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The effects of turbulence and combustion chamber geometry on combustion in a spark ignition engine Mawle, Craig D. 1989

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T H E E F F E C T S O F T U R B U L E N C E A N D C O M B U S T I O N C H A M B E R G E O M E T R Y O N C O M B U S T I O N I N A S P A R K I G N I T I O N E N G I N E By Craig D. Mawle B.A.Sc. in Mechanical Engineering, University of British Columbia, 1986 A THESIS SUBMITTED IN PARTIAL F U L F I L L M E N T OF T H E REQUIREMENTS FOR T H E D E G R E E OF M A S T E R OF A P P L I E D SCIENCE in T H E FACULTY OF GRADUATE STUDIES D E P A R T M E N T OF MECHANICAL ENGINEERING We accept this thesis as conforming to the required standard T H E UNIVERSITY OF BRITISH COLUMBIA September, 1989 © Craig D. Mawle, 1989 In presenting this thesis in partial fulfilment of the requirements for an advanced degree at the University of British Columbia, I agree that the Library shall make it freely available for reference and study. I further agree that permission for extensive copying of this thesis for scholarly purposes may be granted by the head of my department or by his or her representatives. It is understood that copying or publication of this thesis for financial gain shall not be allowed without my written permission. Department of Mechanical Engineering The University of British Columbia 2075 Wesbrook Place Vancouver, Canada V6T 1W5 Date: Octobe r 10, 1989 Abstract An experimental program has been undertaken to investigate the effects of turbulence on combustion in a spark ignition engine and to examine the effects of a new combustion chamber design. The experiments were conducted in a rapid intake and compression machine using the compression stroke alone in order to eliminate the effects of intake stroke generated turbulence. The effects of turbulence scale and intensity on combustion was investigated using perforated plates to generate turbulence in the combustion chamber. The turbulence generated with three different perforated plate hole sizes was measured with a hot-wire anemometer (HWA) at several locations and orientations so that the turbulence could be characterized. Combustion experiments were run for various ignition timings with a premixed stoichiometric methane/air mixture. Pressure measurements and a mass fraction burned (MFB) analysis program were used to calculate the MFB curves and the combustion durations. The turbulence generated by the perforated plates was separated into two stages. The first stage (pre-relaxation stage) was characterized by large anisotropics with a turbulence scale characteristic of the perforated plate hole size. The second stage (relaxed stage) was characterized by homogeneous isotropic turbulence with a turbulence scale characteristic of the chamber height. Although the scales were not measured they were implied by the decay rates and the turbulence intensity. The main combustion duration (5-90% MFB) was found to decrease with increased turbulence intensity and the flame initiation period (FIP, 0-5% MFB) was found to decrease with smaller scales even when counteracted by a lower turbulence intensity. 11 This suggests that the generation of small scale turbulence could be used to decrease the FIP in a spark ignition engine. The compression stroke turbulence generating characteristics of a standard bowl-in-piston design and several new "forced squish-jet" designs were examined using HWA measurements. The "forced squish-jet" designs used a ridge at the top of the piston bowl and a step in the head to force the squish motion through jet slots located in the ridge. Combustion durations and MFB curves were determined from combustion experiments for all the designs as with the perforated plates. The flow field generated by the bowl-in-piston design resulted in the turbulence in-tensity at the centre of the bowl being generated at around top dead centre (TDC). With the "forced squish-jet" designs, although the main part of the motion was directed between the ridge and the step in the head and not through the jet slots, the turbulence intensity at the centre of the bowl was generated around 10 degrees before TDC. The jet slots main influence was to reduce the effectiveness of the ridge in generating turbulence. The combustion with the different combustion chamber designs was significantly af-fected by the generation of turbulence. The fuel/air mixture burned at a slow (laminar) rate until the pistons generated turbulence upon which the combustion rate dramatically increased. The timing of the increase in the combustion rate was determined by the timing of the turbulence generation with the different designs. hi Table of Contents Abstract ii List of Tables vii List of Figures viii Nomenclature xiii Acknowledgments xvi 1 Background and Review 1 1.1 Introduction 1 1.2 Mixture Motion in an Internal Combustion Engine 4 1.3 Objective and Scope 5 1.4 Review of Previous Work 7 2 Experimental Apparatus and Method 15 2.1 Rapid Intake and Compression Machine 15 2.1.1 Existing RICM Apparatus 15 2.1.2 Modifications to the RICM 16 2.2 Data Acquisition 18 2.3 Flow Field Measurements 19 2.4 Combustion Measurements 21 3 Data Analysis 24 IV 3.1 Flow Field Data Analysis 24 3.1.1 Hot-Wire Anemometer Data Analysis 24 3.1.2 Limitations of the HWA in the RICM 29 3.2 Combustion Data Analysis 31 4 Discussion of Experimental Results 34 4.1 Effects of the Perforated Plates 34 4.1.1 Characterization of Turbulence 34 4.1.2 Effects on Combustion 38 4.2 Effects of Different Combustion Chamber Designs 40 4.2.1 Compression stroke Generated Turbulence 41 4.2.2 Combustion with Compression Stroke Generated Turbulence . . . 49 5 Conclusions and Recommendations 55 5.1 Introduction 55 5.2 Conclusions 55 5.2.1 Perforated Plate 55 5.2.2 Different Combustion Chamber Designs 56 5.3 Recommendations 58 5.3.1 Perforated Plate 58 5.3.2 Different Combustion Chamber Designs 59 References 60 Appendices 115 A Description of M F B Computer Program 115 v B Calculations of Squish Velocity 121 vi List of Tables 2.1 Combustion Chamber Specifications and RICM Running Parameters . . 64 2.2 Hot-Wire Anemometer Specifications and Equipment 64 2.3 HWA Measurement Locations Used with the Different Pistons 65 2.4 Number of Successful Combustion Runs with Different Pistons 65 4.1 Combustion Durations with Perforated Plates 66 4.2 Combustion Durations with Piston BIP 67 4.3 Sum of Time to Spark Plus Combustion Duration with Piston BIP . . . 68 4.4 Combustion Durations with Piston 2J 69 4.5 Sum of Time to Spark Plus Combustion Duration with Piston 23 . . . . 70 4.6 Combustion Durations with Piston 2LJ 71 4.7 Combustion Durations with Piston 4J 72 4.8 Combustion Durations with Piston NJ 73 B. l Piston Position, Piston Velocity and Squish Velocity for the standard bowl-in-piston design with 1 mm clearance gap 124 vii List of Figures 2.1 Schematic of Rapid Intake and Compression Machine 74 2.2 Schematic of RICM Intake Valve and Linkage Mechanism 75 2.3 "Forced Squish-Jet" Piston Crowns for the RICM 76 2.4 Cross-Section of the "Forced Squish-Jet" Head Assembly 77 3.1 Ensamble-Averaged Mean Velocity and Turbulence Intensity for N = 2, 5, 10 and 15 Cycles 78 3.2 Comparisons of Ignition Delay (FIP) and Combustion Duration Computed with Krieger-Borman and Rassweiler-Withrow Analysis on Identical Pres-sure Files 79 3.3 Comparisons of Estimates of Apparent MFB to Results from Comprehen-sive Computer Model 79 4.1 Mean Velocity and Turbulence Intensity measured at the centre and near the edge of the combustion chamber lined up with a hole in the 20 mm Perforated Plate 80 4.2 Mean Velocity and Turbulence Intensity measured at the centre and near the edge of the combustion chamber lined up with a hole in the 10 mm Perforated Plate 81 4.3 Mean Velocity and Turbulence Intensity measured at the centre and near the edge of the combustion chamber lined up with a hole in the 5 mm Perforated Plate 82 viii 4.4 Mean Velocity and Turbulence Intensity measured at the centre of a hole and between two holes with the wire parallel and perpendicular to the 20 mm Perforated Plate 83 4.5 Mean Velocity and Turbulence Intensity measured at the centre of a hole and between two holes with the wire parallel and perpendicular to the 10 mm Perforated Plate 84 4.6 Mean Velocity and Turbulence Intensity measured at the centre of a hole and between two holes with the wire parallel and perpendicular to the 5 mm Perforated Plate 85 4.7 Mean Velocity and Turbulence Intensity measured at the centre of a hole with the wire parallel to the 20, 10 and 5 mm Perforated Plates 86 4.8 Log-Log Plot of Mean Velocity and Turbulence Intensity measured at the centre of a hole with the wire parallel to the 20, 10 and 5 mm Perforated Plates 87 4.9 MFB Curves with ignition timing at TDC as well as 10, 20, 30 and 40 degrees BTDC for the 20 mm Perforated Plate 88 4.10 MFB Curves with ignition timing at TDC as well as 10, 20, 30 and 40 degrees BTDC for the 10 mm Perforated Plate 88 4.11 MFB Curves with ignition timing at TDC as well as 10, 20, 30 and 40 degrees BTDC for the 5 mm Perforated Plate 89 4.12 MFB Curves with ignition timing at TDC for the 20, 10 and 5 mm Perfo-rated Plates 89 4.13 MFB Curves with ignition timing at 10 degrees BTDC for the 20, 10 and 5 mm Perforated Plates 90 4.14 MFB Curves with ignition timing at 20 degrees BTDC for the 20, 10 and 5 mm Perforated Plates 90 IX 4.15 MFB Curves with ignition timing at 30 degrees BTDC for the 20, 10 and 5 mm Perforated Plates 91 4.16 MFB Curves with ignition timing at 40 degrees BTDC for the 20, 10 and 5 mm Perforated Plates 91 4.17 Mean Velocity and Turbulence Intensity measured at the 2, 8 and 16 mm depths near the bowl edge for Piston BIP with respect to crank angle . . 92 4.18 Mean Velocity and Turbulence Intensity measured at the 2, 8 and 16 mm depths near the bowl edge for Piston BIP with respect to time 93 4.19 Mean Velocity and Turbulence Intensity measured at the 2, 8 and 16 mm depths at the centre of the bowl for Piston BIP with respect to crank angle 94 4.20 Mean Velocity and Turbulence Intensity measured at the 2, 8 and 16 mm depths at the centre of the bowl for Piston BIP with respect to time . . . 95 4.21 Sketch of Flow Pattern Generated with the Bowl-in-Piston Design . . . . 96 4.22 Mean Velocity and Turbulence Intensity measured at the 2, 8 and 16 mm depths near the bowl edge lined up with a jet slot for Piston 2J 97 4.23 Mean Velocity and Turbulence Intensity measured at the 2, 8 and 16 mm depths near the bowl edge between the jet slots for Piston 2J 98 4.24 Sketch of Flow Pattern Generated with the "Forced Squish-Jet" Designs 99 4.25 Mean Velocity and Turbulence Intensity measured at the 2, 8 and 16 mm depths at the centre of the bowl for Piston 2J 100 4.26 Mean Velocity and Turbulence Intensity measured at the 8 mm depth near the bowl edge lined up with a jet slot for Pistons 2J, 2LJ and 4J 101 4.27 Mean Velocity and Turbulence Intensity measured at the 2 mm depth near the bowl edge away from jet slots for Pistons BIP, 2J, 4J and NJ 102 4.28 Mean Velocity and Turbulence Intensity measured at the 8 mm depth near the bowl edge away from jet slots for Pistons BIP, 2J, 2LJ, 4J and NJ . . 103 x 4.29 Mean Velocity and Turbulence Intensity measured at the 2 mm depth at the centre of the bowl for Pistons BIP, 2J, 2LJ, 4J and NJ 104 4.30 Mean Velocity and Turbulence Intensity measured at the 8 mm depth at the centre of the bowl for Pistons BIP, 2J, 2LJ, 4J and NJ 105 4.31 Mean Velocity and Turbulence Intensity measured at the 16 mm depth at the centre of the bowl for Pistons BIP, 23, 2LJ, 4J and NJ 106 4.32 MFB Curves with ignition timing at 20 degrees BTDC for a spark probe depth of 2, 8 and 16 mm for Piston BIP 107 4.33 MFB Curves with ignition timing at 30 degrees BTDC for a spark probe depth of 2, 8 and 16 mm for Piston BIP 107 4.34 MFB Curves with ignition timing at 40 degrees BTDC for a spark probe depth of 2, 8 and 16 mm for Piston BIP 108 4.35 MFB Curves with ignition timing at 20 degrees BTDC for a spark probe depth of 2, 8 and 16 mm for Piston 2J 108 4.36 MFB Curves with ignition timing at 30 degrees BTDC for a spark probe depth of 2, 8 and 16 mm for Piston 2J 109 4.37 MFB Curves with ignition timing at 40 degrees BTDC for a spark probe depth of 2, 8 and 16 mm for Piston 2J 109 4.38 MFB Curves with ignition timing at 40 degrees BTDC for a spark probe depth of 2 mm for Pistons BIP, 2J, 2LJ, 4J and NJ 110 4.39 MFB Curves with ignition timing at 40 degrees BTDC for a spark probe depth of 8 mm for Pistons BIP, 2J, 2LJ, 4J and NJ 110 4.40 MFB Curves with ignition timing at 40 degrees BTDC for a spark probe depth of 16 mm for Pistons BIP, 2J, 2LJ, 4J and NJ I l l 4.41 MFB Curves with ignition timing at 30 degrees BTDC for a spark probe depth of 2 mm for Pistons BIP, 2J, 2LJ and NJ I l l xi 4.42 MFB Curves with ignition timing at 30 degrees BTDC for a spark probe depth of 8 mm for Pistons BIP, 2J and 2LJ 112 4.43 MFB Curves with ignition timing at 30 degrees BTDC for a spark probe depth of 16 mm for Pistons BIP, 2J and 2LJ 112 4.44 MFB Curves with ignition timing at 20 degrees BTDC for a spark probe depth of 2 mm for Pistons BIP and 2J 113 4.45 MFB Curves with ignition timing at 20 degrees BTDC for a spark probe depth of 8 mm for Pistons BIP and 2J 113 4.46 MFB Curves with ignition timing at 20 degrees BTDC for a spark probe depth of 16 mm for Pistons BIP and 2J 114 A. l Comparison between the MFB curves calculated with a time step of 0.016 and 0.16 milliseconds 117 A. 2 Comparison between the MFB curves calculated using a polytropic coef-ficient of compression of 1.25, 1.30, 1.35 and 1.40 117 B. l Squish Velocity and Piston Velocity based on a simple model for clearance heights of 1, 2 and 5 mm for the standard bowl-in-piston design at 1000 rpm 125 Xll Nomenclature a Thermal coefficient of resistance BTDC Before Top Dead Centre FIP Flame Initiation Period (0-5% MFB) i Cycle index At Micro time scale Xx Micro length scale Lt Integral time scale Lx Integral length scale MFB Mass Fraction Burned N Number of cycles Nu Nusselt number $ Equivalence ratio ( , \ T , a t i o ) * * stoichiometric fuel/air ratio ' APcomtmttion Pressure change due to combustion APpi,ton(caicuiated) Calculated pressure change due to piston motion APtotai(mea,ured) Total measured pressure change &Pvintt.rval Pressure change at constant volume Vi n t e r „ 0 j APvTDC Pressure change at constant volume VTDC r Spacial separation R(r) Spacial correlation coefficient R(T) Autocorrelation coefficient R a m b Resistance at ambient temperature Rop Resistance at operating temperature Xll l Re Reynolds number Re\ Microscale Reynolds number t Time T Window size r Correlation time 6 Crank angle TDC Top Dead Centre tg Time at crank angle 6 Tom{, Ambient temperature Operating temperature tg Time at crank angle 9 tw Time at centre of window U(i,t) Velocity in cycle i at time t U(i,tw) Window-averaged mean velocity in cycle i at time tw U(i,t) Mean velocity in cycle i at time t u(i,t) Turbulent velocity fluctuation in cycle i at time t u'(i,tw) Window-averaged turbulence intensity in cycle i at time tw u'(i, t) Turbulence intensity in cycle i at time t UE(^) Ensamble-averaged mean velocity at time t u'E(t) Ensamble-averaged turbulence intensity at time t Uc(i,t) Cyclic variation of mean velocity in cycle i at time t U'c(t) Cyclic variation intensity at time t U(i, 6) Mean velocity in cycle i at crank angle 6 u'{i,0) Turbulence intensity in cycle i at crank angle 6 Uc(i,Q) Cyclic variation of mean velocity in cycle i at crank angle 6 xiv UE{9) Ensamble-averaged mean velocity at crank angle 9 u'E{9) Ensamble-averaged turbulence intensity at crank angle 6 UQ(9) Cyclic variation intensity at crank angle 9 u(xp) Turbulent velocity fluctuation at point xp u2 Mean square velocity U(T) Turbulent velocity fluctuation at time r ^interval Volume during interval V T D C Volume at TDC xp Representitive coordinate of base point P xv Acknowledgments I would like to express my sincere gratitude to my supervisor Dr. R. L. Evans for his contributions and encouragement during the course of this work. I would also like to thank J. Richards for his work on the ignition system and T. Besic for the machining of the modifications to the apparatus. Further thanks are due Professor I. S. Gartshore and fellow graduate students for their valuable contributions during various discussions. xvi Chapter 1 Background and Review 1.1 Introduction Internal combustion engines have provided a portable power source for the past century and will probably continue to do so for the foreseeable future. Over the past few decades concerns have arisen over the effects of the internal combustion engine on our society and environment. In the seventies the "oil crisis" brought engine efficiency improvements into the forefront of engine researchers minds and more recently environmental effects have become a major concern. The lack of internal combustion engine alternatives suggests that both efficiency and emissions control will continue to be important research topics. Internal combustion engine research has established that advancements in efficiency and exhaust emissions can be produced through fast and lean combustion. Lean