclusive knock resistance comparison the fuels would have to be examined under the same boundary conditions. However, the results are noted and are to be explained in the following figure 5.5.

78

5 Experimental Results and Evaluation

10 RON = 125 RON = 120 RON = 115 RON = 110 RON = 105 RON = 100 (Isooctane) 5

0 -4 TDC 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 42 50 % Mass fraction burned in ° CA a. TDC [MFB50]

Figure 5.5: Isooctane-TEL: n = 1500 rpm, p2 = 1000 mbar, T2 = 45 ◦C The arrangement of the result curves shows per se a conclusive behavior, but if compared to pure isooctane, which is shown as a grey-dashed curve in the diagram, it is obvious that the TEL mixtures are not of the alleged RON and therefore obviously not as knock-resistant as stated in the ASTM mixture tables. On the contrary, the added TEL has even reduced the knock resistance of the blends. That was apparently due to the organic compound TEL, which decays over time and gradually loses its knock-increasing effect [44]. The results at higher engine speeds are shown in appendix A2.2. Since these measurements are neither applicable for comparison with gaseous fuels nor for correlation between the MN and the RON, the results of the selected additionally highly knock-resistant liquid fuels are now discussed in detail. Figure 5.6 shows the measurement results of the liquid fuels examined at the original WOT point EOP1 corresponding to the first part of the measurement campaign. The variation of the combustion position at knock (MFB505 % ) re-

5.1 Evaluation of the Knock Rate

79

sults from the fact that different fuels are used and not as before from the variation of the percentage share of one component within a binary mixture. Nevertheless, a larger crank angle range for various knock limits could be examined with this fuel selection, which made it possible to cover a wide overlapping area with the gaseous fuels according for the MFB505 % position at the same operating point and the same boundary conditions as for the gaseous fuels. The knock-resistant fuel benzene allows for a knocking operation at MFB50 positions close to TDC. All other result curves are arranged to the right of this curve, thus behaving accordingly less knock-resistant. Dicyclopentadiene (DCPD) completes the measurement sequence as the most knock sensitive fuel with MFB505 % positions around 15 ◦ CA a. TDC. Methyl tertbutyl ether (MTBE), the P1 fuel, the mesitylene and also the xylenes show their knock limit within a narrow crank angle degree range. Although, except for the P1 fuel, fuels consisting of only one chemical compound were investigated, none of these fuels show an abrupt transition into knocking operation, even if the slope of the exponential regression curve is significantly steeper. Figure 5.6 and 5.7 show the results for higher engine speeds. At 2000 rpm, it was not possible anymore to provoke knocking combustion for benzene by further advancing the ignition timing to early. Since the liquid test fuels were only available in limited quantities for the project, a reduced selection was investigated at higher engine speeds.

80

5 Experimental Results and Evaluation

10

5

Benzene Acetone Toluene Xylene_p Xylene_m MTEB Mesitylene P1Fuel DCPD

0 -2 TDC 2 4 6 8 10 12 14 16 18 20 50 % Mass fraction burned in ° CA a. TDC [MFB50]

22

10

5

Benzene Acetone Toluene Xylene_p Xylene_m Mesitylene P1Fuel DCPD

0 -2 TDC 2 4 6 8 10 12 14 16 18 20 50 % Mass fraction burned in ° CA a. TDC [MFB50]

22

Figure 5.6: Selected liquid test fuels EOP1: n = 1500 rpm (top) n = 2000 rpm (bottom), RON compare table 3.4 on page 51

5.2 Investigation of Atypical Knocking Combustion Phenomena

81

10 Acetone Toluene Xylene_m Xylene_p P1Fuel 5

0 -2 TDC 2 4 6 8 10 12 14 16 18 20 50 % Mass fraction burned in ° CA a. TDC [MFB50]

22

Figure 5.7: Selected liquid test fuels EOP1: n = 3000 rpm, RON compare table 3.4 on page 51

5.2 Investigation of Atypical Knocking Combustion Phenomena For each atypical knocking combustion phenomenon, selected characteristic parameters are evaluated and compared in a graphical evaluation based on the MFB50 position. The results are to show to what extent the root cause of the phenomena is fuel-specific, i.e. is due to the chemical properties of the fuel, and / or is due to the comparatively early combustion position.

82

5 Experimental Results and Evaluation

5.2.1 Onset of Knocking Combustion The onset of knock is determined by the extended criterion of the combined knock algorithm. In this process, the primary knocking effect on the low-pass filtered heat release trace is shown as a second amplitude peak, which in turn is triggered by the additional heat release of the knocking combustion. Between the first and the second heat release peak, the local minimum is defined in this work as the position of the onset of knock in ◦ CA a. TDC, see the heat release rate in figure 5.8 between 0 and 10 ◦ CA a. TDC . Another method to determine the onset of knock could be, as mentioned in section 2.4, to evaluate the onset of the typical knock induced oscillations based on the high-pass filtered pressure trace. However, due to the signal influencing factors, additional concretization measures are necessary to define the exact onset of knock on the basis of this method [61]. For the following evaluations, the determination of the onset of knock for all self-ignition related working cycles will be based on the method of the extended knock algorithm as described. Figure 5.8 shows the original and the filtered HRR-trace with the corresponding pressure trace. At the low point between the two high points on the grey heat release trace at 5 ◦ CA a. TDC the onset of knock is defined. The area under the curve already indicates to what extent combustion has progressed at the onset of knock and gives an indication of the remaining proportion of unburned mixture.

5.2 Investigation of Atypical Knocking Combustion Phenomena

350 300 250

83

140 Cyl. pressure Filtered heat release 120 Heat release 100

200

80

150

60

100

40

50

20

0

0

-50 -50 -40 -30 -20 -10 0 10 20 30 40 50 Crank angle position in °CA a. TDC [ ]

60

-20 70

Figure 5.8: HRR originally (black), HRR low pass filtered (grey) and unfiltered cylinder pressure trace of a knocking working cycle at EOP1: n = 1500 rpm; compressed natural gas (CNG); MN = 87

The total combustion process is measured by the percentage of fuel energy converted within the working cycle and is given in percent [mass fraction burnt (MFB)]. Figure 5.9 shows the onset of knock of all measured knocking working cycles for the gaseous fuels. The ordinate shows the integrated heat release rate in % MFB. The combustion position relative to TDC is evaluated for the MFB50 position of the working cycle and plotted on the abscissa. Each point in the diagram represents one self-ignited affected working cycle. The coloring of the points shows the working cycles that are identified as knock by the amplitude ratio of the extended algorithm (grey triangles). The distribution of the accumulated points show the development of the onset of knock versus the combustion position towards TDC.