mixtures burn cooler and are less susceptible to autoignition or knock so they can be used in higher compression ratio engines [1] which are thermodynamically more efficient [2]. Lean mixtures produce less carbon monoxide, hydrocarbon and nitrogen oxide emissions [1] which are the main pollutants from spark ignition engines. There are drawbacks with lean combustion however, namely that lean mixtures burn more slowly [3] and lean-burn engines are more prone to misfire due to mixture inconsistencies [1] than are engines running with stoichiometric mixtures. Fast combustion [4] improves efficiency by bringing the combustion phase of the engine cycle closer to the ideal constant volume combustion. Fast combustion also provides greater resistance to knock while reducing cyclic variations 1 Chapter 1. Background and Review 2 in power which leads to better driveability. The drawback with fast burn engines is a higher peak pressure and temperature which results in an increase of nitrogen oxide emissions, although this tendency can be compensated for through timing changes, the use of exhaust gas recirculation (EGR) and the use of lean mixtures. Consequently fast combustion is an important factor in engine advancements particularly with lean mixtures. Fast and lean combustion is of particular importance when used in conjunction with slow burning fuels such as natural gas which consists mainly of methane. Natural gas could become an important alternative fuel especially in areas such as western Canada where there is a large domestic supply. In such a situation there are both economic and political reasons for encouraging the use of natural gas in internal combustion en-gines. Natural gas is presently less than half as expensive as gasoline and becoming less dependent on external energy supply is always politically sound. Unfortunately, when converting to natural gas there is a loss of power due both to the slow burning nature of the fuel and the fact that it is a gas. Gaseous fuels reduce the amount of fuel/air mixture inducted into an engine compared to liquid fuels because the increased volume of the fuel displaces some air causing a loss in power. In order to make natural gas a viable alternative for automotive applications the power loss associated with its use should be reduced. To achieve an improvement in performance it is necessary to have a good understanding of the effects of engine parameters on combustion. There are three factors controlling the burning rate in a spark ignition engine: the chemical kinetics of the fuel, the combustion chamber geometry, and the mixture motion in the cylinder. Chemical kinetics, although an important factor in determining flame speed, are a direct function of the fuel/air mixture being used and are not controlled by engine design parameters. Consequently chemical kinetics are not considered further in this study. Combustion chamber geometry, the second factor, can reduce the burning Chapter 1. Background and Review 3 duration by reducing the distance the flame must travel to engulf the mixture. This can be accomplished by using compact chamber designs with central ignition. This type of chamber geometry also reduces the flame contact area with the chamber walls which has the added effect of reducing heat transfer which also increases burning rate [5]. Chamber geometry can also have an effect on mixture motion which is the final factor considered to control burning rate. Mixture motion in an engine is very complex and has a number of effects: it can improve fuel/air mixing, it can aid in fuel vaporization for liquid fuels and it can increase the flame speed. Most of these effects are from small scale velocity fluctuations known as turbulence. The most important of which is the direct interaction between the turbulence and the flame. Two theories have been developed in order to study the relationship between turbu-lence and flame speed. The first is the wrinkled flame model proposed by Damkohler [6] and the second is the turbulent entrainment model proposed by Blizard and Keck [7] and improved by Tabaczynski et al [8]. The wrinkled flame model states that the increase in flame speed is due to an increased flame front area caused by flame element distortions or wrinkles. The turbulent entrainment model states that there are regions of intermittent increased activity in a flame front which entrain pockets of unburnt mixture which burn at the laminar rate. The regions of increased activity are assumed to be vortex tubes of the Kolmogorov scale as suggested by Tennekes [9] and the regions of laminar burning are characterized by the Taylor microscale which is also consistent with Tennekes' turbulence model. Irrespective of the two mechanisms there is a myriad of factors affecting turbulent combustion in an engine. Although a definite link between turbulence and combustion has been established, a more detailed look at mixture motion is required before proceeding. Chapter 1. Background and Review 4 1.2 Mixture Motion in an Internal Combustion Engine For a simple disc combustion chamber geometry the majority of the mixture motion is known to be generated from the intake process as recently demonstrated by Dohring [10]. A large portion of the energy of the inflow is in the form of turbulence caused by the shear flow past the intake valve. Once the valve closes the turbulence rapidly decays. During the compression stroke the turbulence is sustained at a low level due to compression effects. As the piston approaches top dead centre (TDC) the compression rate decreases causing the turbulence to increase its rate of decay. This is the point at which the mixture is normally ignited, with the turbulent energy much lower than the original energy produced during intake. In a simple disc combustion chamber the turbulent energy is lowest at the most critical point in the combustion process, namely at ignition. Two approaches have been taken to increase the turbulence late in the compression stroke at the time of ignition. The first is through swirl which sustains the intake energy in an organized motion so that it breaks up into turbulence later during the compression stroke. The second approach is through squish which produces a flow from the piston motion near the end of the compression stroke. The actual effects of swirl and squish are not certain and their measured effectiveness in increasing combustion rate does vary between researchers. The effects of swirl alone in a disc combustion chamber were studied by Ma [11] and Witze et al [12]. Ma concluded from his experiments that while swirl could be correlated with smoke emissions from a direct injection diesel engine it only produced marginal effects on performance, economy and emissions in spark ignition engines. Witze, on the other hand, found that with high swirl there was a significant increase in burning speed and his photographs showed that the flame remained attached to the spark plug and Chapter 1. Background and Review 5 was smeared around the combustion chamber by the swirling motion. Another similar study, but without a perfect disc combustion chamber, was done by Nagayama et al [13] in which swirl was found to be beneficial. Although many others have evidence that swirl does have beneficial effects there is still a need for more extensive experiments to determine its exact role in improving combustion. The effects of squish have also been considered by a large number of researchers over a wide range of engine conditions. One popular squish generating combustion chamber geometry is the bowl-in-piston design. This design is symmetric about the centreline of the piston and produces a flow towards the centre of the bowl as the piston reaches TDC. Ma conducted extensive experiments with this configuration and found that although there did not appear to be any orderly movement, there was a beneficial increase in small scale turbulence. It is widely agreed that while squish motion is not directly beneficial to combustion, the increase in turbulence intensity associated with squish generally shortens the early flame growth time. Nagayama and other researchers have shown that a combination of both swirl and squish is more beneficial than either alone. It has been suggested that squish helps the breakup of swirling motion producing a high turbulence intensity at ignition. The effects of both are generally very dependent on combustion chamber geometry and spark location. The main contribution from mixture motion on combustion is seen to be its effect on turbulence intensity around the time of combustion. 1.3 Objective and Scope The objective of this work was to investigate the effects of turbulence on combustion rate in a spark ignition engine environment. Increased turbulence has been known to increase the burning rate in an engine and is a possible means of achieving fast combustion with Chapter 1. Background and Review 6 slow burning fuels and lean mixtures. The focus of the work was separated into two parts. For the first part, the effects on combustion of turbulence scale and intensity was investigated using perforated plates to generate turbulence in the combustion chamber of the rapid intake and compression machine (RICM). The turbulence scale and intensity was of interest so that the kind of turbulence that was most effective in increasing the burning rate in an engine could be determined. For the second part, the effects of several "squish-jet" combustion chamber designs on turbulence and combustion was examined in the RICM. Of particular interest in this part of the study was to investigate the proposed designs ability to generate turbulence during the compression stroke which would improve combustion. A stoichiometric mixture of methane/air was used in both parts for the combustion experiments. Three different sizes of perforated plates were mounted in the RICM so that they would generate turbulence of different scales and intensities during the compression stroke. A hot-wire anemometer (HWA) was used in several locations and at two ori-entations to measure the flow field generated by each perforated plate. Combustion experiments were run with several ignition timings for each perforated plate and the combustion duration was determined from pressure measurements using a mass fraction burned (MFB) analysis program. The different turbulence scales and intensities deter-mined from the measurements were compared with the MFB curves and the combustion durations in order to determine the relative effects of both scale and intensity on com-bustion. Previous work done by Dymala-Dolesky [14] using "squish-jet" combustion chamber designs suggested that a forced jet action was needed to produce strong jets. If the jets were strong enough and had the appropriate timing they could produce turbulence and enhance combustion. The piston designs used in this work were based on the bowl-in-piston design with Chapter 1. Background and Review 7 a squish area of 75%. Four "forced squish-jet" designs, which mated with a step in the cylinder head so that the fuel/air mixture would be forced through the jet holes, were made to fit the RICM. A standard bowl-in-piston design, which was used with a flat cylinder head, was also made to fit the RICM for a basis of comparison with the new "forced squish-jet" designs. With the compression stroke by itself, HWA and combustion measurements were taken with each design. The HWA was located in several positions so that the jet motion, squish motion and motion at the spark locations could be determined. Several spark locations and spark timings were used in an attempt to maximize the performance of the different designs so that they could be compared on an even basis. The effects of the turbulence strength and timing on combustion was investigated and compared to the results produced by the standard bowl-in-piston design. 1.4 Review of Previous Work The first direct measurements of turbulence in an engine were conducted by Semenov [15] in the late fifties using a hot-wire anemometer (HWA). He concluded that the bulk motion was generated by the intake process. The turbulence from the intake stroke was found to decay during the compression stroke and was found to be homogeneous and isotropic near TDC. The conditions in an engine are for the most part far from those for which a HWA has been calibrated. Horvatin and Hussmann [16] and later Hassen and Dent [17] studied heat transfer equations related to a HWA so that wind tunnel calibration could be extended to include the pressure and temperature conditions in an engine. Both groups successfully measured the flow in motored diesel engines but did not relate the flow to any combustion measurements. Chapter 1. Background and Review 8 Ohigashi et al [18] examined the effects of turbulence on combustion by pulling a per-forated plate across a disc shaped combustion chamber before igniting the propane/air mixture. The intensity of the turbulence was measured inside the combustion chamber using a HWA in air at the same conditions as the combustible mixture before igni-tion. High speed motion pictures of the flame development were taken along with the combustion pressure records during the combustion runs. Based on the uniform flame development for all values of turbulence intensity they suggested that the turbulence field was almost uniform (homogeneous). They found that the initial flame speed was affected by the equivalence ratio, and the turbulence intensity; but that once the flame had reached a certain size its burning rate was dominated by the turbulence intensity with the laminar burning rate of the propane/air mixture (a function of $) being of secondary importance.They found that the scale of the turbulence, implied from the perforated plate hole size, had no significant effect on the burn rate. Winsor and Patterson [19] examined the effects of turbulence on combustion in an engine with a disc combustion chamber. Using a HWA at three different orientations and at several depths they found the turbulence to be homogeneous and isotropic near TDC. The mean turbulence velocity was found to increase linearly with engine speed. A model was developed relating the combustion duration to the cylinder velocity variations at the spark plug. Based on this model the critical flame radius for which turbulence became effective was found to be about 0.4 inches (10 mm). A combustion model they developed suggested that a reduction in eddy size could reduce the critical flame radius resulting in a shorter flame initiation period. Tsuge et al [20] studied the characteristics of turbulence in a closed vessel using a perforated plate to generate the turbulence in a manner similar to that used by Ohigashi et al. A HWA was used to make velocity measurements at several locations across the perforated plate in order to determine the homogeneity of the turbulence generated. They Chapter 1. Background and Review 9 found that the decaying process could be considered in two parts: the pre-relaxation stage and the relaxed stage. The short pre-relaxation stage consisted of strong local anisotropics associated with the method of turbulence generation while the relaxed stage consisted of homogeneous isotropic turbulence characteristic to the form of the vessel. This could explain why Ohigashi found no effects from different scales (hole sizes) if he was in the relaxed region. Dent and Salama [21] investigated the turbulence structure in two engines, using a HWA and random signal analysis, and presented an improved procedure for estimating the turbulence parameters. They found that large structures were beneficial in enhancing the propagation of an established flame, but tended to quench early flame kernel growth. Small scale turbulence was found to be important for early flame kernel growth. Andrews et al [22] in a broad review paper of turbulent combustion theories proposed that the microscale Reynolds number, R\, is important in determining the ratio of turbulent burning velocity to laminar burning velocity. Available experimental data supported a correlation which indicated the importance of turbulence scale on combustion. Lancaster [23] measured the turbulence parameters in a motored engine with a triaxial HWA over a wide range of engine conditions. He evaluated the data using ensemble-averaging and time-averaging techniques as well as with a nonstationary time-averaging technique which he developed. Near TDC the turbulence was found to be isotropic and both turbulence scale and intensity were found to be functions of intake volume flow rate for the geometries tested. From a plot of the turbulence energy spectrum most of the turbulence energy was found to be of a frequency below 1000 Hz. The turbulence near TDC was also found to scale with engine speed but compression ratio and vol\imetric efficiency changes had little effect. In an accompanying paper [24] the effects of the turbulence on combustion were analyzed. Pressure data was used in conjunction with a combustion heat release model to calculate turbulent flame speed. The ratio of turbulent Chapter 1. Background and Review 10 flame speed to laminar flame speed was found to be a linear function of turbulence intensity. Tabaczynski [25] reviewed the features of turbulent flow in an engine during intake and compression and the influence of the turbulence field on the combustion process. He concluded that turbulent scale was a function of chamber height and that some mea-surements previously associated with isotropic flow also indicated a situation of no mean flow. He stated that the equality of HWA measurements regardless of wire orientation implied that no mean flow existed and that there was isotropic turbulence. The HWA's inability to determine flow direction caused low frequency turbulence with no mean flow to produce a HWA signal normally interpreted as a mean flow. A technique was pre-sented to determine the turbulence intensity under such conditions but it was necessary to first know that there was no mean flow. The effects of turbulence on combustion was concluded to be consistent with most of the previous work. Small scale turbulence was important for the flame initiation period but the main combustion rate was a function of turbulence intensity only. A suggestion given by Tabaczynski was that conventional turbulent flows, such as the flow behind a grid, could correlate well with engine flows and thus be useful in engine research. Witze [26] measured the spatial distribution of turbulence in an engine with a HWA and found the flow field to be inhomogeneous near TDC in contrast to earlier findings. The effects of squish were also investigated but were found to be insignificant compared to the effects of the intake flow. The length scales were found to be a function of geometry only while mean velocity, turbulence intensity, and time scales were functions of engine speed. The laser doppler anemometer (LDA) was introduced for measurements inside an engine in the late seventies. Due to low data rates, only ensemble-average data process-ing techniques were used in early studies [27] so that turbulence intensity could not be Chapter 1. Background and Review 11 separated from cyclic fluctuations. Cole and Swords [28] used a conditional sampling technique to successfully measure both mean velocity and turbulence intensity in a fired engine which is not possible with a HWA. A strong correlation was found between the mean velocity at the spark plug and the peak pressure. Witze [29] compared LDA and HWA measurements in a motored engine to check the validity of HWA measurements. In so doing, better parameters were devised for the analysis of the heat transfer from the HWA. The result was a good agreement between the LDA and HWA velocity measurements. Rask [30] compared different LDA analysis techniques in order to eliminate systematic errors involved in the techniques (such as "crankangle broadening"). He then suggested appropriate parameters in order to produce— sufficient accuracy of results. ^ ^ " ^ A detailed study on the effects of turbulence on combustion was performed by Checkel turbulence of various scales and intensities. The turbulence generated by the perforated plates was measured in a wind tunnel and in the bomb using a HWA. The technique for measuring a non-flowing turbulence field was further developed to produce turbulence intensity and decay rates from the HWA signal in the bomb. The turbulence in the bomb was successfully compared to the wind tunnel turbulence making the inference of turbulent length scales in the closed bomb possible. Flows with high rates of energy dissipation were found most effective in enhancing combustion. The scale was again found important in the flame initiation period since the turbulent eddies must be of a size in the order of the flame kernel before they become effective. From turbulence decay observations Checkel proposed that any useful small scale turbulence from the intake stroke would clearly decay before TDC. Large structures such as swirl could last long enough but to become beneficial are required to break up near TDC. [1]. Perforated plates with different hole sizes were pulled across a bomb to generate Chapter 1. Background and Review 12 Liou and Santavicca [31] measured time and length scales as well as the energy spec-trum in an engine with and without swirl. Their cycle resolved LDA measurements showed that turbulence intensity and time scales both scale with engine speed as ob-served in earlier studies. The energy spectrum was found to shift to higher frequencies with higher engine speeds and swirl was found to reduce cyclic variations in the velocity field by a factor of two. Martin et al [32] used a LDA to measure the effect of the flame front on the flow field in an engine. In order to eliminate cyclic variation biases and complex flow fields it was necessary to control the overall flow field in the engine [33]. This was accomplished by installing a grid across the entire cylinder to produce isotropic homogeneous turbulence with no mean velocity at the time of ignition. The flame was found to amplify the turbulence inhomogeneously in the direction of flame travel just ahead of the flame. The flame front had little effect on the turbulence parallel to the flame surface. Despite the popularity of the LDA measurement system, the HWA was still being used and improved for engine measurements. Catania and Mittica [34] developed a cycle-by-cycle nonstationary time-averaging method of data analysis which could separate the mean velocity, turbulence intensity and cyclic fluctuations from a HWA signal taken in an engine. The method was more straightforward and general than previous cycle-by-cycle data analysis techniques and in particular cases would reduce to conventional time-averaging or ensemble-averaging procedures. This new method was used by Cameron [5] where it was also compared to Lancasters nonstationary method. Catania and Mittica's method was found to be the better of the two because of its treatment of the cyclic variation in the mean velocity. Fraser et al [35] applied an LDA system making two-point spatial correlation mea-surements of velocity fluctuations to the flow field in an engine. This produced the first direct measurements of the integral length scale in an engine using an LDA system. The Chapter 1. Background and Review 13 integral length scale was found to be about one fifth the chamber height at TDC for the disc combustion chamber tested. Although comparison with other results using different techniques is suspect, length scales measured by others agreed within a factor of two. Hall and Bracco [36] have measured the flow field in an engine under both fired and motored conditions using cycle resolved LDA measurements. The measurements were made in an engine having a disc combustion chamber with and without swirl. The results agreed with those of Martin et al in so far as the flame front had little effect parallel to the flame surface. In the direction of the flame travel they found that there was little or no amplification of turbulence intensity in front of the flame but there was a magnification across the flame. With and without swirl the turbulence was found to be nearly isotropic and homogeneous but with swirl the turbulence intensity was up to sixty percent larger than without swirl. It is generally difficult to compare results from different authors because of the variety of techniques used and conditions tested. Never the less there appears to be a general consensus on several points: • The intake process is the major source of turbulent energy. • Turbulence rapidly decays during the compression stroke although the effects of compression may help to sustain it to some extent. • As the stroke approaches TDC on compression the turbulence tends to be isotropic. • Although the turbulence near TDC appears to tend to be homogeneous there is no clear evidence that this is the case. • The turbulence scale is generally a function of chamber geometry or clearance height. Chapter 1. Background and Review 14 • Squish appears to have a minimal effect on the mean flow pattern near TDC al-though it is usually beneficial to combustion. • Swirl reduces cyclic velocity variations and increases turbulence intensity near TDC in most engines. • Mean velocity, turbulence intensity and turbulence scales are generally functions of engine speed. • Small scale turbulence is important to early flame development. • Turbulence intensity, independent of turbulence scale, is the most important factor controlling combustion rate in the main combustion period. Chapter 2 Experimental Apparatus and Method 2.1 Rapid Intake and Compression Machine All of the experiments were conducted in the rapid intake and compression machine (RICM) at the University of British Columbia's Department of Mechanical Engineering. The RICM was designed by Dohring and a complete description of the machine is given in his M.A.Sc. thesis [10]. A brief description of the apparatus is given in the next section followed by a description of modifications necessary for the present work. 2.1.1 Existing R I C M Apparatus The RICM was built in order to investigate the relative effects of intake and compression stroke generated turbulence on the combustion duration in a spark ignition engine. The RICM can be run with both an intake and compression stroke or with a compression stroke alone. In either case the RICM operates at a maximum simulated engine speed of 1000 rpm with the piston beginning and ending at rest. For the two stroke case the piston starts at TDC before intake and the intake mixture is drawn from a small tank which is initially at ambient pressure and temperature. For the one stroke case the piston starts at bottom dead centre (BDC) before compression with the cylinder pressure set to 50.8 kPa in accordance with the pressure at BDC in the two stroke case (as was determined by Dohring [10]). In both cases the piston stops at TDC after the compression stroke to provide for constant volume combustion. 15 Chapter 2. Experimental Apparatus and Method 16 The process is energized by a pneumatic driving cylinder on one side of the rack and controlled through a hydraulic ram on the other side of the rack. The rack is meshed with a pinion connected to the crank shaft which moves the piston. A schematic of the assembly is shown in figure 2.1. Once the rack has rapidly accelerated, through the triggering of a fast response solenoid valve, the constant flow rate orifice on the hydraulic side of the assembly keeps the rack at a constant velocity. A braking pin, which blocks the flow through the orifice, was designed to bring the rack to a rapid stop at the end of the compression stroke without a destructive collision between the ram and the orifice. A groove in the rack was milled so that a pin in the groove would control the intake valve through the linkage assembly (see figure 2.2). The combustion chamber specifications and RICM running parameters are given in table 2.1. The piston, cylinder and head assembly were designed to be easily changed for future modification. The piston has a removable crown which can be replaced to provide a variety of combustion chamber geometries although only the disc combustion chamber crown was designed. The cylinder was made from a pipe available off-the-shelf with a micro-honed inside surface so that it can easily be replaced if necessary. The head assembly was machined from mild steel and no cooling is required. This makes alterations or replacement much simpler than with conventional water or air cooled engine heads. 2.1.2 Modifications to the R I C M The cylinder and head plate of the RICM were modified to accommodate perforated plates located in the cylinder just above the TDC piston location. Three perforated plates with hole sizes of 20, 10 and 5 millimeters were made in an attempt to generate turbulence of different scales by forcing the fuel/air mixture through the perforated plate holes with the piston as it approached TDC. The blockage ratio of each perforated plate was 61 % to be consistent with previous research by Checkel [1]. Chapter 2. Experimental Apparatus and Method 17 The head of the RICM was replaced with a 2 inch (50.8 mm) thick quartz window to provide full field optical access to the combustion chamber for both flow and com-bustion visualization experiments using high speed motion pictures. Unfortunately early attempts at flow visualization and flame photography were unsuccessful so that both were abandoned. The spark probe and pressure transducer were relocated into the head plate and an intake port was relocated into the cylinder just above the piston BDC lo-cation. The fuel/air mixture was introduced into the cylinder when the piston was at BDC through a manually controlled valve mounted in the cylinder intake port. Once the cylinder had been charged, the intake valve was closed and the RICM could be triggered for the compression stroke only tests. The perforated plates were removed from the combustion chamber and a standard bowl-in-piston piston crown (denoted piston BIP)was built to replace the disc piston crown used in the perforated plate experiments. The cylinder was also replaced in order to accommodate the longer length of piston BIP. The bowl-in-piston design had a squish area of 75%, a clearance gap of 1 millimeter and retained the 8 to 1 compression ratio. Because piston BIP came within 1 millimeter of the head the HWA, pressure transducer and spark probe had to be moved to the head. A mild steel disc with a pressure transducer fitting and two HWA or spark probe ports was made to replace the quartz window. Figure 2.3 shows the four different "forced squish-jet" designs used with the RICM. A step in the head was required for the ridge on the pistons to fit into so that the squish fluid would be forced through the jet slots. This was accomplished by inserting a spacer the thickness of the step (5 millimeters) between the head and the head plate (see figure 2.4). All the different "forced squish-jet" designs retained the 75% squish area, 1 millimeter clearance gap and the 8 to 1 compression ratio. Piston 2J was the base "forced squish-jet" design from which the other "forced squish-jet" designs were compared in order to determine the effects of changing some design Chapter 2. Experimental Apparatus and Method 18 parameters. Piston 2J had two 2 millimeter wide 5 millimeter high opposing jet slots. Piston 2LJ was built to examine the influence of the width of the jet slots so it had two 4 millimeter wide 5 millimeter high opposing jet slots. Piston 4J was built to examine the influence of the number of jet slots so it had four 2 millimeter wide 5 millimeter high jet slots which would produce two sets of opposing jets 90 degrees to each other. Finally piston NJ was built to examine the influence of the ridge by itself after the previous designs showed minimal effects from the jet slots. Piston NJ had no jet slots but still retained the 5 millimeter ridge of the "forced squish-jet" design. 2.2 Data Acquisition A Cyborg Isaac 2000 high speed data acquisition system controlled through an IBM PC was used to collect data on three channels simultaneously. A data acquisition rate of 62.5 kHz was used to be consistent with previous work [10] and to maintain a good resolution in terms of crank angle position. Each channel had a twelve bit resolution for a voltage ranging from —10 to 10 volts producing a digitization error of less than 0.005 volts. Data collection was externally triggered from an optical pickup located on the rack to collect 20,000 points of data over 106.7 milliseconds for the one stroke measurements. The data were transferred from the acquisition system through a serial port to the IBM PC and then transferred to diskette. For each test condition a minimum of 10 sets of data were collected for averaging of the results. The data were transferred from diskette through an Ethernet connection to a VAX computer network for processing. A Nicolet 3091 digital storage oscilloscope was connected in parallel with two channels of the system for immediate viewing of the data before storage. A second Nicolet 3091 digital storage oscilloscope was used to measure the third channel for some cases. The oscilloscopes and acquisition system were triggered by the same optical pickup. When Chapter 2. Experimental Apparatus and Method 19 running combustion tests it was found that the primary ignition voltage should not be measured with an oscilloscope because of a grounding problem between the two which caused premature ignition. For the flow field measurements the three parameters measured were: 2. the hot-wire anemometer bridge output voltage 3. the crank angle optical pickup voltage. For the combustion measurements the three parameters measured were: 1. the pressure transducer charge amplifier output voltage 2. the crank angle optical pickup voltage 3. the ignition system primary voltage. 