84

5 Experimental Results and Evaluation

90

Amplitude ratio < 0.7 Amplitude ratio > 0.7 knocking IWC

80 70 60 50 40 30 20 10 -8

-6

-4 -2 0 2 4 6 8 10 12 14 16 50 % Mass fraction burned in ° CA a. TDC [MFB50]

18

Figure 5.9: Onset of knocking combustion vs. MFB50 position as a function of the amplitude ratio (knock intensity) according to the extended algorithm for all examined gaseous fuels at EOP1; cycle-based evaluation

The first self-igniting related working cycles here occur as the MFB50 position below 13 ◦ CA a. TDC. In the direction towards TDC, i.e. the direction towards the efficiency optimum combustion position, the largest accumulation of points is observed. By further approaching towards TDC, which is only achievable with highly knock-resistant fuels, a gradual decrease in the early onset of knock and the total amount of knocking working cycles is shown. This results in a narrowing pattern of the point distribution which is oriented towards a MFB of 60 % to 70 %. According to the ordinate, the highest density of distributed points is accumulated in a MFB range of 35 % to 70 %. Moreover, it is apparent that as soon as 65 % of the mass fraction has been converted by the regular combustion, no distinct self-igniting related knocking working cycles occur anymore. Basically, early onset knock means that there is still a large

5.2 Investigation of Atypical Knocking Combustion Phenomena

85

quantity of unburned fuel mixture in the end gas at the time of self-ignition. The common understanding of incipient knock by gradually advancing the ignition timing starting at non-knocking engine operation, describes late onset of knock with initially low knock rate for conventional gasoline-like fuels. On this assumption, an extremely rapid heat release of the remaining large amount of unburned mixture would be expected within a very short time leading to severe engine damage. However, this behavior could not be observed here, nor are there any late onsets of knocking combustion after MFBs beyond >70 %. Considering the atypical behavior of gaseous fuels, the question arises even more which behavior the liquid fuels show in this regard at MFB positions near TDC. Figure 5.10 shows the corresponding evaluation.

90

Amplitude ratio < 0.7 Amplitude ratio > 0.7 knocking IWC

80 70 60 50 40 30 20 10 -8

-6

-4 -2 0 2 4 6 8 10 12 14 16 50 % Mass fraction burned in ° CA a. TDC [MFB50]

18

Figure 5.10: Onset of knocking combustion vs. MFB50 position as a function of the amplitude ratio (knock intensity) according to the extended algorithm for all examined liquid fuels at EOP1; cycle-based evaluation

86

5 Experimental Results and Evaluation

It is obvious that the number points in the diagram are less. On the one hand, this is due to the fact that compared to gaseous fuels, fewer liquid fuels and measureing points were investigated at this high load operating point. On the other hand, the characteristic shape of the second heat release peak could differ as the evaluation of the knock intensity is based on the amplitude ratio. A comparison of the knock detection methods is discussed in chapter 5.2.3. Between 6 ◦ CA a. TDC and 13 ◦ CA a. TDC, the most significant accumulation of measuring points with a high knock intensity can be found. However, a definable overall contour of the accumulation can be identified in the figure, which qualitatively corresponds to the results of the gaseous fuels. Consequently, it can be concluded that for the underlying boundary conditions, liquid fuels show a comparable behavior with regard to the onset of knock, both in relation to the MFB50 position as well as to the MFB. As already mentioned, the majority of the working cycles show very early onsets of knock between 40 % to 70 %. The most significant accumulation of severe knocking working cycles, i.e. knocking working cycles of which amplitude ratio exceeds the predefined threshold, are in the same MFB50 range between 35 % to 55 % for liquid and gaseous fuels. Even in the investigation of the liquid test fuels, no gradually appearing knock with a late onset of knock (as anticipated at approximately MFB90) with low knock intensities is observed. However, onset of knock at moderate intensities (i.e. with a small amplitude ratio) occur at very early MFB positions, which indicates that additional to the available unburnt fuel mass, there are other decisive influences on the development of the knock intensity which attenuate, slow down or in some other manner inhibit the distinct second heat release peak. The unburned fuel mass at the onset of knock seems not to be an absolute parameter to describe whether abnormal combustion will occur and at which reaction rate the energy will be converted. The quantitative evaluation of the influencing parameters regarding the onset of knocking combustion discussed above, makes it possible to now investigate the development of the self-igniting working cycles over all measuring points considering the effect of the varied parameters. In the following section, the heat release trace of a knocking working cycle is evaluated for a gaseous as well as for liquid fuel that has been measured at the same EOP. The comparison shows the onset of knock and the distinct shape of the second heat release peak in detail.

5.2 Investigation of Atypical Knocking Combustion Phenomena

87

5.2.2 Comparison of Knocking Behavior of Different Fuels at Similar MFB505 % Position For the comparison of individual knocking cycles, a liquid and a gaseous fuel of similar MFB505 % positions are selected. Figure 5.11 shows the results for toluene and the gaseous mixture with a corresponding MFB505 % position chosen from the database of binary methane-ethane mixtures.

10 MN = 74.8 Toluene

5

0

5

6

7 8 9 10 11 12 13 14 15 16 50 % Mass fraction burned in ° CA a. TDC [MFB50]

17

Figure 5.11: Progression of the knock rate for a CH4 / C2 H6 mixture and toluene EOP1: n = 1500 rpm, similar MFB505 % position The heat release traces of the binary methane-ethane mixture with a MN of 74.8 and the one of toluene are compared in figure 5.12. In addition to the quantitative evaluation of the onset of knock by means of characteristic calculated values, the qualitative course is to be investigated next.

88

5 Experimental Results and Evaluation

350 300 250

Toluene MN = 74.8 Cyl. pressure Heat-release

140 120 100

200

80

150

60

100

40

50

20

0

0

-50 -40 -30 -20 -10 0 10 20 30 40 50 60 70 Crank angle position in °CA a. TDC [ ]

Figure 5.12: Methane-ethane mixture and toluene for EOP1: n = 1500 rpm, MFB505 % = 9.5 ◦ CA a. TDC, unfiltered pcyl trace and filtered HRR (3 kHz-filtered)

The focus is set on the particularity of the heat release trace, especially on the onset of knock and the characteristic shape of the second heat release peak. In fact, the qualitative comparison of the two plotted heat release traces shows that the observed combustion phenomena for gaseous fuels appear to the same extent for the liquid fuel. The comparison to methane-based gaseous fuels can only be realized with very knock-resistant liquid fuels. Fuels of lower knock resistance, e.g. as in the category of conventional gasoline, could hardly be investigated under the same boundray conditions. The same behavior could be observed by comparing the even more knockresistant methane and acetone. Acetone behaves slightly more knock-resistant than methane, as shown in figure A2.3 in the appendix. Both fuels show a clear

5.2 Investigation of Atypical Knocking Combustion Phenomena

89

but relatively late onset of knock. Further comparisons are shown in figures A2.3 to A2.4 in the appendix. The qualitative and quantitative comparison shows comparable characteristics for different types of fuel and chemical stages regarding the onset of knock. In the following section, the second atypical knocking combustion phenomenon is investigated.