2.3 Flow Field Measurements A constant temperature hot-wire anemometer (HWA) was used to make point velocity measurements of the flow field inside the combustion chamber of the RICM. The equip-ment used is given in table 2.2. The wire was welded to the probe and then annealed for about seven hours until the wire had stabilized. Once the wire had stabilized the resistance was measured and the overheat ratio was calculated using the relationship: where a is the thermal coefficient of resistance. The bridge resistance was then set to operate the wire at the desired temperature of 600°C. The gain, HF filter and cable compensation settings were adjusted as outlined in the bridge manual to produce the 1 the pressure transducer charge amplifier output voltage Hop — ^ a m b ( l ^(-^op -^ amfc )) Chapter 2. Experimental Apparatus and Method 20 best frequency response for the system. The low-pass filter was set to 20 kHz (less then half the sampling frequency of 62.5 kHz) in order to avoid any problems of aliasing. Calibration measurements were made in a wind tunnel at atmospheric conditions over speeds ranging from 0.5 to 18 m/s using a pitot-static tube to measure the reference velocity. In order to develop an understanding of the flow field generated by the perforated plates several point measurements were required. For each of the three perforated plates the HWA measurements were taken in at least three locations with two wire orientations in each location. The three locations chosen were: 1. lined up with a hole in the perforated plate at the centre of the combustion chamber, 2. between the hole in the perforated plate at the centre of the combustion chamber and an adjacent hole, 3. lined up with a hole in the perforated plate closer to the wall of the combustion chamber. The two wire orientations were: 1. with the wire parallel to the perforated plate in order to measure the velocity along the axis of the cylinder, 2. with the wire perpendicular to the perforated plate in order to measure the velocity across the head and piston surface. The three locations were selected in order to determine the homogeneity of (lie flow field and the two orientations were selected to determine if the turbulence was isotropic according to the technique recommended by Tabaczynski [25]. Chapter 2. Experimental Apparatus and Method 21 In order to determine the squish motion, the jet motion and the motion at the spark location HWA measurements were taken in several locations for each combustion chamber design. The HWA was inserted into the combustion chamber through the disc head which replaced the quartz window at two locations on axes perpendicular to the head. These ports allowed HWA measurements to be taken: 1. on the centreline of the cylinder, 2. 3.5 millimeters inside of the bowl edge. On each axis the HWA probe was adjusted so that it was 2, 8 or 16 millimeters from the head surface. At the centre location the rotation of the probe was such that the wire was perpendicular to a ray from a jet slot to the probe so that any jet motion that reached this location would be detected. At the location by the edge of the bowl the rotation of the probe was such that the wire was parallel to the bowl edge closest to the wire. In this way the squish motion was measured when the disc was rotated so that the probe was between the jet slots and the jet motion was measured when the disc was rotated so that the probe was lined up with a jet slot. Table 2.3 lists which HWA probe positions were used in the different piston design tests. 2.4 Combustion Measurements The rate of combustion in the RICM was determined from pressure and crank angle data. The analysis of this pressure and crank angle data is given in a later section (see section 3.2). A piezoelectric pressure transducer (Kistler type 6123 pressure transducer serial num-ber 177004 and a Kistler 5504 Dual Mode Charge Amplifier) was mounted in either the head or the head plate of the RICM in order to measure the pressure in the combustion Chapter 2. Experimental Apparatus and Method 22 chamber. The transducer was recessed half of its diameter to protect it from thermal shock, as recommended by Benson and Pick [37]. A slotted disc and optical pickup assembly was mounted on the end of the crank shaft in order to measure the crank angle position. This system produced a square wave, 2 crank angle degrees in period, which was used to count the degrees of rotation from the initial triggering position at BDC. A stoichiometric mixture of methane/air was prepared in a storage tank for the com-bustion tests. The composition was measured with a Hewlett Packard model 5750b Research Gas Chromatograph as described by Pierik [38]. The calibration curves given by Pierik were also used. The mixture composition used was found to be within 0.25 % of stoichiometric for all the combustion tests according to the gas chromatograph mea-surements. A Delta products inc. Mark Ten B capacitive discharge ignition system was used to ignite the mixture. The timing was controlled through a digital timer built in the department. The timing could be adjusted in one degree increments for ignition from 99 to 0 degrees before TDC (BTDC) and a delay timer could be adjusted for ignition after TDC. For the experiments with the perforated plates a long thin spark probe was made which would fit into a HWA access port in the head plate so that there could be central ignition with the quartz window in place. For each of the three perforated plates, the combustion tests were run with ignition timing set for TDC as well as 10, 20, 30 and 40 degrees BTDC so that the effects of the different turbulence intensities at these timings could be determined. The long thin spark probe was mounted through the HWA access port in the centre of the steel disc head, which replaced the quartz window, so that there could be cen-tral ignition with the bowl-in-piston chamber designs. The spark probe could not be Chapter 2. Experimental Apparatus and Method 23 mounted in the head plate because the pistons extended to the top of the cylinder past the head plate HWA access ports. The spark probe was adjusted so that it was 2, 8 or 16 millimeters from the head surface (in accordance with the HWA locations) and the ignition timing was set for 20, 30 or 40 degrees BTDC so that the best practical spark timing could be found for each design for an unbiased comparison. Table 2.4 lists which spark positions and timings were used in the different piston design tests. Chapter 3 Data Analysis 3.1 Flow Field Data Analysis 3.1.1 Hot-Wire Anemometer Data Analysis Since the temperature and pressure in the combustion chamber of the RICM vary widely and are, for the most part, substantially different from the conditions for which the calibration measurements were made for the HWA, heat transfer equations are required to extend the calibration of the HWA into the engine-like conditions. The original work on heat transfer from a wire with applications to the HWA was done by King [39] in the early part of this century where the Nusselt-Reynolds relationship of the form Nu = A + B(Re)n was developed where A, B and n are constants. The heat transfer equations were devel-oped further by Collis and Williams [40] and by Davies and Fisher [41]. The application of this technique to a HWA in an engine was demonstrated by Horvatin and Hussmann [16] and by Hassan and Dent [17] near the end of the sixties. Witze [29] compared HWA measurements to LDA measurements in an engine and found the results were very close if all gas properties in the heat transfer equations for the wire were evaluated at the free-stream conditions. A detailed explanation of the equations and procedure used for this work to determine the instantaneous velocity from a HWA signal, based on an anal-ysis performed by Lancaster [23] and using the recommendation of Witze, was given by Cameron [5]. 24 Chapter 3. Data Analysis 25 The calibration constants were evaluated from the wind tunnel velocity measurements based on the pitot-static pressure readings. The temperature in the RICM was calculated from the pressure measurements assuming a polytropic relationship. An evaluation of the limitations of this technique is given in the next section. The interpretation of the velocity measurements in order to understand the mixture motion effects has become associated with the characterization of the turbulent flow field in the combustion chamber. The main parameters considered are: mean velocity, turbulence intensity and length scales. These parameters are from the theory of isotropic turbulence for a stationary process, such as in a wind tunnel, but are not easily applicable to the nonstationary motion in an engine or the RICM. This has led to the development of special techniques and definitions for determining the turbulence parameters in an engine-like environment. The definitions and parameters chosen for this work were developed by Catania and Mittica [34] and were used by Dohring [10] in previous work with the RICM. The mean velocity profile for one cycle of the RICM was calculated by time averaging the instantaneous velocity over short time intervals known as windows. The window-averaged mean velocity in cycle i at time tw is defined as: U(i,tw) = ± ftw+" U(i,t)dt where U(i,t) is the instantaneous velocity in cycle i at time t, T is the window size and tw is the time in the centre of the window. A window size of 2 milliseconds, as selected by Dohring [10], was used. This corresponded to a cut-off frequency of 500 Hz so that all velocity fluctuations below 500 Hz were considered to be changes in the mean velocity while all fluctuations above 500 Hz were considered to be turbulence fluctuations. The cut-off frequency was within the range suggested by Liou and Santavicca [31] of between 300 and 900 Hz. Considering that the speed of the RICM was approximately equivalent Chapter 3. Data Analysis 26 to 1000 rpms, the window size corresponded to 12 degrees of crank angle revolution. This is within the range suggested by Catania and Mittica [34] of between 8 and 20 degrees. Many other researchers have also used similar window sizes [5, 10, 14, 29, 30, 42, 43]. The window-averaged mean velocities for each cycle were taken to represent the mean velocity at the centre of each window (ie. at time tw). A cubic-spline curve fitting routine with zero tension was used to interpolate between these mean velocity points so that the curve produced passed through each window averaged point, U(i,tw). The result was a mean velocity profile for each cycle i at any time t denoted by U(i,t). The turbulent velocity fluctuations u(i, t) were then found by subtracting the mean velocity from the instantaneous velocity in each cycle as follows: The root-mean-square (rms) of the turbulent velocity fluctuations, better known as the turbulence intensity, was also found using window averaging. The window averaged turbulence intensity in cycle i at time tw is denned as: The same interpolation method used to find the mean velocity was then used to find the The mean velocity and turbulence intensity found for each cycle was ensemble-averaged over a number of cycles for each test condition. The ensemble-averaged mean velocity at time t is defined as: u{i,t) = U(i,t) - U(i,t) turbulence intensity for each cycle i at any time t denoted by u'(i,t). and the ensemble-averaged turbulence intensity at time t is defined as: Chapter 3. Data Analysis 27 where N is the number of cycles for the test condition. The effects of the sample size was investigated by varying N between 2 and 15 cycles. Since the ensemble-averaged values did not significantly vary when N was 10 or greater (see figure 3.1) a minimum of 10 sets of data were collected for each test condition (as stated in section 2.2). The cyclic variation of the mean velocity is defined as: Uc(i,t) = U(i,t)-0E(t) The standard deviation in the cyclic variation of the mean over N cycles is defined as: U'c(t) 1 N and is sometimes referred to as the cyclic variation intensity. The mean velocity, turbulence intensity and cyclic variation of the mean velocity were more useful when presented with reference to crank angle rather than time. Since the degrees of crank angle revolution were measured at the same time as the velocity this was a simple conversion process using the definitions given below. The mean velocity in cycle i at crank angle 6 is defined as: U{i,6) = U(i,t9) where tg is the time corresponding to the crank angle position 6. Similarly, the turbulence intensity in cycle i at crank angle 6 is defined as: it'(i, 8) = u'(i,t8) and the cyclic variation of the mean velocity in cycle i at crank angle 6 is defined as: Uc(i,6) = Uc(i,te) The ensemble-averaged mean velocity and turbulence intensity at crank angle 9 are defined as: = i E a n d = iSriyM) i v t=l i v 1=1 Chapter 3. Data Analysis 28 respectively. The cyclic variation intensity at crank angle 0 is denned as: urn = N Using spatial correlations, the integral length scale is defined as: Lx = / R{r)dr Jo where R(r) is the spatial correlation coefficient R(r) u(xp)u(xp + r) xp is the measuring coordinate and r is the separation distance. The integral length scale characterizes the large eddies limited by the system boundaries which create velocity gradients resulting in turbulent stresses. The smaller scales are measured by the micro length scale, defined as: A - I 2 ' i <m provided that R(r) is nearly symmetrical near r = 0. The micro length scale characterizes the smallest eddy size outside the dissipation length. Unfortunately, the spatial correlation used to determine the length scales are diffi-cult to measure so that the length scales are usually determined from their time scale counterpart. The integral time scale is defined as: LT = [°° R(r)dr Jo where R(T) is the autocorrelation coefficient R(T) = u(r)u(t + r) t is time and r is the correlation time. The autocorrelation coefficient can easily be obtained using point measurements. The micro time scale is defined as: Chapter 3. Data Analysis 29 Taylors hypothesis [44] defines the relationship between length and time scales as : Lx — ULt and Xx = UXt when U is constant and U u'. This relationship holds quite well in wind tunnels but in an engine environment the constraints are not often met so that the length scales found in this manner are questionable at best. The scales in the present work were not calculated because of the inherent error but were implied from the dimensions of the apparatus. Although these definitions are not particularly good under engine like conditions, at present they are the best available. Consequently, the use of these turbulence parameters for describing the turbulent flow field inside an engine should only be used for qualitative analysis. 3.1.2 Limitations of the H W A in the R I C M There are two areas in which errors associated with HWA measurements can be cate-gorized. The first area was based on the measurement accuracy and was related to the calibration of the wire and the environment in which the measurements were taken. The second area was based on the interpretation of the measurements and was related to the directional ambiguity and the determination of a mean and turbulent component from the HWA signal. Because of the limitations of the HWA in engine applications the results should only be considered to be representative of the flow field. Therefore mean velocity and tur-bulence intensity measurements made with a HWA should only be used for qualitative comparisons. An obvious limitation with the application of the HWA to an engine-like environment is in the calibration. Although a great deal of work has been done on the heat transfer characteristics of the wire, the calibration of the HWA was extrapolated far beyond the Chapter 3. Data Analysis 30 calibration range based on an analytical model and consequently is a major source of uncertainty. Even if the heat transfer relationships were exact over the temperature and pressure range of the measurements the correct values of these parameters were not precisely known. Although the pressure change in the cylinder was accurately measured with a piezo-electric pressure transducer, the absolute pressure depended on an initial pressure mea-surement. The temperature, on the other hand, was not directly measured but calculated from the pressure measurements using a polytropic relationship and an assumed initial value. Although this method of determining temperature was quite crude, Lancaster [23] found through direct temperature measurements with a 5.0 fim resistance thermometer that the temperature in an engine could be calculated quite accurately using a polytropic relationship. For this work a polytropic exponent of 1.35 was selected as recommended by Dohring [10] who performed a sensitivity analysis for a variation in polytropic expo-nent from 1.30 to 1.40 and found at worst a 23 % change in the gas velocity at TDC. To avoid temperature errors caused by the thermal boundary layer, found by Hey wood [45] to be approximately 2 millimeters thick at TDC, HWA measurements near the wall should be restricted. The high temperature of the engine environment requires the wire to be operated at a high temperature; in this work 600° C. The temperature was limited by the properties of the wire which would change if the wire were heated above this value. As the piston approaches TDC on the compression stroke the temperature in the engine can reach values above 500°C. This reduces the sensitivity of the wire by decreasing the effective overheat ratio near TDC. The wire sensitivity was also reduced in this region because of the low mean flow near TDC. Consequently, the HWA measurements near T D C are not very accurate. A major limitation associated with the