5.2.3 Distinct Second Heat Release Peak The second atypical knocking combustion phenomenon is investigated on the basis of the knock intensity (ratio of amplitude) and the shape of the second heat release peak respectively, which is predominantly seen during knocking combustion. The area of the second heat release peak corresponds to the converted energy, initiated by the self-ignition. The amplitude ratio is one measure of the knock intensity. Considering the knock intensity, a differentiation is necessary for the present application, since the evaluation criteria and thus also the definitions of the knock intensity are different for the two used algorithms. The United DEUTA-OTA (VDO) algorithm is based on the secondary knocking effect, i.e. the oscillations on the pressure trace, and identifies the knocking working cycles that do not necessarily show a primary knocking effect on the heat release trace, i.e. a distinct second heat release peak. Whereas the extended algorithm evaluates the intensity of the low-oscillating knocking working cycles. Hence, for the following quantitative evaluations, the intensity according to the definition of the amplitude ratio will be applied. A comparison with the knock intensity according to the VDO algorithm is shown later in this chapter. The knock intensity of both algorithms are expressed in nondimensional key figures. For the VDO algorithm, 1.815 is the threshold of an identified knocking working cycle by definition. Whereas the amplitude ratio threshold of the extended algorithms is 0.7, see chapter 4.2. A working cycle is identified as knock if one of these thresholds is exceeded [55]. Figures 5.13 to 5.16 show the knock intensity as a function of the MFB50 position for both knock intensity definitions for all identified knocking work-

90

5 Experimental Results and Evaluation

ing cycles. On the ordinate the knock intensity is plotted in normalized values. That means both threshold values (VDO: KRATlimit = 0.815; extended algorithm = 0.7) are normalized to 1. Thus, working cycles with knock intensities below the threshold of 1 do not exceed the threshold, but show the characteristic attributes of knocking combustion at a lower extent and the gradual development of the increasing knock intensity.

Figure 5.13: Knock intensity of gaseous fuels (standardized, cycle-based) according to the VDO-algorithm; identified knocking working cycles are colored in grey

5.2 Investigation of Atypical Knocking Combustion Phenomena

91

Figure 5.14: Knock intensity of gaseous fuels (standardized, cycle-based) according to the extended algorithm; identified knocking working cycles are colored dark grey

Regarding the MFB50 position, the absolute number of points detected as knock (highlighted in grey) are distributed over the same crank angle range. For the liquid fuels, it can be seen that the absolute number detected as knocking working cycles were identified according to the VDO method. The density of points above the knock threshold identified by the extended algorithm is comparatively thin. Hence, it can be assumed that liquid fuels possibly show the primary knocking effect less. The highest knock intensities occur in a range between 5 and 12 ◦ CA a. TDC. Adjusting the combustion position (MFB50) towards TDC results in increasing cylinder peak pressures. This leads to a rising temperature level in the end gas, which in turn increases the probability of self-ignition as well as the resulting knock intensity. Simultaneously, the onset of knock is shifted towards early

92

5 Experimental Results and Evaluation

Figure 5.15: Knock intensity of liquid fuels (standardized, cycle-based) according to the VDO-algorithm; identified knocking working cycles are colored in grey MFB positions [16] [28]. At the same time, the combustion process is accelerated, which in turn decreases the remaining time in which self-ignition can occur in the end gas. Generally, a short and TDC close combustion increases the IMEP and the thermodynamic efficiency, which shows a parabolic correlation to the MFB50 position. Increasing wall heat losses and losses of pressure work due to leakage and friction at early ignition time points on the one hand, and increasing exhaust gas enthalpy losses in the combustion process at very late ignition timing on the other hand, result in the efficiency-optimal range in between. [3]. In figures 5.17 and 5.17, the characteristic of the second heat release peak at knocking combustion is being evaluated quantitatively. The knocking working cycles are attributed to the algorithm according to which they were identified.

5.2 Investigation of Atypical Knocking Combustion Phenomena

93

Figure 5.16: Knock intensity of liquid fuels (standardized, cycle-based) according to the extended algorithm; identified knocking working cycles are colored dark grey On the ordinate, the knock intensity evaluated by the amplitude ratio (extended algorithm) is plotted. If a working cycle has been detected exclusively by the extended algorithm, it is marked with grey triangle, all those detected exclusively by the VDO algorithm are marked with wihite dots and the cycles detected by both detection algorithms are colored in black. The grey triangles, i.e. the low-oscillation knocking working cycles, tend to be distributed towards TDC. The white dots, i.e. working cycles which show clear oscillations on the pressure curve but not necessarily the distinctive second heat release peak, occur predominantly at late MFB50 positions.

94

5 Experimental Results and Evaluation

2.0

1.8

bothe algorithms exclusively VDO exclusively ext. alg

1.6

1.4

1.2

1.0 -4

2.0

1.8

-2 0 2 4 6 8 10 12 14 50 % Mass fraction burned in ° CA a. TDC [MFB50]

16

bothe algorithms exclusively VDO exclusively ext. alg

1.6

1.4

1.2

1.0 -4

-2 0 2 4 6 8 10 12 14 50 % Mass fraction burned in ° CA a. TDC [MFB50]

16

Figure 5.17: Exclusively as knock identified IWC assigned to the respective detection algorithm for gaseous fuels (top) liquid fuels (bottom): Intensity according to the amplitude ratio vs. MFB50

5.2 Investigation of Atypical Knocking Combustion Phenomena

95

While the gaseous fuels show result points from all three modes, the distribution for the liquid fuels is slightly one-sided and strongly thinned out. Therefore, it can be assumed that the second atypical combustion phenomenon, a distinct second heat release peak, occurs less frequently with liquid fuels. The majority of the knocking working cycles were detected by the VDO algorithm, i.e. by evaluating the oscillation on the pressure trace and less by the extended algorithm, see figure 5.18. Here, the knock intensities are plotted as determined by the VDO method. All white dots show the working cycles which were detected exclusively by the VDO algorithm and the black dots show all knocking working cycles additionally detected by the extended algorithm. There is no working cycle that was exclusively identified by the extended algorithm. The second atypical knock combustion phenomenon with liquid fuels shows the distinct second heat release peak only sporadically. The knocking working cycles are mainly detected by the secondary knocking effect, the oscillations on the pressure trace.

5.0

bothe algorithms exclusively VDO

4.0

3.0

2.0

1.0 -4

-2

0 2 4 6 8 10 12 14 50 % Mass fraction burned in ° CA a. TDC [MFB50]

16

Figure 5.18: Exclusively as knock identified IWC assigned to the respective detection algorithm for liquid fuels: Intensity according to VDO evaluation method vs. MFB50

96

5 Experimental Results and Evaluation

In order to divide the proportion of non-oscillating knocking working cycles, which only show the primary knocking effect of the distinct second heat release peak, also regarding different influencing variation parameters, a statistical evaluation is shown in the following chapter.

5.2.4 Non-Oscillating Knocking Working Cycles The third atypical knocking combustion phenomenon will be investigated by means of the proportion of low-oscillating knocking working cycles. Already in the previous section, differences for gaseous and liquid fuels have been indicated in this context.

10 Liquid fuels Gaseous fuels

5 3.1 2.4 1.0

0.4

0

0.4

0.2

1500

2000 Engine speed in rpm [n]

3000

Figure 5.19: Proportion of low-oscillating knocking working cycles for liquid and gaseous fuels versus the engine speeds

5.2 Investigation of Atypical Knocking Combustion Phenomena

97

Figure 5.19 shows the average proportion of working cycles identified as knock exclusively by the extended algorithm relative to the total number of working cycles, including those recorded by the VDO method. The proportion of low-oscillating knocking working cycles for the liquid fuels is small for all engine speeds investigated. In the averaged calculation, a decreasing proportion of non-oscillating knocking cycles can be observed over all engine speed for the gaseous fuels. The proportion within the liquid fuels is consistently low with a slight increase for 3000 rpm. Figure 5.20 shows the proportion of non-oscillating knocking working cycles as a function of the knock resistance of the fuels and the engine speeds. The gaseous fuels are subdivided into three methane number categories, the liquid fuels into two knock resistance categories. The division is made at the knock limit MFB505 % = 10 ◦ CA a. TDC, which is based on the experimental measurements in chapter 5.1.2. Benzene, acetone and toluene constitute the one group of the knock-resistant liquid fuels.

10

MN < 70 70 < MN < 85 MN > 85 Remaining. liq. fuels Ben, Ace, Tol

8.9

6.8

5 3.6

3.1 2.1

1.7 0.8

0

0.0

1500

1.6 0.5

0.0

0.3

2000 Engine speed in rpm [n]

0.7

0.3

0.0

3000

Figure 5.20: Proportion of low-oscillating knocking working cycles for liquid and gaseous fuels versus the MN respectively RON

98

5 Experimental Results and Evaluation

All other liquid fuels investigated are included in the other group, the more knock sensitive fuels. The liquid fuels confirm the behavior shown in figure 5.19. In contrast, the gaseous fuels in the methane number range between 70 and 85 show a high proportion of non-oscillating knocking working cycles, which decreases with increasing engine speeds. The liquid fuels show significantly lower proportions of low-oscillating knocking working cycles. In contrast, the gas mixtures show a trend with increasing engine speed, which is also shown in the qualitative development regarding the knock resistance of the gaseous fuels. This leads to the assumption that the decisive influencing factor for the nonoscillating working cycles depends on both, engine speed and knock the resistance of the fuel. The literature shows that in experimental engine investigations on knocking combustion, such low-oscillating working cycles were also observed occasionally. In the work of [49], the absence of oscillation on the pressure trace was interpreted as insufficiently fast energy conversion of the self-ignited end gas. In a series of experiments on an optically accessible two-stroke single-cylinder engine, knock investigations were made on the basis of Schlieren recordings by König and Sheppard [38], in which similarly knocking working cycles were identified without the typical oscillations on the pressure trace, see figure 5.21.

5.2 Investigation of Atypical Knocking Combustion Phenomena

99

Figure 5.21: Low-oscillating cylinder pressure trace and Schlieren photographic sequence showing of deflagration similar combustion of a hot spot auto-ignited working cycle (left to right, 1.08 ◦ CA) [37]

The irregular pressure increase in figure 5.21 is attributed to the primary knocking effect, triggered by the self-ignition processes in the end gas. Due to the number of exothermic centers, the authors observed a diffuse flame front in the Schlieren recordings, describing this type of combustion as “autoignition initiated at several centers in the end-gas, with the character of a homogeneous volume reaction rather than a propagating flame”. However, it is also noted that the complete auto-ignition process takes a relatively long time compared to a volume reaction. The reason for the delayed reaction could be the low temperature of the end-gas compared to severe knock. This leads to the assumption that low-oscillating knock is caused by the formation of a kind of diffuse deflagrative combustion, which could originate from exothermic cen-

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ters in the end gas of comparatively low temperatures [37]. Consequently, if an increased temperature (decreasing temperature gradient) and mixture homogeneity in the end gas is assumed with increasing engine speed, this could explain the decreasing proportion of low-oscillating knocking working cycles with increasing engine speeds in the experiments of gaseous fuels.

5.2.5 Influencing Parameters on the Atypical Knocking Combustion Phenomena Besides the reaction kinetic processes, thermodynamic interactions and inhomogeneities of the end gas, the combustion chamber geometry also influences the knocking characteristics. This multidimensional interaction can lead to a variety of knocking characteristics, which, due to the highly unsteady process in the end gas, can lead to self-ignition followed by combustion and pressure waves. The resulting self-ignition-induced pressure oscillations are thus stochastic. As the flame extinguishing process close to the wall begins, pockets of different temperature gradients are formed to several partially separated end gas areas. If an exothermic center initiates self-ignition in an end gas pocket, it has only a limited proportion of the total remaining unburned fuel mass at its disposal, which volume limits the energy conversion rate and probability of a pressure wave. In the case of adjacent but separated end gas pockets, interactions can also occur, since the pressure wave caused by a selfigniting center is only weakened but not stopped by passing the burnt gas. The oscillating pressure amplitude results from the pressure disturbance potential due to chemical reactions and the speed of sound, which increases as density and temperature are increased. In addition, the volume of the end gas pocket determines the range of these variables. The temperature gradient is furthermore characterized by the relative thermal inhomogeneity in the volume of the unburned mixture [36] [37].

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101

cylinder wall

end gas pocket Text Text

combustion gas

regular flame front

Figure 5.22: Characteristics of an end gas pocket according to [36]

Especially for engines with a low stroke-bore ratio but high compression ratio, an increased surface-to-volume ratio at the top dead center exists. Additional squish crevices may be unavoidable when the geometric compression ratio is increased which, with an early onset of knock close to TDC, as observed in these investigations, can influence the combustion process. As observed in chapter 5.2.1, the knocking combustion occurs at early MFB fractions within the working cycle. Almost the same behavior was observed for the liquid fuels, in contrast to the understanding of incipient knock when approaching the TDC gradually, late onset of knock and low knock intensities of individual working cycles are expected. The fact of exclusively early onsets of knock at varying intensities, and none beyond MFB fractions of 70 % regardless of the chemical properties within the test fuels, show that the influence of the MFB50 position plays a role in the knocking characteristics. For onset of knock at very early MFB50 positions, only sporadic knocking work cycles occur, where most of the cylinder charge (MFB) has already been converted at the onset of knock, see figures 5.9 and 5.10. In the relevant literature, investigations could be found in which the onset of knock at early positions was observed as well. In the investigations of Chun [16], the early MFB50 positions are examined as a function of the thermodynamic temperature history and the prevailing temperature in the end gas. Furthermore, temperature differences caused by the thermodynamic combustion history of the previous working cycle could influ-

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ence the mechanisms in the end gas of the subsequent working cycle (residual gas fraction and heated combustion chamber walls). If knock occurs close to TDC, it can be assumed that at this point in time, the piston has only moved slightly down and the turbulence generated by the squish area could influence the self-ignition related process in the end gas. Therefore, the resulting backflows can either accelerate the normal combustion process by generating turbulence at the squished crevice or slow down the reaction kinetic processes in the end gas due to the lower temperature of the reverse flowing mixture [65]. The CFD simulations in the investigations of Imaoka [31] show that the backflows caused by the squish crevice have a temperature reducing effect on the unburned mixture during the flame propagation and thereby influence the pre-reaction processes in the end gas. The investigations were performed for a combustion chamber with pent-roof geometry and show“that a large quantity of unburned fuel remains in the squish are”. According to the author, the temperature of the unburned mixture decreases “markedly” in the vicinity of the squish areas due to the reverse flowing mixture out of the squish crevices. A comparable effect is achieved by storing the unburned mixture in the top land piston area during the compression stroke, which returns to the combustion zone as a cooling backflow after the pressure reversal point (PRP) has been exceeded [56]. Due to the high surface-to-volume ratio of the top land, the back flowing mixture has reached a comparatively low temperature [5]. The backflows have a retarding effect on the reaction kinetic processes taking place in the end gas and thus cause a knock-reducing effect. The results in the following figure show to what extent these effects could influence the onset of knock and the second heat release peak in the present investigations. If the “freezing” of the self-ignition-relevant pre-reactions is taking place for a short time, the regular flame consumes further mixture volume that is no longer available to the end gas for the development of a possible self-ignition.