interpretation of HWA measurements under Chapter 3. Data Analysis 31 engine-like conditions was the inability of the HWA to determine flow direction. Because of this directional ambiguity a low frequency turbulence fluctuation could produce the same signal as a mean flow. This problem is particularly acute in an engine where the turbulent fluctuations can be very large compared to the mean flow. Techniques have been developed to deal with the measurement of the turbulence intensity in situations of no mean flow but they can only be used if it is already known that there was no mean flow. 3.2 Combustion Data Analysis The rate of combustion in the RICM was determined from pressure and crank angle data (see section 2.4). The method used for this work was based on the procedure developed by Rassweiler and Withrow [46]. The computer program written for the analysis and a complete description is given in Appendix A while the general approach is described below. Between the closing of the intake valve and the opening of the exhaust valve, the pressure rise in the combustion chamber can be caused from both piston motion and combustion. To separate the two effects the pressure and crank angle data were first broken up into small intervals. In each interval the pressure rise due to piston motion was calculated based on a polytropic compression of the fuel/air mixture starting from the pressure at the beginning of the interval. This pressure rise was subtracted from the total measured pressure rise in the increment leaving the pressure rise due to combustion. &Pzombustion — ^Ptotal(meatured) ~~ APpiston(calculated) This pressure rise due to combustion was assumed to have taken place at a. constant volume after the compression, namely at the volume at the end of the interval. Since the pressure rise produced at a constant volume from the combustion of a given mass of Chapter 3. Data Analysis 32 fuel/air mixture is inversely proportional to the volume [47], the pressure rise due to com-bustion in each interval was scaled to an equivalent pressure rise due to the combustion of the same mass of fuel/air mixture at one consistent volume. The volume chosen was the volume at TDC because in every case the majority of the combustion was expected to take place after the piston had stopped at TDC. ^interval &pvTDC = APvintcrval VTDC From the time of spark to the end of combustion the pressure rise due to combus-tion up to the end of each interval was found through the summation of these pressure rises over the preceding intervals. Using photographic measurements to determine mass fraction burned, Rassweiler and Withrow found that, throughout the entire combustion period, the percent pressure rise due to combustion was very nearly equal to the percent mass burned as predicted by simple combustion theory. Consequently, the mass fraction burned was approximated by dividing the pressure rise due to combustion up to the end of each interval by the total pressure rise due to combustion. Although this simple method was developed over 50 years ago, the mass fraction burned results produced with it agree very well with results produced through com-plicated methods using heat-release analysis. Young [48] in 1980 compared the simple method developed by Rassweiler and Withrow to a complex heat-release method de-veloped by Krieger and Borman [49]. The method developed by Krieger and Borman inferred the mass fraction burned from an energy balance among the work done, heat losses, internal energy of the mixture, and the fuel energy released during the combus-tion. This complex method required detailed information about the engine operating conditions, mixture composition, and cylinder heat transfer characteristics as well as the pressure and crank angle data required by the simple method. A good correlation was found with the two methods for the flame initiation periods and the main combustion Chapter 3. Data Analysis 33 durations, see figure 3.2. Amann [50] in 1985 compared several mass fraction burned analysis methods including the simple method developed by Rassweiler and Withrow and a sophisticated computer-based method developed by Young and Lienesch [51]. The mass fraction burned curve produced by the two methods were found to be in good agreement, see figure 3.3. Based on these comparisons it was decided that the results produced by the simple method developed by Rassweiler and Withrow would be accurate enough for the burn rate analysis in the present work. The separation of the MFB curves into the flame initiation period (FIP) and main combustion duration is some what arbitrary. The FIP was denned as the time from the spark to the point of 5% MFB. This selection was based on an error analysis performed by Kapil [52] and was considered the minimum MFB percentage which could be reliably calculated from pressure data. The FIP was an indication of the time required for the flame to develop, from the point of ignition, into a developed flame kernel and was expected to be a function of both turbulence scale and intensity (see section 1.4). The main combustion duration was defined as the time from the end of the FIP to the point of 90% MFB. This selection of 90% MFB was selected to eliminate the errors associated with determining the end of combustion (see section 4.2.2). The main combustion duration was an indication of the burning rate of the developed flame and was expected to be a function of turbulence intensity only (see section 1.4). Chapter 4 Discussion of Experimental Results 4.1 Effects of the Perforated Plates 4.1.1 Characterization of Turbulence In order to characterize the flow field and turbulence generated with the perforated plates the HWA velocity measurements were first separated into their mean and turbulent components as described in section 3.1.1. Three measurement locations and two wire orientations were chosen in order to help determine whether or not the flow field was homogeneous and/or isotropic. Figures 4.1, 4.2 and 4.3 show the mean velocity and turbulence intensity measured at the centre and near the edge of the combustion chamber for the 20, 10 and 5 millimeter perforated plates respectively. In each case the HWA was lined up with a hole in the perforated plate so that the velocity component through the hole was measured. The results for the 20 and 10 millimeter perforated plates show that the velocities are fairly similar for the two locations on each perforated plate. This indicates that for the 20 and 10 millimeter perforated plates the flow through all the holes in each perforated plate was similar across the entire perforated plate. The results from the 5 millimeter perforated plate show that the peak velocity through the centre hole was about twice that through the edge hole. The reason for this discrepancy is most likely caused by the cylinder wall because in this case the HWA was less that 3 millimeters from the cylinder wall. Based on the results of the other two perforated plates it was assumed that the flow around any 34 Chapter 4. Discussion of Experimental Results 35 hole for each of the perforated plates was similar (except very close to the wall where the flow was not considered to be important) so that the measurements taken around the centre hole of each perforated plate were representative of almost any hole in the perforated plate. The difference in the velocity profiles in figures 4.1, 4.2 and 4.3 are thought to be related to the HWA location and to the window size. In all three cases the same HWA location was used which caused the relative location based on hole diameter or Reynolds number (Re) to be different for each case. The window size used to separate the turbu-lence intensity from the mean velocity was based on engine parameters and not related to the perforated plate hole size. Since the scale and frequency of the turbulence generated by the perforated plates was related to the hole size, the window size which defines the cut-off frequency for the turbulence should have been a function of hole size as well. If the HWA location and window size were non-dimensionalized based on the hole size for each perforated plate the profiles could be more similar and may collapse onto a single curve. Checkel [1] was quite successful in collapsing the velocity and turbulence intensity profiles behind perforated plates with different hole sizes and at different speeds using the hole diameter and the "equivalent plate speed" to non-dimensionalize his data. Unfortu-nately, because the HWA was mounted through the side of the head plate measurements at different depths could not be taken so that this kind of analysis was not possible. Figures 4.4, 4.5 and 4.6 show the mean velocity and turbulence intensity measured at the centre of a hole and between two holes with the wire parallel and perpendicular to the perforated plate in each location for the 20, 10 and 5 millimeter perforated plates respectively. For all three perforated plates the figures show that the mean velocity and turbulence intensity could be separated into a pre-relaxation stage and a relaxed stage as defined by Tsuge et al [20]. The pre-relaxation stage consists of strong local anisotropics and the relaxed stage consists of homogeneous isotropic turbulence. The turbulence Chapter 4. Discussion of Experimental Results 36 was considered to be homogeneous where it became independent of HWA location and was considered to be isotropic where it became independent of the wire orientation as suggested by Tabaczynski [25] (at least in the two directions which could be measured with the present apparatus). For the 20 millimeter perforated plate (figure 4.4) the flow appears to become relaxed at around the 40 millisecond mark. Consequently, past this time any point measurement could be considered to be representative of the entire flow field [20]. For the 10 mil-limeter perforated plate (figure 4.5) the flow appears to become relaxed after around 35 milliseconds and for the 5 millimeter perforated plate (figure 4.6) the relaxed stage starts before the 30 millisecond mark. The different times at which the flow became relaxed for the three different perforated plates implies that the scale of the turbulence generated by each perforated plate was a function of the hole size as expected. Figure 4.7 shows the mean velocity and turbulence intensity for the three different perforated plates. In each case the HWA was lined up with the hole in the centre of the perforated plate with the wire parallel to the perforated plate (so that the velocity component through the hole was measured). This figure clearly shows the difference between the flows generated by the different hole sizes in the perforated plates. For each perforated plate the turbulence intensity peaks at the point where the mean velocity was rapidly decreasing. The time of the peaks were also coincident with the transition between the pre-relaxation stage and the relaxed stage which was where the strong local anisotropics had broken up into turbulence. Since the flow field in the pre-relaxation stage had strong local anisotropics, the HWA location must be considered when interpreting the data. For each of the three cases shown, the HWA was set up to measure the flow through the holes in the perforated plates. For each perforated plate there appeared to be a jet produced through the holes which peaked at approximately the 25 millisecond mark which coincided with the Chapter 4. Discussion of Experimental Results 37 expected time of maximum mass flux past the perforated plates based on piston position. These jets broke up into homogeneous isotropic turbulence a short time later depending on the perforated plate hole size. Based on these observations the scale of the turbulence was considered to be a function of the perforated plate hole size, in the pre-relaxation region, for the purpose of relating turbulence scale to combustion. In the pre-relaxation stage it was difficult to assess the turbulence intensity because of the strong influence of the HWA location on the results (see figures 4.4, 4.5 and 4.6). Although there was no apparent trend in the maximum turbulence intensity for the dif-ferent perforated plate hole sizes when looking at the measurements at any one location, by examining the turbulence intensities in figures 4.4, 4.5 and 4.6 it became apparent that the turbulence intensity was generally highest with the 20 millimeter perforated plate and lowest with the 5 millimeter perforated plate. This was consistent with the observation made by Checkel [1] that for the same velocities and decay times larger perforated plate hole sizes produce higher turbulence intensities and significantly larger integral scales. The determination of the turbulence intensity in the relaxed region was reasonably straight forward because the turbulence was homogeneous and isotropic in this region. Consequently, a single point measurement could be used to determine the turbulence intensity in the entire flow field. Figure 4.7 clearly shows that the turbulence intensity in the relaxed stage for the three perforated plates was a function of perforated plate hole size at any one time but that the decay rates were similar. Figure 4.8 is a plot of the turbulence intensities from figure 4.7 for the three perforated plates but on a log-log scale. The slopes in figure 4.7 for the different perforated plate hole sizes axe very similar indicating that the decay rates were approximately the same. This implies that the turbulence had decayed to a scale characteristic of the combustion chamber geometry in the relaxed region. These observations were consistent with the expectations that in the relaxed stage the turbulence was not influenced by the way in which it was generated, but Chapter 4. Discussion of Experimental Results 38 was a function of the vessel constraints as determined by Tsuge [20]. This dependence of the turbulence scale on the vessel dimensions was also consistent with the notion that the turbulence scale near TDC of an engine was generally a function of chamber geometry or clearance height (implying that the intake generated turbulence was relaxed near TDC). It should be noted that the combustion chamber volume does not change after TDC, which corresponds to around the 34 millisecond mark. 4.1.2 Effects on Combustion The mass fraction burned (MFB) curves and the combustion durations were calculated from pressure and crank angle measurements for the three different perforated plates as described in section 3.2. For each perforated plate five different ignition timings were used in order to cover a range of turbulence intensities for each scale of turbulence (perforated plate hole size). Table 4.1 shows the combustion durations for the three different perforated plates with the five ignition timings. The combustion duration was separated into two parts: the flame initiation period (FIP, 0-5% MFB) and the main combustion duration (5-90% MFB), see section 3.2. The FIP was an indication of the time required for the flame to develop, from the point of ignition, into a developed flame kernel and was expected to be a function of both turbulence scale and intensity (see section 1.4). The main combustion duration was an indication of the burning rate of the developed flame and was expected to be a function of turbulence intensity only. The standard deviation of the combustion durations (er) in all cases were relatively low, indicating the repeatability of the combustion process in the RICM and justifying the use of a small number of runs for each case. Figures 4.9, 4.10 and 4.11 show the MFB curves, with the five ignition timings, for the 20, 10 and 5 millimeter perforated plates respectively. The effect of the ignition Chapter 4. Discussion of Experimental Results 39 timing became much more pronounced as the perforated plate hole size was decreased. Looking back at the HWA measurements it can be seen that the turbulence intensity, over the same time span, also decreased more as the perforated plate hole sizes decreased. This trend indicates that the combustion rate was a function of turbulence intensity as concluded by previous researchers. With the 20 millimeter perforated plate all of the MFB curves are fairly tightly packed but it can still be seen that the slopes of the curves increase with more advanced ignition timing (fig 4.9). This trend was also reflected in the main combustion durations (5-90% MFB) on table 4.1. Although the turbulence did not decay very much over the short time span covered by the combustion period with this perforated plate, the turbulence intensity apparently decayed enough to affect the main combustion duration. Unfortunately, since the combustion occurred during the pre-relaxation stage, the characterization of the turbulence in this stage could not be well enough defined to substantiate this hypothesis. The FIP with the 20 millimeter perforated plate had no consistent trend. It was thought that the fuel/air mixture temperature at the time of ignition reduced the effect of the turbulence during the FIP. The low temperature at ignition with advanced timing would increase the FIP where the turbulence was highest and the high temperature at ignition with retarded ignition would decrease the FIP where the turbulence was lowest. The effect of ignition timing would have to be eliminated before the effect of turbulence during the FIP could be determined. With the 10 and 5 millimeter perforated plates (figures 4.10, 4.11 and table 4.1) the same trend was observed during the main combustion duration as with the 20 millimeter perforated plate but was more clearly defined because of the larger separation between the MFB curves. The influence of the turbulence intensity was also more pronounced during the FIP because of the wider range of turbulence intensities around the ignition period due to the higher decay rates with the smaller scales. With the 5 millimeter Chapter 4. Discussion of Experimental Results 40 perforated plate the FIP was reduced, presumably by higher turbulence intensity, as the ignition was advanced except for with the most advanced ignition timing. It was thought that the effect of the fuel/air mixture temperature outweighed the effect of the turbulence intensity at this point. The effects of the perforated plate hole size independent of ignition timing can be seen in figures 4.12 to 4.16 where the MFB curves were grouped by ignition timing instead of by perforated plate hole size. These figures show that the main combustion rates (slopes of the curves) increase with increased perforated plate hole size. The reason for this increase was thought to be the higher turbulence intensity with the larger perforated plates. The effect of the turbulence scale on the FIP could now be seen to some extent in the figures since the fuel/air mixture temperature should be the same for each perforated plate with the same ignition timing. Even though the turbulence intensity was higher with the larger perforated plates the FIP was shorter with the smaller perforated plates for some ignition timings, particularly with advanced ignition timings where the small scale turbulence did not have time to decay before ignition. This shows that the small scale turbulence was effective in reducing the FIP even when opposed by the turbulence intensity. Consequently, small scale turbulence generation could be an effective tool in reducing the FIP even if the intensity is lower than if larger scale turbulence had been generated by similar means. 