5.2 Investigation of Atypical Knocking Combustion Phenomena

15

103

Amplitude ratio < 0.7 Amplitude ratio > 0.7 knocking IWC

10 5 0 -5 -10 -15 -8

-6

-4 -2 0 2 4 6 8 10 12 14 16 50 % Mass fraction burned in ° CA a. TDC [MFB50]

18

Figure 5.23: Relative position of the onset of knock to the pressure reversal point as a function of the MFB50 position for all gaseous fuels at EOP1, cycle-based

Figure 5.23 shows the relative position in degrees crank angle of the onset of knock to the PRP, as a function of the combustion position MFB50. Consequently, at zero on the ordinate, the knock onset (KO) and the PRP are right next to each other, for positive values, self-ignition occurs after the PRP and vice versa for negative values. Figure 5.24 shows almost the same behavior for liquid fuels. Despite the reduced number of measurements with liquid fuels, the similarity to gaseous fuels is shown by the qualitative distribution of the points and the overall outline. As before, the grey triangles are the working cycles identified as knock. It is obvious that, regardless of the type of fuel and the engine speed, almost no knocking working cycles above the defined threshold occur if the onset of knock has not started before the PRP.

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15

Amplitude ratio < 0.7 Amplitude ratio > 0.7 knocking IWC

10 5 0 -5 -10 -15 -8

-6

-4 -2 0 2 4 6 8 10 12 14 16 50 % Mass fraction burned in ° CA a. TDC [MFB50]

18

Figure 5.24: Relative position of the onset of knock to the pressure reversal point as a function of the MFB50 position for all liquid fuels at EOP1, cycle-based

Figure 5.25 and 5.26 show the relative position in degrees crank angle of the second heat release peak to the pressure reversal point in a similar illustration for the gaseous and liquid fuels. According to the ordinate definition here, the PRP and the 2nd heat release peak of a knocking working cycle are coincident at a value of zero. The distinct amplitude peaks are within a narrow difference range from 0 ◦ CA to 3 ◦ CA. At the time of the PRP, the reverse flowing cooling unburned mixture components could be responsible for lower knock intensities. Conversely, this means that the onset of knock (or the necessary pre-reactions of auto-ignition) has to be initiated before the PRP is reached in order to enable a distinct second heat release peak. The effect of storage and removal of unburned hydrocarbons depends on the volume of the top land piston area and the piston crown design. Since the knocking combustion is not consuming the mixture completely in the crevice and top land areas as the

5.2 Investigation of Atypical Knocking Combustion Phenomena

105

pressure gradient decreases the displaced mixture of the top land area could be converted in the course of a gradual burnout.

15

Amplitude ratio < 0.7 Amplitude ratio > 0.7 knocking IWC

10 5 0 -5 -10 -15 -8

-6

-4 -2 0 2 4 6 8 10 12 14 16 50 % Mass fraction burned in ° CA a. TDC [MFB50]

18

Figure 5.25: Relative position of the 2nd HR-peak to the pressure reversal point as a function of the MFB50 position for all gaseous fuels at EOP1, cycle-based

Finally, it can be summarized that the investigated atypical knocking combustion phenomena, as well as knock itself, depends on complex, mutually influencing parameters. These investigations have shown that the atypical knocking combustion phenomena observed here with gaseous fuels could also be observed with liquid fuels of comparable knock resistance, early MFB50 positions and the same engine boundary conditions.

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15

Amplitude ratio < 0.7 Amplitude ratio > 0.7 knocking IWC

10 5 0 -5 -10 -15 -8

-6

-4 -2 0 2 4 6 8 10 12 14 16 50 % Mass fraction burned in ° CA a. TDC [MFB50]

18

Figure 5.26: Relative position of the 2nd HR-peak to the pressure reversal point as a function of the MFB50 position for all liquid fuels at EOP1, cycle-based

5.3 Motor Methane and Octane Number Correlation The knock resistance of fuels is determined in experimental investigation on a special engine designed for this purpose, as explained in chapter 2.6, using standardized methods, with the result of a classification number [2]. The methods differ for gaseous and liquid fuels as they are carried out under modified conditions and with different primary reference fuel components. These methods were developed in the first part of the last century. Therefore, the key figures can only approximately indicate the anti-knock properties of the fuel for modern highly supercharged combustion engines. The original determina-

5.3 Motor Methane and Octane Number Correlation

107

tion of the classification number was carried out in a CFR engine by increasing the compression ratio continuously until a previously defined knock limit was achieved. The amount of methane or isooctane contained in the primary reference fuel is by definition the MN, respectively RON. These procedures are still the best-known method of rating the knock resistance of a fuel. There is no correlation between the MN of gaseous fuels and the RON of liquid fuels known in literature for modern supercharged combustion engines. Based on the experimental investigations of gaseous and liquid fuels in the present project, a correlation between the MN and the RON is derived. Since the results of the isooctane TEL mixture experiments in chapter 5.1.2 cannot be used for direct comparison with the gaseous fuels, the correlation is derived by the MMN. Therefore, methane-hydrogen mixtures, which are known to correspond to the components of the primary reference fuels in the determination method of MN were investigated. In this work, the reference to knock resistance of the liquid fuels can be established using the MMN relation. In parallel, the deviation between the actual knock resistance and the calculated MN according to the AVL method for the methane-ethane mixtures can be evaluated. For this purpose, the relation of the MFB505 % of a test fuel to the corresponding composition of the CH4 / H2 (primary reference fuel) of equivalent knocking behavior is determined. These MFB505 % are the interpolation points of the MMN correlation curve.

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10

0 % H2 1 % H2 2 % H2 3 % H2 4 % H2 6 % H2 8 % H2 10 % H2 20 % H2 30 % H2 40 % H2 50 % H2

5

0

4

5

6 7 8 9 10 11 12 13 14 15 50 % Mass fraction burned in ° CA a. TDC [MFB50]

16

Figure 5.27: Methane-hydrogen primary reference fuels EOP1: n = 2000 rpm, MN = 50 – 100

5.27 shows the test results of all methane-hydrogen mixtures as a reference mixture variation, starting from pure methane going up to a hydrogen content of 50 %, at the engine speed of 2000 rpm. A plausible knock resistance behavior of the mixtures is shown in the order of the result curves relative to each other. The results of other engine speeds are given in appendix A3.1. According to the procedure described in chapter 5.1 all MFB505 % positions are determined and plotted against the hydrogen content of the mixture in figure 5.28. The knock resistance decreases as the proportion of hydrogen increases, which results in a quasi-linear course of the interpolation points over all investigated mixtures. Similar behavior of the knock resistance with increasing admixing components could also be observed in the literature [43] [34]. In the range of the smallest hydrogen admixtures, the MMN regression line overestimates the measured points. In this range, the gas mixtures are more knock-resistant than the calculated methane number indicates. In order to confirm this behavior or

5.3 Motor Methane and Octane Number Correlation

109

to trace it back to measuring deviations within the tolerance range, additional measurements have been carried out for the low hydrogen admixture range in smaller increments at a reduced engine load point.

20 2000 rpm

18 16 14 12 10 8 6 4 2 0 -2

0

10

20

30 40 50 60 70 80 Proportion of H2 in mol % in CH4

90

100

Figure 5.28: Regression-straight: MFB505 % positions as a function of the H2 content in the mixture for methane-hydrogen fuel matrix at EOP1: n = 2000 rpm, MN = 50 – 100

The results in figure 5.29 confirms that the same knock-resistant increasing effect can be observed for small amounts of admixed hydrogen. The knockresistant increasing effect is sensitive and, with few exceptions, detectable for the majority of the measurements. Further results are presented in appendix A.4. Compared to methane, hydrogen has a faster flame propagation speed and at the same time is itself extremely sensitive to knock, and thus by definition has a methane number of zero. Using pure methane with increasing hydrogen admixture quantities in the single-digit percentage scale, the nonlinear effect of the increasing flame speed is disproportionate to the increasing

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knock tendency of the mixture. In the intervals between two grid points along the approximation curve, the MMN values for the test fuels are determined by linear interpolation; extrapolation is performed outside the grid points. Thus, according to the MFB505 % positions for all investigated fuels, the MMNs are determined at the one operating point. Figure 5.30 shows the MMN for liquid fuels as a function of engine speed.

14 2000 rpm 12 10 8 6 4 2 0 -2

0

10

20

30 40 50 60 70 80 Proportion of H2 in mol % in CH4

90

100

Figure 5.29: Regression-straight: MFB505 % positions as a function of the H2 content in the mixture for methane-hydrogen fuel matrix at EOP2: n = 2000 rpm, MN = 50–100

For the fuels benzene and acetone, the relation results in an MMN > 100 according to the extrapolated values of the motor methane number relation line. This shows how much more knock-resistant these fuels behave in comparison to pure methane. It was not possible to consider the three engine speeds for all mixtures, but the results show a higher knock resistance with increasing engine speed for the investigated liquid fuels. This behavior can be attributed

5.3 Motor Methane and Octane Number Correlation

111

to the knock-reducing effect, which is the increasing turbulent flame speed and the lack of time in which self-ignition can occur in the end gas. Additionally, the cooling effect of the latent heat evaporation during the mixture preparation could have a stronger effect with increasing turbulence at higher engine speeds, if it is not completely absorbed at the time inlet closes. In absolute terms, the mixture has less time left for homogeneity preparation within the intake manifold section as the engine speed increases. In figure 5.31, this behavior is not equally apparent for the methane-ethane mixture over engine speed. There are variations in the knock resistance for different speeds, but the deviation to the MMN is smallest for an engine speed of 2000 rpm. In contrast to liquid fuels, the increasing knock resistance with higher engine speeds could be overcompensated by the generally higher temperature level, which can lead to a reduced knock resistance for specific gaseous compositions [14].

160 150 140 130 120 110 100 90 80 70 60 50 40 30 20 10 0

158

1500 rpm 2000 rpm 3000 rpm 114 115 105

84 73 65

62 48

72

68 53

53

52

44

42 33

31

27 15

9

Benzole Toluole Xylole_m Mesitylene DCPD Acetone Xylole_p MTBE P1 Fuel

Figure 5.30: Motor methane number to the corresponding liquid fuel at EOP1: n = 1500 rpm, 2000 rpm and 3000 rpm

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The MMN correlation confirms the knocking behavior of the mixtures and shows the deviation of the calculated methane number versus the actual knocking behavior relative to the bisector line. For all three engine speeds, the deviation increases with decreasingly knock-resistant fuels. The two methaneethane mixtures, which showed pre-ignited working cycles when determining the knock rate for 3000 rpm in chapter 5.1.1, are consequently not considered for the MMN correlation.

0 10 20

1500 rpm 2000 rpm 3000 rpm

30 40 50 60 70 80 90 100 100

90

80

70

60 50 40 30 Methane number [MN]

20

10

0

Figure 5.31: Deviation of the MMN relative to the calculated AVL MN for methane-ethane mixtures at EOP1: n = 1500 rpm, 2000 rpm and 3000 rpm

The actual knock resistance of the liquid fuels can now be linked to the RON via the MMN correlation. However, data from the literature has to be applied, as the direct correlation to the experimental RON measurement of the reference fuels (isooctane TEL) was not possible, as already mentioned. The data in the literature vary considerably for the same fuel over a wide range, see table 3.4

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113

on page 51. The determination of the RON for a highly knock-resistant fuels under identical conditions as the original procedure according to the ASTM method can hardly be carried out with TEL reference fuels, since TEL is not available anymore for these examinations. 130 Liquid Fuels 1500 rpm 2000 rpm 3000 rpm

120 110 100 90 80 70 60 50 40 30 20 115

116

117

118 119 120 121 122 123 Research octane number [RON]

124

125

Figure 5.32: Methane number according to the RON values found in the literature. Evaluated as a function of MMN

Figure 5.32 shows the correlation for the calculated AVL methane numbers and the RON values of the literature. The correlation is limited to the range in which sufficient fuels with reliable RON could be investigated, i.e. benzene, acetone and dicyclopentadiene could not be considered. The data of the knock resistance of a fuel based on index numbers, as determined by the RON and MON methods, show deviations from the expected knocking behavior in the test engine. Investigations in literature show similar findings regarding these discrepancies, which can vary considerably for different engine concepts [47]. Due to the limited validity of the RON for modern engines compared

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to the CFR engine, corresponding deviations between the determined MN and the RON literature values occur. Therefore, the anti-knock performance of liquid fuels over the entire range of the experimental measurements is shown in 5.30 using the correlation of the MMN. Using the motor methane number correlation, it was possible to classify the investigated fuels with regard to their knock resistance and to compare them across the different chemical states. Further investigations with the primary reference fuels according to the original determination procedure might therefore be useful here as well to gain a detailed correlation between the MN and the RON. Apart from that, the deviation between the actual knock resistance of the liquid fuels in a modern engine and the knock resistance according to the RON determined in the CFR engine could be investigated.

6 Conclusion and Outlook This investigation at hand is based on the observations regarding specific heat release phenomena associated with knocking combustion, which were observed during the thermodynamic evaluation in a previous project investigating the knocking behavior of methane-based fuels. These phenomena have been classified into three atypical knocking combustion types. With regard to the objective of this research project, a comprehensive fuel matrix of liquid and gaseous fuels was established which is representative for sustainable future fuels according to their fuel-specific properties. Based on systematic measurements, fuels of different compositions were examined with regard to their knock resistance and knocking behavior in a modern gasoline engine. The comparison of chemically different fuels at similarly early MFB50 positions at the knock limit shows the cause and effects of these phenomena. The results presented in this work have been completed with a correlation between gaseous and liquid fuels which have shown an overlapping knock resistance range. In summary, the following aspects of the investigations are particularly interesting: Onset of Knocking Combustion The first of the mentioned atypical heat release phenomena is an unusually early onset of knock observed in both, the previous and this project as well. The general understanding of incipient knocking with gradual early adjustment of the ignition timing from the non-knocking operating range describes late onset of knock for gasoline-like fuels with low knocking intensities at sporadic working cycles and thus initially low knocking rates. In the present investigations, comparative measurements with liquid fuels at MFB50 positions close to TDC have been carried out with regard to the onset of knock. In the evaluation, the thermodynamically relevant parameters regarding the onset of knock and knocking intensity as a function of the MFB50 and MFB position were examined. The results of the comparison of gaseous and liquid fuels show that these early onsets of knock also occur for knock-resistant liquid fuels if their combustion position could be set in ranges close to TDC, as in the case of © The Author(s), under exclusive license to Springer Fachmedien Wiesbaden GmbH, part of Springer Nature 2021 M. Eberbach, Knocking Combustion of Methane-Based and Highly Knock Resistant Liquid Fuels, Wissenschaftliche Reihe Fahrzeugtechnik Universität Stuttgart, https://doi.org/10.1007/978-3-658-35178-6_6