4.2 Effects of Different Combustion Chamber Designs The effects of different combustion chamber designs were studied using the R I C M with the compression stroke alone. In this way the compression stroke generated turbulence of each design could be evaluated without interference from the intake generated turbulence. Both the turbulence generating characteristics and the subsequent effects on combustion Chapter 4. Discussion of Experimental Results 41 of the turbulence generated, amongst the different designs, were of interest. 4.2.1 Compression stroke Generated Turbulence The standard bowl-in-piston design (piston BIP) was the first to be measured so that there was a base from which the "forced squish-jet" designs could be compared. With all of the designs, HWA measurements were taken at the centre and at the edge of the bowl so that entire flow field could be evaluated. The first of the "forced squish-jet" designs to be measured was with piston 2J which had two 2 millimeter wide jet slots on opposing sides of the bowl (see figure 2.3). This configuration was expected to produce jets that would meet at the centre of the bowl close to the spark location. The effects of the jet slot size and the effects of the number of jet slots used were examined with two other "forced squish-jet" designs. Piston 2LJ had two larger jet slots 4 millimeters wide and piston 4J had four jet slots 2 millimeters wide (see figure 2.3). Due to some similarities in the HWA measurements with the three "forced squish-jet" designs it was suspected that the ridge, and not the jet slots, was the main turbulence generator. A fourth "forced squish-jet" design (piston NJ) which had no jet slots (see figure 2.3) was tested to determine the effects of the ridge alone. The flow patterns of all five pistons were evaluated so that the effects of the different design components could be compared. Bowl-in-Piston Measurements HWA measurements were taken at three depths near the edge of the bowl and nt the centre of the bowl for a total of six measurement locations with the standard bowl-in-piston design (piston BIP). The HWA measurements near the edge of the bowl were expected to show the squish velocity. Both the peak value of the squish velocity and the Chapter 4. Discussion of Experimental Results 42 point at which it occurred were of interest. The measurements at the centre of the bowl were at the same points that the spark probe would be located so that the turbulence characteristics at these points could be related to the combustion rates. The data from all six measurement locations were used in the evaluation of the overall flow pattern generated with this combustion chamber design. The mean velocity and turbulence intensity measured at all three depths at the bowl edge is shown in figure 4.17 with respect to crank angle between BDC (180 degrees) and TDC (360 degrees). At TDC the RICM stops so that data collected after TDC can only be viewed with respect to time. Figure 4.18 shows the data with respect to time with TDC corresponding to around the 34 millisecond mark. At a crank angle of around 27 and 35 degrees BTDC there were mean velocity peaks at the 8 and 16 millimeter depths respectively. These points corresponded to the crank angles at which the top of the piston passed the HWA (see Appendix B for calculations of piston position). A short time after these mean velocity peaks there were peaks in the turbulence intensity; most likely from the breakup of the flow over the edge of the bowl. At approximately 15, 10 and 5 degrees BTDC there was a peak in the mean velocity at the 2, 8 and 16 millimeter depths respectively. These peaks corresponded to around the point at which the squish velocity was expected to be a maximum, based on the simple model in Appendix B, but at the 2 millimeter depth the peak also corresponded to the point at which the piston passed the HWA. There was a peak in the turbulence intensity at the 2 millimeter depth a short time later. The turbulence intensity peak at the 2 millimeter depth could have been from the breakup of the squish motion or from the flow over the edge of the bowl. Since the piston velocity was very slow at this point (less than 1 m/s) and the squish motion had just peaked it was more likely that the turbulence intensity peak was from the breakup of the squish motion. After this point the mean velocity and turbulence intensity at all locations decayed with no more distinguishing features. Chapter 4. Discussion of Experimental Results 43 The mean velocity and turbulence intensity measured at all three depths in the centre of the bowl is shown in figure 4.19 with respect to crank angle and in figure 4.20 with respect to time. At the 8 and 16 millimeter depths there was a peak in the mean velocity shortly after the point at which there was a peak in the mean velocity at the bowl edge indicating that the squish motion generated at the bowl edge reached the centre of the bowl. At the 2 millimeter depth there was no significant peak in the mean velocity. Since the squish motion was generated around the entire circumference of the bowl and directed towards the centre it would be expected that the squish motion would either be deflected at the centre or break up into turbulence as it approached the centre. The low mean velocity at the 2 millimeter depth and the large mean velocity at the two lower depths suggested that the squish motion was deflected downwards into the bowl. The peak in the turbulence intensity at the centre of the bowl for all three depths occurred very close to TDC but the magnitude at the 8 millimeter depth was much larger than at the other two depths. If the squish motion were directed downwards into the bowl, the turbulence intensity measurements would be expected to decrease as the measurement depth increased, becoming farther away from the point of generation. This was the case with the lower two measurement depths but at the 2 millimeter depth the deflection of the squish motion was most likely responsible for the low turbulence intensity. The overall trend, based on the HWA measurements, was that the squish motion was generated late during the compression stroke at the edge of the bowl and then moved downwards as it moved towards the centre of the bowl (see figure 4.21). The concentration of the squish motion at the centre of the bowl resulted in the maximum menu velocity measurement being located in the centre of the bowl. The peak turbulence intensity at each measurement location occurred when the mean velocity first started to decay. Chapter 4. Discussion of Experimental Results 44 "Forced Squish-Jet" Measurements The mean velocity and turbulence intensity produced by a jet slot and between the jet slots at the bowl edge of piston 2J for all three depths is shown in figures 4.22 and 4.23 respectively. Piston 2J was the base "forced squish-jet" design and the flow generated with it was typical of the flow generated with all "forced squish-jet" designs. The HWA was aligned with a jet slot (figure 4.22) so that the jet velocity could be measured and was between jet slots (figure 4.23) so that the squish motion could be measured. At the 16 millimeter depth the mean velocity and turbulence intensity in each figure were approximately the same independent of the jet slots. The shape of the curves were also similar to the shape produced with the standard bowl-in-piston design at the same depth (figure 4.17). At the points where the top of the piston passed the HWA there was a peak in the mean velocity for all three cases. Away from the jet slots and with piston BIP the magnitude of the peaks were similar but in front of the jet slot the peak was larger. This indicated the presence of a jet generated by the jet slot. The presence of a jet was also supported by the larger turbulence intensity in front of the jet slot. The second peak in the mean velocity, associated with the peak squish velocity, was independent of the slot alignment but was much larger with the "forced squish-jet" design than with the standard bowl-in-piston design. This indicated that the ridge was affecting the squish velocity more than the jet slots. At the 8 millimeter depth there was a large peak in the mean velocity and turbulence intensity when the HWA was aligned with the jet slot (figure 4.22) indicating that a strong jet had developed at this depth. The peak mean velocity and turbulence intensity was generated at around 20 degrees BTDC which coincided with the point in the compression stroke when the jet slot passed the HWA. At this point the ridge on the piston and the step in the head had already started to trap fluid in the squish area resulting in the forced Chapter 4. Discussion of Experimental Results 45 squish-jet action as expected. Away from the jet slots (figure 4.23) there was no strong peak in the mean velocity or turbulence intensity and the shape and magnitude of the curve was similar to the standard bowl-in-piston results at the same depth. At the 2 millimeter depth the mean velocity and turbulence intensity was not sub-stantially affected by the alignment of jet slot. Although the ridge was the main factor contributing to the motion, there was a slight difference in the crank angle location at which the peaks occurred when the HWA was aligned with the jet slot. The jet slot tended to delay the squish motion which was probably due to the jet slot relieving an early squish action produced by the ridge. Comparing the mean velocity away from the jet slots to that produced by the standard bowl-in-piston design at the same depth it can be seen that the peak was higher and occurred slightly earlier with the "forced squish-jet" design. There was a drastic difference between the two designs when comparing the turbulence intensity at the 2 millimeter depth. With the "forced squish-jet" design there was a large peak in the turbulence intensity at around 20 degrees BTDC which was not present with the standard bowl-in-piston design at this depth although a similar peak was present at the lower depths. This indicated that the peak in the turbulence intensity with the "forced squish-jet" design was caused by a slightly earlier squish action which was redirected towards the top of the bowl by the ridge (see figure 4.24). This caused the turbulence intensity produced from the squish motion to be measured at the 2 millimeter depth where it had previously only been measured below this depth with the bowl-in-piston design. Figure 4.25 shows the mean velocity and turbulence intensity measured at all three depths at the centre of the bowl with piston 2J. The peaks in the mean velocity and turbulence intensity all occur at about 10 degrees B T D C independent of the depth. This was in contrast to the peaks with the standard bowl-in-piston design which occurred at around TDC. The ridge in the piston and not the jet slots were probably responsible for Chapter 4. Discussion of Experimental Results 46 this change in squish timing at the centre of the bowl. The ridge was designed to trap fluid in the squish area so that the fluid would be forced through the jet slots but the fluid apparently found a path of less resistance through the clearance gap. The reduced area through which the squish fluid passed (the clearance gap area) caused an increase in the magnitude of the squish velocity with the "forced squish-jet" design when compared to the standard bowl-in-piston design. This increased velocity caused the squish motion to reach the centre of the bowl earlier and with a larger magnitude. The lower mean velocity at the 2 millimeter depth compared to the 8 and 16 millimeter depths indicated that the squish velocity was directed downwards into the bowl as it approached the centre. This was also supported by the decreased turbulence intensity at the 8 and 16 millimeter depths as with the standard bowl-in-piston design. The effects of increasing the jet slot size and the number of jet slots on the generation of a jet are shown in figure 4.26 for pistons 2J, 2LJ and 4J where the mean velocity measurements were representative of the jet velocity generated by the different jet slot configurations. The mean velocity curves in all three cases were quite similar in timing and shape but varied in magnitude. The magnitude of the mean velocity and turbulence intensity increased when the slots were larger (piston 2LJ) and decreased when there were more jet slots (piston 4J). The larger jet slots produced a stronger jet and a signif-icantly higher turbulence intensity, as was the case when the perforated plate hole size was increased in the previous set of experiments (see section 4.1). The increased number of jet slots produced a weaker jet and a slightly lower and later turbulence intensity com-pared to the base case although the difference in magnitude could have been caused by some measurement discrepancies. The reason for the discrepancy was most likely n mea-surement error associated with the HWA. The HWA wire broke after the measurements with the bowl-in-piston design and the first "forced squish-jet" designs (piston 2J). A l l of the measurements made with the first wire were a little higher than the measurements Chapter 4. Discussion of Experimental Results 47 made with the replacement wire. This can be seen by looking at the early section of the mean velocity plot before there was any effect from the combustion chamber design (see figure 4.27). For the measurements taken with the first wire (piston BIP and 2J) there was a small mean velocity measured and with the replacement wire (pistons 2LJ, 4J and NJ) there was no mean velocity measured. The reason for this inaccuracy was thought to be related to the calibration of the two wires which was somewhat dependent on the selection of calibration points used to determine the constants A, B and n, especially at low velocities where natural convection can be an increased source of error. A comparison between the mean velocity and turbulence intensity measured at the 2 millimeter depth away from the jet slots with pistons BIP, 2J, 4J and NJ is given in figure 4.27. The effect of using the two different wires with the HWA was particularly clear in this figure between 250 and 330 degrees crank angle where the mean velocities measured with the first wire was consistently larger than with the replacement wire for no other apparent reason. The magnitude of the mean velocity peaks were larger with the "forced squish-jet" designs than with the bowl-in-piston design particularly when the problem with the HWA was taken into account. This indicated that the magnitudes were related to the ridge as well as the jet slots. The jet slots tended to relieve the squish effects of the ridge and produce a lower mean velocity with an increase in number or size of the jet slots. This trend can more clearly be seen by looking at the turbulence intensity measurements. The first peak in the turbulence intensity with all the "forced squish-jet" designs was probably caused by the restriction in the flow area