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the gaseous fuels. Based on the information from available literature, it was possible to show a correlation between the thermodynamic end gas history and the influence of engine-specific current effects in relation to the pressure reversal point. The investigations of the liquid fuels show that the observed effect within the varied fuel-specific properties is not solely due to the chemistry of the fuels, but is additionally superimposed by reaction kinetic and/or physically dominated effects. Further investigations, for example with different piston crowns geometries and thermodynamic evaluations regarding the HRR at the end of the working cycle could be useful to clarify these correlations further. Distinct 2nd Heat Release Peak For both fuel types, the root cause of the second atypical combustion phenomenon i.e. a gradually increasing amplitude peak on the heat release trace was investigated, beginning in the non-knocking operating range and ending at the knock limit. By comparing liquid and gaseous fuels, a specific approach was used to show the extent to which this so-called primary knocking effect on the heat release trace differs for the liquid fuels by attributing the knocking working cycles to the algorithm that identified them. It was found that, in contrast to the gaseous fuels, the knocking working cycles for liquid fuels are mainly identified by the secondary knocking combustion effect, i.e. the typical oscillation amplitudes on the pressure trace provoked by the very fast conversion of the end gas and less by the distinct 2nd heat release peak (primary knocking effect) caused by a second moderately fast propagating self-ignited flame front Non-Oscillating Knocking Working Cycles With regard to the third investigated atypical knocking combustion phenomenon, the statistical evaluation of the proportion of non-oscillating knocking working cycles relative to the total number of identified cycles showed the development as a function of different variation parameters. In the pressure trace analysis, the non-oscillating knocking working cycles clearly show the primary knocking effect in the development of the second heat release peak on the heat release trace but without the typical oscillation on the pressure trace. Since conventional knock algorithms are based on the analysis of the latter, i.e. secondary knocking effect, such knocking working cycles are not detected.

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In the evaluation of these self-ignited, but not conventionally knocking working cycles, the liquid fuels appeared throughout to be relatively insensitive regarding this phenomenon, whereas gaseous fuels show tendencies over engine speeds and even more significantly over the knock resistance of the fuel to be more sensitive regarding this phenomenon. The occurrence of this possibly gas-specific knocking characteristic was compared with the individual findings of similar behaviors in the available literature and the underlying mechanisms were related to the measurements. This leads to the assumption that the reaction kinetic processes in the end gas could differ between the conversion rate of liquid and gaseous fuels to such an extent that in the transition area from barely knocking and very strongly knocking combustion a kind of slow volume reaction takes place and therefore no oscillations are caused on the heat release rate. Due to the limited availability of very knock-resistant liquid fuels and the associated health hazards, the fuel variation carried out in the present investigations comprises a restricted number of fuels from individual chemical groups. Further investigations by varying the fuel-specific properties and, ideally, additional optical measurement technology could contribute to a profound understanding with regard to the type of flame propagation and energy conversion in the end gas pockets. Motor Methane Number – Octane Number Correlation On the basis of the MFB505 % positions determined in section 5.1 for gaseous and liquid fuels of different knock resistance at the knock limit, a correlation between the knock resistance indices of both types of fuel was derived for a supercharged modern combustion engine. The motor methane number correlation was used to classify the fuels investigated in terms of their theoretical knock resistance, i.e. its MN respectively RON, and to compare them with each other across the different chemical states of the fuels (liquid and gaseous). While for the gaseous fuels, differences between the actual knock resistance and the calculated methane number could be shown, for the liquid fuels, the RON data from the literature was correlated with the knock resistance behavior from the experimental investigations. As table 3.4 on page 51 shows, the RON specifications for the same fuel vary considerably in the literature. Further investigations with the primary reference fuels according to the original determination procedure might therefore be useful here as well to gain a detailed correlation between the MN and the RON. Apart from that, the

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deviation between the actual knock resistance of the liquid fuels in a modern engine and the knock resistance according to the RON determined in the CFR engine could be investigated. The evaluation of the measurement results shows to what extent the atypical knocking combustion phenomena of gaseous fuels occur at the same MFB50 positions for the extremely knock-resistant test fuels investigated. Since conventional gasoline engine fuels could not be investigated at all under similar boundary conditions, the investigations and the fuel-specific correlations determined at the IFS of the University of Stuttgart provide an essential contribution to the knocking behavior of the fuels, whose specific properties on a chemical basis might be relevant for future synthetic fuels.

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Appendix A.1 Appendix 1 Appendix A.1.1 shows the results of the preliminary tests for determining the engine operating point. The figures show the knock resistant as a function of inlet air temperature, boost pressure and engine speed. Appendix A1.2 shows the investigations on liquid fuel injection; the knock resistance as a function of fuel temperature, injection pressure and injection timing.

© The Editor(s) (if applicable) and The Author(s), under exclusive license to Springer Fachmedien Wiesbaden GmbH, part of Springer Nature 2021 M. Eberbach, Knocking Combustion of Methane-Based and Highly Knock Resistant Liquid Fuels, Wissenschaftliche Reihe Fahrzeugtechnik Universität Stuttgart, https://doi.org/10.1007/978-3-658-35178-6

128

Appendix 1

10 p2 = 1.2 p2 = 1.4 p2 = 1.6 p2 = 1.8 p2 = 2.0 5

0 -4 -3 -2 -1 0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 50 % Mass fraction burned in ° CA a. TDC [MFB50]

10 T2 = 25 °C T2 = 45 °C T2 = 65 °C

5

0 -3 -2 -1 0 1 2 3 4 5 6 7 8 9 10 11 12 50 % Mass fraction burned in ° CA a. TDC [MFB50]

Figure A1.1: Influence of inlet air pressure and temperature variation on the knocking rate as a function of MFB50 position: EOP1; methane

Appendix 1

129

10 Original 5 bar HDEV1 120 bar, T Krst = Tamb. Injection timing opt. HDEV1 120 bar, T Krst = 70 °C

5

0

7

8 9 10 11 12 13 14 15 16 50 % Mass fraction burned in ° CA a. TDC [MFB50]

17

10 Original 5 bar HDEV1 200 bar HDEV1 120 bar HDEV1 120 bar HDEV1 80 bar HDEV1 40 bar 5

0

7

8 9 10 11 12 13 14 15 16 50 % Mass fraction burned in ° CA a. TDC [MFB50]

17

Figure A1.2: Fuel temperature and rail pressure variation for toluene: EOP1, n = 1500 rpm

130

Appendix 1

10 n = 1500 rpm n = 2000 rpm n = 2500 rpm n = 3000 rpm n = 3500 rpm 5

0

5

6 7 8 9 10 11 50 % Mass fraction burned in ° CA a. TDC [MFB50]

12

Figure A1.3: Knocking rate vs. the MFB50 position as a function of engine speeds: EOP1, methane

Appendix 1

131

10 = 0.8 = 0.9 = 1.0 = 1.1 = 1.2 = 1.3 = 1.4

5

0

2

3 4 5 6 7 8 9 10 11 50 % Mass fraction burned in ° CA a. TDC [MFB50]

12

10 = 0.8 = 0.9 = 1.0 = 1.1 = 1.2 = 1.3 = 1.4

5

0 -4

-2

0 2 4 6 8 10 12 14 16 18 50 % Mass fraction burned in ° CA a. TDC [MFB50]

20

Figure A1.4: Knock rate vs. the MFB50 position as a function of λ = 0.8 to 1.4: EOP1, methane (top) and toluene (bottum)

132

Appendix 2

A.2 Appendix 2 Appendix A2 shows the results of the knock investigations for TEL-enriched isooctane fuels, albeit at an adjusted lower operating point.

A2.1 Knock Rate versus MFB50

10

5

MN = 100.0 MN = 94.71 MN = 90.07 MN = 84.95 MN = 79.47 MN = 74.80 MN = 69.89 MN = 64.54 MN = 59.14 MN = 54.11 MN = 48.68 MN = 44.38

0 -3 -2 -1 TDC 1 2 3 4 5 6 7 8 9 10 11 12 13 50 % Mass fraction burned in ° CA a. TDC [MFB50]

Figure A2.1: Methane-ethane mixtures: n = 2000 rpm, IMEP = 14 bar to 15 bar

Appendix 2

133

10 RON = 125 RON = 120 RON = 115 RON = 110 RON = 105 RON = 100 (Isooctane) 5

0 -4

TDC 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 50 % Mass fraction burned in ° CA a. TDC [MFB50]

10 RON = 125 RON = 120 RON = 115 RON = 110 RON = 105 RON = 100 (Isooctane) 5

0 -4

TDC 4 6 8 10 12 14 16 18 20 22 24 26 28 30 32 34 36 38 40 50 % Mass fraction burned in ° CA a. TDC [MFB50]

Figure A2.2: Isooctane-TEL mixture: n = 2000 rpm (top) and 3000 rpm (bottom), p2 = 1000 mbar, T2 = 45 ◦C RON = 100 – 125

134

Appendix 2

10 Methane Acetone

5

0

4

350 300 250

5

6 7 8 9 10 11 12 13 14 15 16 50 % Mass fraction burned in ° CA a. TDC [MFB50]

Methane Acetone Cyl. pressure Heat-release

17

140 120 100

200

80

150

60

100

40

50

20

0

0

-50 -40 -30 -20 -10 0 10 20 30 40 50 60 70 Crank angle position in °CA a. TDC [ ]

Figure A2.3: Liquid (acetone) and gaseous (methane) fuel with similar MFB505 % position; n = 1500 rpm, heat release (3 kHz-filtered)

Appendix 2

250 200

135

Liquid fuel (Toluene) Methane based fuel (MN = 54) Methane based fuel (MN = 59)

150 100 50 0 -40

-20 0 20 40 60 Crank angle position in °CA a. TDC [ ]

Figure A2.4: Comparison of a liquid fuel and gaseous fuels: n = 1500 rpm, filtered HRR (3 kHz-filtered)

80

EOP1,

136

Appendix 3

A.3 Appendix 3 Appendix A3 contains the results of the methane-hydrogen mixtures which were used for the motor methane number correlation. The figures show the development of the knock rate with increasing hydrogen content in the methanehydrogen mixture. Results are available for different speeds and load points or engine operating points.

Appendix 3

137

10

0 % H2 1 % H2 2 % H2 3 % H2 4 % H2 6 % H2 8 % H2 10 % H2 20 % H2 30 % H2 40 % H2 50 % H2

5

0

4

5

6 7 8 9 10 11 12 13 14 15 50 % Mass fraction burned in ° CA a. TDC [MFB50]

16

10

0 % H2 1 % H2 2 % H2 3 % H2 4 % H2 6 % H2 8 % H2 10 % H2 20 % H2 30 % H2 40 % H2 50 % H2

5

0

4

5

6 7 8 9 10 11 12 13 14 15 50 % Mass fraction burned in ° CA a. TDC [MFB50]

16

Figure A3.1: Methane-hydrogen primary reference fuels: EOP1, n = 2000 rpm (top) and 3000 rpm (bottom), MN = 40 – 100

138

Appendix 3

10

0 % H2 1 % H2 2 % H2 3 % H2 4 % H2 6 % H2 8 % H2 10 % H2 20 % H2 30 % H2 40 % H2 50 % H2 60 % H2

5

0

-3

-2

-1 0 1 2 3 4 5 6 7 8 50 % Mass fraction burned in ° CA a. TDC [MFB50]

9

10

0 % H2 1 % H2 2 % H2 3 % H2 4 % H2 6 % H2 8 % H2 10 % H2 20 % H2 30 % H2 40 % H2 50 % H2 60 % H2

5

0

-1

0

1 2 3 4 5 6 7 8 9 10 50 % Mass fraction burned in ° CA a. TDC [MFB50]

11

Figure A3.2: Knock rate vs. MFB50 position as a function of content H2 in the methane-hydrogen mixture 1 % to 8 % for EOP2, n = 1500 rpm (top), 2000 rpm (bottom)

Appendix 3

139

10

0 % H2 1 % H2 2 % H2 3 % H2 4 % H2 6 % H2 8 % H2 10 % H2 20 % H2 30 % H2 40 % H2 60 % H2

5

0

0

1

2 3 4 5 6 7 8 9 10 11 50 % Mass fraction burned in ° CA a. TDC [MFB50]

12

Figure A3.3: Knock rate vs. MFB50 position as a function of content H2 in the methane-hydrogen mixture 1 % to 8 % for EOP2, n = 3000 rpm

140

Appendix 4

A.4 Appendix 4 Appendix A4 contains a compilation of the regression lines for different methane-hydrogen mixtures at different engine speeds and load points. A regression line shows the development of the knock rate with increasing hydrogen content of the investigated mixtures for an operating point and is an intermediate step in the determination of the motor methane number.

Appendix 4

141

20 1500 rpm

18 16 14 12 10 8 6 4 2 0 -2

0

10

20

30 40 50 60 70 80 Proportion of H2 in mol % in CH4

90

100

20 3000 rpm

18 16 14 12 10 8 6 4 2 0 -2

0

10

20

30 40 50 60 70 80 Proportion of H2 in mol % in CH4

90

100

Figure A4.1: Regression-straight: MFB505 % positions vs. the H2 content of methane-based mixture at EOP1: n = 1500 rpm (top) and 3000 rpm (bottom), MN = 50 – 100

142

Appendix 4

14 1500 rpm 12 10 8 6 4 2 0 -2

0

10

20

30 40 50 60 70 80 Proportion of H2 in mol % in CH4

90

100

14 3000 rpm 12 10 8 6 4 2 0 -2

0

10

20

30 40 50 60 70 80 Proportion of H2 in mol % in CH4

90

100

Figure A4.2: Regression-straight: MFB505 % positions vs. the H2 content of methane-based mixture at EOP2: n = 1500 rpm (top) and 3000 rpm (bottom), MN = 60 – 100

Appendix 4

143

A4.1 Methane-Ethane MMN – MN Correlation

0 10 20

1500 rpm 2000 rpm 3000 rpm

30 40 50 60 70 80 90 100 100

90

80

70

60 50 40 30 Methane number [MN]

20

10

0

Figure A4.3: Deviation of the MMN relative to the calculated AVL MN for methane-ethane mixtures at EOP2: n = 1500 rpm, 2000 rpm and 3000 rpm