as the ridge started to mesh with the step in the head. The second peak with the "forced squish-jet" designs and the only peak with the standard bowl-in-piston design was most likely associated with the later squish motion which was not related to the ridge. The maximum second peak was produced when there were no jet slots in the ridge (piston NJ) so that all the fluid behind the ridge was forced through the clearance space between the ridge in the Chapter 4. Discussion of Experimental Results 48 piston and the step in the head. When the jet slots were present (pistons 2J, 4J) and when there was no ridge (piston BIP) the turbulence intensity was considerably lower at this point presumably because there was more space for the fluid to escape through. Comparing the mean velocity and turbulence intensity measured away from the jet slots at the 8 millimeter depth (figure 4.28), with all the combustion chamber designs, there was only a minimal difference caused by the different designs, particularly in the turbulence intensity. The differences in the generation of squish motion near the edge of the bowl, with the different designs, were limited to the region near the top of the bowl where the squish was most effective. The mean velocity and turbulent intensity for all the pistons at the centre of the bowl is shown in figures 4.29, 4.30 and 4.31 at depths of 2, 8 and 16 millimeters respectively. At the 16 millimeter depth (figure 4.31) there was a definite increase in the mean velocity with the "forced squish-jet" designs compared to the standard bowl-in-piston design. The peak in the turbulence intensity also occurred earlier with all the "forced squish-jet" designs as was found with piston 2J. The magnitudes of the mean velocity peaks with the "forced squish-jet" designs were in an order consistent with the earlier finding that the ridge was the main factor contributing to the increase in squish motion and that the jet slots primarily reduced the effectiveness of the ridge by providing an alternate flow path for the fluid behind the ridge. The one exception from this trend, based on the magnitude of the peaks, was with the first "forced squish-jet" design measured (piston 2J) which was the only "forced squish-jet" design measured using the first wire. Consequently, the magnitude of the peak with piston 2J was expected to be high compared to the magnitudes measured with the other "forced squish-jet " designs because of the wire used. At the 2 and 8 millimeter depths (figures 4.17 and 4.18) the same trends amongst the designs were continued but with a lower mean velocity due to the downwards deflection Chapter 4. Discussion of Experimental Results 49 of the squish motion as it approached the centre of the bowl (see figure 4.24). With most of the "forced squish-jet" designs, the turbulence intensity at the 8 millimeter depth was slightly higher than at the other two depths indicating that this was the depth at which the squish velocity came together. Above the 8 millimeter depth the low mean velocity due to the presence of the wall contributed to the low turbulence intensity while below this depth the decay of the turbulence caused the lower turbulence intensity. The best compromise between maximum mean velocity and the decay of the turbulence was reached at the 8 millimeter depth. 4.2.2 Combustion with Compression Stroke Generated Turbulence The MFB curves and combustion durations were calculated from pressure and crank angle measurements (see section 3.2) taken with the spark probe at three different depths and with ignition timings of 20, 30 and 40 degrees BTDC. The tests were conducted over this range of conditions so that the maximum performance with each design could be found for an unbiased comparison. The ignition timings used were limited to this range because it was the range in which ignition normally occurs in a spark ignition engine. Bowl-in-Piston Measurements The combustion durations for piston BIP are listed in table 4.2 for all spark probe depths and ignition timings tested. In all cases the FIP (0-5% MFB) was quite long indicating that the fuel/air mixture was burning at the laminar rate well past the time of ignition. Looking back at the HWA measurements (figures 4.19 and 4.20) this was confirmed by the very low turbulence intensity before around 10 degrees BTDC corresponding to 30 milliseconds after BDC. During the experimental measurements it was noticed from the pressure traces on the oscilloscope that the ignition timing had little effect on the total time between the Chapter 4. Discussion of Experimental Results 50 triggering of data collection and the end of combustion. When the time from BDC (point at which data collected was triggered) to the point of ignition was added to the total combustion duration (0-90% MFB), for each ignition timing at a single spark probe depth, the results confirmed the observation (see table 4.3). This indicated that the fuel/air mixture was burned at a laminar rate until the turbulence intensity started to increase at which point the fuel/air mixture rapidly burned (at a turbulent rate). The effects on combustion of the spark probe depth with piston BIP can be seen in figures 4.32, 4.33 and 4.34 for ignition timings of 20, 30 and 40 degrees BTDC respectively. The differences between the combustion durations at the different depths were quite small with an ignition timing of 20 degrees BTDC but the differences increased as the ignition timing was advanced. This increase was evidently due to the longer time available for the flame to develop at a laminar rate with advanced ignition timing before the turbulence intensity rapidly increased the burning rate. During the early "laminar" burning stage the flame kernel developed much quicker when the spark probe was located at the 16 millimeter depth (see figure 4.34) than at the other two depths. This trend could have been caused by the restriction in the growth of the flame kernel by the head when the flame was developing close to the head. Or this trend could have been associated with the larger mean velocity measurements at the lower depths (see figure 4.19) but in any case the combustion rate was not increased enough compared to the increase due to turbulence to be of much importance. The transition from the slow "laminar" burning stage to the rapid "turbulent" burning stage was coincident at each depth with the point where the turbulence intensity started to increase. Once the "turbulent" burning stage had developed, the magnitude of the turbulence intensity measured at different depths in the HWA experiments did not have much influence on the combustion rate (slope of the MFB curve). The reason for this poor correlation between the turbulence intensity at the spark location and the main burning Chapter 4. Discussion of Experimental Results 51 rate was because the burning rate was dependent on the overall turbulence intensity and not just the turbulence intensity at one point. Figure 4.19 shows that the turbulence intensity at the 8 millimeter depth was the first to increase; quickly followed by the turbulence intensity at the 16 millimeter depth and a short time later followed by the turbulence intensity at the 2 millimeter depth. For the two most advanced ignition timings (figures 4.33 and 4.34) the "turbulent" burning stage also started first when the spark probe was at the 8 millimeter depth followed by the 16 millimeter depth and then the 2 millimeter depth. This indicated that the turbulence intensity at the spark probe was controlling the transition from "laminar" to "turbulent" burning rate. When the ignition timing was the most retarded (figure 4.32) this trend was no longer found. For this case the flame was not fully developed (ie not past the FIP) when the turbulence was encountered so the influence of the turbulence would not be expected to immediately change the burning rate. For this retarded ignition timing the 2 millimeter spark probe depth produced the earliest "turbulent" burning stage because, by the time the flame kernel had developed, the turbulence intensity at the 2 millimeter depth was just developing while at the other two depths the turbulence was already decaying. The turbulence that was just developing at the 2 millimeter depth was more likely to have smaller turbulence scales because it had not decayed yet, and thus be more effective in decreasing the FIP. "Forced Squish-Jet" Measurements The combustion durations for piston 2J are listed in table 4.4 for all spark probe depths and ignition timings tested. The total time from BDC to the end of combustion was again independent of ignition timing (see table 4.5) indicating that the flow field was controlling the combustion rate as before. The MFB curves for the three spark probe Chapter 4. Discussion of Experimental Results 52 depths with piston 23 are shown in figures 4.35, 4.36 and 4.37 for ignition timings of 20, 30 and 40 degrees BTDC respectively. The FIP's at all ignition timings were shortest with a spark probe depth of 2 mil-limeters. This trend could have been caused by the redirecting of the squish motion towards the top of the bowl by the ridge. The turbulence nearest the generating source (near the top of the bowl) would be expected to have a smaller scale than the turbulence farther down stream which had had more time to decay. The smaller scale turbulence would reduce the FIP because it would become effective with a smaller flame kernel as suggested by Winsor and Patterson [19] and supported by Checkel [1]. Although the main combustion durations were all very close and quite short, they were consistently shorter with a spark probe depth of 8 millimeters with piston BIP. This was evidently associated with the more central ignition source and the turbulence intensity at that central point. At the 2 millimeter spark probe depth the main combustion duration was likely affected by the distance the flame would have to travel to engulf the chamber. At the 16 millimeter spark probe depth the combustion duration was apparently affected more by the lower level and late timing of the turbulence intensity (see figure 4.25) than by the distance that the flame had to travel. These two factors worked against each other to produced the best compromise at the 8 millimeter spark probe depth. The combustion durations for pistons 2LJ, 4J and NJ are listed in tables 4.6, 4.7 and 4.8, respectively, for all the spark probe depths and the limited number of ignition timings tested. Since the combustion duration was mainly controlled by the generation of turbulence by the different piston designs, the full range of tests were only conducted with an ignition timing of 40 degrees BTDC. The most advanced timing was chosen so that the flame kernel would be fully developed by the time the turbulence was generated. The comparisons between the MFB curves for all the piston designs at an ignition timing of 40 degrees BTDC are shown in figures 4.38, 4.39 and 4.40 for spark probe Chapter 4. Discussion of Experimental Results 53 depths of 2, 8 and 16 millimeters respectively. In all cases the bowl-in-piston design was the slowest to increase the combustion rate so that it was clearly the least effective design. The FIP with the "forced squish-jet" designs were all very similar because of the similarities found with the turbulence measurements. The main variation with the "forced squish-jet" designs was found in the final stages of combustion. One important factor controlling the final stages of the MFB curve was the determination of the end of combustion. In this work the end of combustion was defined as the point at which there was no longer a pressure rise due to combustion. Since the MFB curve was normalized based on this pressure at the "end" of combustion, the final stages of the MFB curve could have been drastically shifted up or down by the selection of the end point. The determination of the end point was particularly poor when the slope of the MFB curve was low. The reason for the long tail in the MFB curves for some designs was believed to be related to the fuel/air mixture being trapped in the clearance volume at TDC. Since the piston of the RICM stopped at TDC the fuel/air mixture trapped in the 1 millimeter clearance gap above the squish area would be slow to burn. The slow burning of this trapped fuel/air mixture would produce the long tails found in the MFB curves. Many factors including the gap size, the gap shape, the heat transfer rate from the flame, the pressure and temperature in the chamber and the exact fuel/air ratio would affect the end of combustion with the different piston designs. The variation in these factors was believed to be the source of the differences found during the final stages of combustion with the different pistons. If these piston designs were tested in a real engine the differ-ences due to the clearance gap would be eliminated and the final stages of combustion would most likely be very similar for all the "forced squish-jet" designs. The MFB curves for the different piston designs at an ignition timing of 30 and 20 degrees BTDC are shown in figures 4.41 to 4.46 for all the spark probe depths. Because Chapter 4. Discussion of Experimental Results 54 the rate of combustion with the first few piston designs was found to be controlled by the turbulence and not by the ignition timing not all pistons were tested at these more retarded ignition timings. The MFB curves that were produced at these two ignition timings show the same trends as did the MFB curves with advanced timing. Chapter 5 Conclusions and Recommendations 5.1 Introduction The objective of this research was to investigate the effects of turbulence on combus-tion and how turbulence can be generated to achieve fast combustion in a spark ignition engine. The focus of the work was separated into two parts. In the first part, perfo-rated plates were used in the rapid intake and compression machine (RICM) to generate different turbulence scales and intensities so that the best combination could be deter-mined. In the second part, the compression stroke turbulence generating characteristics of several "forced squish-jet" combustion chamber designs and the resultant effect on combustion were investigated and compared to a standard bowl-in-piston design. All of the experiments were conducted in the RICM with a stoichiometric methane/air mixture being used in the combustion runs. 5.2 Conclusions 5.2.1 Perforated Plate Characterization of Turbulence • The turbulence produced by the perforated plates in the R I C M became relaxed (characterized by homogeneous isotropic turbulence) a short time after the piston stopped at TDC. 55 Chapter 5. Conclusions and Recommendations 56 • The turbulence intensity was highest when the mean velocity was rapidly decaying. • In the pre-relaxation stage the turbulence scale was characteristic of the perforated plate hole size. • The turbulence intensity was generally higher when larger scale turbulence was generated, all other factors being the same. • Turbulence characterized by a larger scale decayed more slowly, although the effects of the turbulence intensity on the decay rate could not be separated from the effects of scale. • In the relaxed stage the turbulence decay rate implies that the turbulence scale was characteristic of the combustion chamber dimensions. Effects on Combustion • The main combustion duration was decreased by an increase in the turbulence intensity. • The flame initiation period (FIP) was decreased by the generation of smaller scale turbulence even when the generation of the small scale turbulence resulted in a lower turbulence intensity than if large scale turbulence was generated. 5.2.2 Different Combustion Chamber Designs Compression Stroke Generated Turbulence • With the standard bowl-in-piston design the squish motion was generated late in the compression stroke at around 10 degrees BTDC. er 5. Conclusions and Recommendations 57 With the standard bowl-in-piston design the squish motion was deflected down-wards as it approached the centre of the bowl. The peak turbulence intensity at each measurement location with the standard bowl-in-piston design occurred when the mean velocity first started to break up at TDC. Jets were developed near the edge of the bowl by the jet slots and were a maximum at around the 8 millimeter depth at 20 degrees BTDC. The "forced squish-jet" design produced a slightly earlier squish motion which was directed towards the top of the bowl by the ridge at around 20 degrees BTDC. At the centre of the bowl the "forced squish-jet" designs produced an earlier mean velocity and turbulence intensity peak at around 10 degrees BTDC because of the higher squish velocity generated by the ridge at the edge of the bowl. The squish motion was also directed downwards as it approached the centre of the bowl with the "forced squish-jet" designs. An increase in the jet slot size increased the mean velocity of the jet and significantly increased the turbulence intensity produced by the jet slots. An increase in the number of jet slots reduced the mean velocity of the jet and slightly reduced the turbulence intensity produced by the jet slots. The ridge produced the major increase in motion between the standard bowl-in-piston design and the "forced squish-jet" designs while the jet slots reduced the effectiveness of the ridge. Chapter 5. Conclusions and Recommendations 58 • The maximum turbulence intensity at the centre of the bowl was measured at the 8 millimeter depth at around 10 degrees BTDC for most "forced squish-jet" designs. Effects on Combustion • The burning rate was rapidly increased by the generation of turbulence with the different combustion chamber designs. • The rapid (turbulent) combustion rate could not be related to single point turbu-lence measurements because it was dependent on the overall turbulence. • The transition from slow (laminar) combustion to rapid (turbulent) combustion was controlled by the timing of the turbulence generated and not by the ignition timing once the flame kernel had developed. • The "forced squish-jet" designs increased the combustion rate earlier than the stan-dard bowl-in-piston design because of the timing of the turbulence generated. • The differences in the main combustion rate between the different "forced squish-jet" designs were not related to the jet slots but to the variations in clearance volume at TDC. 5.3 Recommendations 5.3.1 Perforated Plate It is suggested that further work be conducted in order to better characterize the tur-bulence generated by the perforated plates in the RICM. The optical access to the com-bustion chamber provides an ideal setting for a full range of LDA measurements. The determination of the length scale with each perforated plate hole size should be included Chapter 5. Conclusions and Recommendations 59 in the measurement scheme along with an accurate determination of the turbulence in-tensity. More combustion experiments should also be run so that the effects on the combustion of the turbulence parameters can be determined. Special attention should be given to the FIP which has been shown to be particularly sensitive to turbulence scale. Possibly, an ideal combination of scale and intensity could be found which best reduces the FIP. 5.3.2 Different Combustion Chamber Designs It is suggested that the "forced squish-jet" design be refined so that strong jets can be produced without excessive flow over the ridge. Both a smaller clearance gap and larger jet slots could be used. The effects of the ridge alone in generating turbulence during the compression stroke should also be investigated. LDA measurements would also be useful in determining the flow direction for analysis of both the standard bowl-in-piston design and new designs. The effects of intake generated turbulence on the development of compression gen-erated turbulence could also be examined using the side valve designed for the RICM along with LDA measurements. The problem with the long tail produced on the MFB curves should be investigated in either a real engine or by adjusting the stroke of the RICM so that the piston stops after TDC. In this way the small gap which trapped some fuel/air mixture could be eliminated. Care should be taken if the piston of the RICM is stopped after TDC so that an excessive torque is not placed on the crankshaft. References M. David Checkel, "Turbulence-Enhanced Combustion of Lean Mixtures", Ph.D. Thesis, University of Cambridge, 1981. Gordon J. Van Wylen and Richard E. Sonntag, "Fundamentals of Classical Ther-modynamics", Second Edition (John Wiley &: Sons), 1978. M. P. Halstead, D. B. Pye, and C. P. Quinn, "Laminar Burning Velocity and Weak Flammability Limits Under Engine-Like Conditions", Combustion and Flame, Vol. 22, 1974. James N. Mattavi, "The Attributes of Fast Burning Rates in Engines", SAE 800920, 1980. C. D. Cameron, "An Investigation of Squish Generated Turbulence in I. C. Engines", M.A.Sc. Thesis, University of British Columbia, Report AFL-85-02, 1985. G. Damkohler, "The Effects of Turbulence on the Flame Velocities in Gas Mixtures", NACA T M 1112, 1947. N. C. Blizard and J. C. Keck, "Experimental and Theoretical Investigation of Tur-bulent Burning Model for Internal Combustion Engines", SAE 740191, 1974. Rodney J. Tabaczynski, Colin R. Ferguson, and Krisna Radnakrishnan, "A Turbu-lent Entrainment Model for Spark Ignition Engine Combustion",SAE 770647, 1977. H. Tennekes, "Simple Model for the Small- Scale Structure of Turbulence", Physics of Fluids, Vol. 11, No. 3, 1968. Klaus Dohring, "The Relative Effects if Intake and Compression Stroke Generated Turbulence on I. C. Engine Combustion Duration", M.A.Sc. Thesis, University of British Columbia, Report AFL-86-01, 1986. T. H. Ma, "Effects of Cylinder Charge Motion on Combustion", I. Mech. E. paper C81/75, 1975. Peter 0. Witze and Fernando R. Vilchis, "Stroboscopic Laser Shadowgraph Study of the Effect of Swirl on Homogeneous Combustion in a Spark-Ignition Engine", SAE 810226,1981. Isao Nagayama, Yashushi Araki, and Yasuo Iioka, "Effects of Swirl and Squish on S. I. Engine Combustion and Emission", SAE 770217, 1977. R. Dymala-Dolesky "The Effects of Turbulence Enhancement on the Performance of a Spark-Ignition Engine", M.A.Sc. Thesis, University of British Columbia, Report AFL-86-09, 1986. 60 References 61 E. S. Semenov, "Studies of Turbulent Gas Flow in Piston Engines", NASA Technical Translation F97, 1963. M. Horvatin and A. W. Hussmann, "Measurement of Air Movements In Internal Combustion Engine Cylinders", DISA Information, No. 8, 1969. H. Hassan and J. C. Dent, "The Measurement of Air Velocity in a Motored Internal Combustion Engine Using a Hot-Wire Anemometer", Proc. I. Mech. E., Vol. 185 50/71, 1971. S. Ohigashi, Y. Hamamoto, and A. Kizima, "Effects of Turbulence on Flame Prop-agation in Closed Vessel", Bulletin of the JSME, Vol. 14, No 74, 1971. Richard E. Winsor and Donald J. Patterson, "Mixture Turbulence — A Key to Cyclic Combustion Variation", SAE 730086, 1973. M. Tsuge, H. Kido, and H. Yanagihara, "Decay of Turbulence in a Closed Vessel", Bulletin of the JSME, Vol. 16, No. 92, 1973. J. C. Dent and N. Salama, "Turbulence Structure in the Spark Ignition Engine", I. Mech. E. paper C83/75, 1975. G. E. Andrews, D. Bradley, and S. B. Lwakabamba, "Turbulence and Turbulent Flame Propagation — A Critical Appraisal", Combustion and Flame, Vol. 24, 1975 David R Lancaster, "Effects of Engine Variables on Turbulence in a Spark-Ignition Engine", SAE 760159, 1976. David R. Lancaster, Roger B. Krieger, Spencer C. Sorenson, and William L. Hull, "Effects of Turbulence on Spark-Ignition Engine Combustion", SAE 760160, 1976 Rodney J. Tabaczynski, " Turbulence and Turbulent Combustion in Spark-Ignition Engines", Prog. Energy Combustion Science, Vol. 2, 1976. Peter O. Witze, "Measurements of the Spatial Distribution and Engine Speed De-pendence of Turbulent Air Motion in an I. C. Engine", SAE 770220, 1977. Rodney B. Rask, "Laser Doppler Anemometer Measurements in an Internal Com-bustion Engine", SAE 790094, 1979. J. B. Cole and M. D. Swords, "An Investigation of the Ignition Process in a Lean-Burn Engine Using Conditionally Sampled laser Doppler Anemometry", SAE 800043, 1980. Peter 0. Witze, "A Critical Comparison of Hot-Wire Anemometrv and Laser Doppler Velocimetry for I. C. Engine Applications", SAE 800132, 1980. Rodney B. Rask, "Comparison of Window, Smoothed-Ensemble, and Cycle-by-Cycle Data Reduction Techniques for Laser Doppler Anemometer Measurements of In-Cylinder Velocity", Symposium of Fluid Mechanics of Combustion Systems, ASME FED Spring Meeting, 1981. References 62 T. M. Liou and D. A. Santavicca, "Cycle Resolved Turbulence Measurements in a Ported Engine With and Without Swirl", SAE 830419, 1983. Jay K. Martin, Peter 0. Witze, and Claus Borgnakke, "Combustion Effects on the Preflame Flow Field in a Research Engine", SAE 850122, 1985. Frederic A. Matekunas, "Modes and Measures of Cyclic Combustion Variability", SAE 830337, 1983. A. E. Catania and A. Mittica, "A Contribution to the Definition and Measurement of Turbulence in a Reciprocating I. C. Engine", ASME 85-DGP-12, 1985. R. A. Fraser, P. G. Felton, F. V. Bracco, and D. A. Santavicca, "Preliminary Tur-bulence Length Scale Measurements in a Motored I. C. Engine", SAE 860021, 1986. M. J. Hall and F. V. Bracco, "A Study of Velocities and Turbulence Intensities Measured in Firing and Motored Engines", SAE 870453, 1987. R.S. Benson and R. Pick, "Recent Advances in Internal Combustion Engine Instru-mentation with Particular Reference to High Speed Data Acquisition and Automated Test Bed", SAE 740695, 1974. Ronald Jay Pierik, "Swirling Combustion of Premixed Gaseous Reactants in a Short Cylindrical Chamber", M.A.Sc. Thesis University of British Columbia, 1987. L. V. King, "On the Convection of Heat from Small Cylinders in a Stream of Fluid: Determination of the Convective Constants of Small Platinum Wires with Applica-tions to Hot-Wire Anemometry", Proc. Roy. Soc, Vol. 214a, No. 14, 1914. D. C. Collis and N. J. Williams, "Two- Dimensional Convection from Heated Wires at Low Reynolds Numbers", Journal of Fluid Mech., Vol. 6, 1959. P. O. A. L. Davies and M. J. Fisher, "Heat Transfer from Electrically Heated Cylin-ders", Proc. R. Soc. A., Vol. 280, 1964. Peter 0. Witze, Jay K. Martin and Claus Borgnakke, "Measurements and Predic-tions of the Precombustion Fluid Motion and Combustion Rates in a Spark Ignition Engine", SAE 831697, 1983. Esther C. Tippett, "The Effects of Combustion Chamber Design on Turbulence, Cyclic Variation and Performance in an SI Engine", M.A.Sc. Thesis, University of British Columbia, 1989. G. I. Taylor, "The Spectrum of Turbulence", Proc. Roy. Soc. of London, Vol. A 164, 1938. John B. Heywood, "Fluid Motion Within the Cylinder of Internal Combustion En-gines — The 1986 Freeman Scholar Lecture", Journal of Fluids Engineering, Vol. 109, 1987. G.M. Rassweiler and L. Withrow, "Motion Pictures of Engine Flames Correlated with Pressure Cards", SAE Journal (Trans), Vol. 42, No. 5, 1938. References 63 [47] G. G. Brown, E. H. Leslie and J. V. Hunn, "Gaseous Explosions", Industrial and Engineering Chemistry, Vol. 17, 1925. [48] Michael B. Young, "Cyclic Dispersion — Some Quantitative Cause-and-Effect Re-lationships", SAE 800459, 1980. [49] R. B. Krieger and G. L. Borman, "The Computation of Apparent Heat Release for Internal Combustion Engines", ASME 66-WA/DGP-4, 1966. [50] Charles A. Amann, "Cylinder-Pressure Measurement and Its Use in Engine Re-search", SAE 852067, 1985. [51] M. B. Young and J. H. Lienesch, "An Engine Diagnostic Package (EDPAC) — Software for Analyzing Cylinder Pressure-Time Data", SAE 780967, 1978. [52] Anil Kapil, "Cycle-to-Cycle Variations in Spark-Ignition Engines", M.A.Sc. Thesis, University of British Columbia, Report AFL-88-01, 1988. Tables 64 Combustion Chamber: Bore 101.6 mm Stroke 100.0 mm Compression Ratio 8 : 1 Initial Pressure (1-Stroke) 50.8 kPa Initial Pressure (2-Stroke) Ambient Bowl-in-Piston Squish Area 75 % Intake Valve: Valve Opens 24° ATDC Valve Closes 20° ABDC Maximum Lift 9.4 mm Driving Cylinder Pressure 900 kPa Simulated Revolutions 1000 rpm Table 2.1: Combustion Chamber Specifications and RICM Running Parameters Probe: TSI Model 1226 No. 44313 Wire: Material TSI Platnum Iridium, PI2.5 Diameter 6.3 fim Thermal Coeff. of Resistance 0.0009/°C Length 1.5 mm Operating Temperature 600° C Bridge: DISA Type 55M10 CTA Standard Bridge Filter: Type DISA Type 55D25 Auxiliary Unit Low Pass Setting 20 kHz High Pass Setting Off Table 2.2: Hot-Wire Anemometer Specifications and Equipment Tables 65 Piston HWA Measurement Location at the centre at bowl edge at bowl edge of the bowl between jet slots lined-up with jet slot 2 mm 8 mm 16 mm 2 mm 8 mm 16 mm 2 mm 8 mm 16 mm BIP • • • • • • — — — 2J • • • • • • • • • 2LJ • • • • • 4J • • • • • • • NJ • • • • • — — — Table 2.3: HWA Measurement Locations Used with the Different Pistons Piston Spark Probe Depths and Ignition Timings 40 degrees BTDC 30 degrees BTDC 20 degrees BTDC 2 mm 8 mm 16 mm 2 mm 8 mm 16 mm 2 mm 8 mm 16 mm BIP 7 10 9 8 10 8 10 10 7 2J 10 10 8 10 10 10 6 5 10 2LJ 10 10 5 10 9 10 4J 10 10 10 2 2 2 2 2 NJ 10 10 8 4 Table 2.4: Number of Successful Combustion Runs with Different Pistons Tables 66 5 millimeter Perforated Plate Hole Size Spark Timing 0-5% MFB 5-90% MFB MFB No. of Degrees Time cr <T Time Runs BTDC (ms) (ms) (ms) (ms) (ms) (ms) 0 33.408 4.145 .523 10.314 .911 14.149 10 10 29.168 3.178 .527 7.017 .787 10.195 10 20 26.848 3.002 .355 6.195 .578 9.197 10 30 25.168 2.038 .731 5.552 .678 7.590 10 40 23.488 3.221 .341 4.836 .256 8.057 10 10 millimeter Perforated Plate Hole Size Spark Timing 0-5% MFB 5-90% MFB MFB No. of Degrees Time <r Time Runs BTDC (ms) (ms) (ms) (ms) (ms) (ms) 0 33.520 2.630 .251 6.745 .394 9.375 10 10 29.136 2.534 .269 5.045 .307 7.579 10 20 26.864 3.020 .399 4.133 .210 7.173 10 30 25.136 2.928 .310 3.662 .187 6.590 10 40 23.424 3.501 .488 3.388 .270 6.889 10 20 millimeter Perforated Plate Hole Size Spark Timing 0-5% MFB 1 5-90% MFB MFB No. of Degrees Time (T V> Time Runs BTDC (ms) (ms) (ms) (ms) (ms) (ms) 0 33.792 1.876 .228 4.650 .252 6.526 10 10 29.312 2.657 .250 3.606 .093 6.263 10 20 26.992 2.419 .143 3.327 .066 5.746 10 30 25.056 2.386 .289 3.375 .385 5.761 10 40 23.520 2.968 .233 2.995 .154 5.963 10 Table 4.1: Combustion Durations with Perforated Plates Tables 67 Bowl-in-piston Design with 2 millimeter Spark Probe Depth Spark Timing 0-5% MFB 5-90% MFB MFB No. of Degrees Time Runs BTDC (ms) (ms) (ms) (ms) (ms) 20 4.868 .260 3.178 .991 8.046 10 30 7.584 .846 2.915 .199 10.499 8 40 8.957 .367 3.385 .888 12.342 7 Bowl-in-piston Design with 8 millimeter Spark Probe Depth Spark Timing 0-5% MFB 5-90% MFB MFB No. of Degrees <r <r Time Runs BTDC (ms) (ms) (ms) (ms) (ms) 20 5.072 .114 2.763 .081 7.835 10 30 6.340 .230 2.936 .162 9.276 10 40 6.571 .460 4.113 .429 10.684 10 Bowl-in-piston Design with 16 millimeter Spark Probe Depth Spark Timing 0-5% MFB 5-90% MFB MFB No. of Degrees tr er Time Runs BTDC (ms) (ms) (ms) (ms) (ms) 20 5.222 .214 3.033 .194 8.255 7 30 5.613 .265 3.782 .187 9.395 8 40 4.359 2.571 6.535 2.536 10.894 9 Table 4.2: Combustion Durations with Piston BIP Tables 68 Bowl-in-piston Design with 2 millimeter S 3ark Probe De pth Spark Degrees BTDC Timing Time to Spark (ms) Combustion Duration (ms) Total Time (ms) No. of Runs 20 30 40 27.056 25.400 23.977 8.046 10.499 12.342 35.102 35.895 36.319 10 8 7 Bowl-in-piston Design with 8 millimeter S 3ark Probe De pth Spark ' Degrees BTDC Timing Time to Spark (ms) Combustion Duration (ms) Total Time (ms) No. of Runs 20 30 40 27.136 25.216 23.536 7.835 9.276 10.684 34.971 34.492 34.220 10 10 10 Bowl-in-piston Design with 16 millimeter Spark Probe Depth Spark Timing Combustion Total No. of Degrees Time to Duration Time Runs BTDC Spark (ms) (ms) (ms) 20 27.086 8.255 35.341 7 30 25.040 9.395 34.435 8 40 23.556 10.894 34.450 9 Table 4.3: Sum of Time to Spark Plus Combustion Duration with Piston BIP Tables 69 "Forced Squish-Jet" Design 2J with 2 millimet er Spark Probe Depth Spark Timing 0-5% MFB 5-90% MFB MFB No. of Degrees a a Time Runs BTDC (ms) (ms) (ms) (ms) (ms) 20 2.076 .201 2.063 .241 4.138 6 30 3.590 .109 2.105 .184 5.695 10 40 5.150 .165 2.244 .223 7.394 10 "Forced Squish-Jet" Design 2J with 8 millimeter Spark Probe Depth Spark Timing 0-5% MFB 5-90% MFB MFB No. of Degrees <T <r Time Runs BTDC (ms) (ms) (ms) (ms) (ms) 20 2.440 .235 2.036 .183 4.476 5 30 4.120 .195 2.013 .099 6.133 10 40 5.797 .157 2.000 .092 7.797 10 "Forced Squish-Jet" Design 2J with 16 millimeter Spark Probe Depth Spark Timing 0-5% MFB 5-90% MFB MFB No. of Degrees a <r Time Runs BTDC (ms) (ms) (ms) (ms) (ms) 20 2.486 .293 2.245 .228 4.731 10 30 4.325 .217 2.277 .179 6.602 10 40 5.939 .156 2.474 .149 8.413 8 Table 4.4: Combustion Durations with Piston 2J Tables 70 "Forced Squish-Jet" Design 2J with 2 millimeter Spark Probe Depth Spark Timing Combustion Total No. of Degrees Time to Duration Time Runs BTDC Spark (ms) (ms) (ms) 20 27.280 4.138 31.418 6 30 25.376 5.695 31.071 10 40 23.632 7.394 31.026 10 "Forced Squish-Jet" Design 2J with 8 millimeter Spark Probe Depth Spark Timing Combustion Total No. of Degrees Time to Duration Time Runs BTDC Spark (ms) (ms) (ms) 20 27.008 4.476 31.484 5 30 25.264 6.133 31.397 10 40 23.532 7.797 31.429 10 "Forced Squish-Jet" Design 2J with 16 millimeter Spark Probe Depth Spark Timing Combustion Total No. of Degrees Time to Duration Time Runs BTDC Spark (ms) (ms) (ms) 20 26.784 4.731 31.515 10 30 25.264 6.602 31.866 10 40 23.640 8.413 32.053 8 Table 4.5: Sum of Time to Spark Plus Combustion Duration with Piston 2J Tables "Forced Squish-Jet" Design 2LJ with 2 millimeter Spark Probe Depth Spark Timing 0-5% MFB 5-90% MFB MFB No. of Degrees cr cr Time Runs BTDC (ms) (ms) (ms) (ms) (ms) 20 30 3.237 .076 2.443 .141 5.680 10 40 4.891 .157 2.569 .154 7.460 10 "Forced Squish-Jet" Design 2LJ with 8 millimeter Spark Probe Depth Spark Timing Degrees BTDC 0-5% MF f1 (ms) B cr (ms) 5-90% M (ms) FB cr (ms) MFB Time (ms) No. of Runs 20 30 40 3.563 5.273 .185 .189 ' 2.275 2.170 .170 .169 5.838 7.443 9 L 1 0 "Forced Squish-Jet" Design 2LJ with 16 millimeter Spark Probe De pth Spark Timing Degrees BTDC 0-5% MF (ms) B (ms) 5-90% M (ms) FB cr (ms) MFB Time (ms) No. of Runs 20 30 40 3.982 5.675 .248 .069 2.184 2.164 .150 .019 6.166 . 7.839 10 5 Table 4.6: Combustion Durations with Piston 2LJ Tables 72 "Forced Squish-Jet" Design 4J with 2 millimeter Spark Probe Dept] I Spark Timing 0-5% MFB 5-90% MFB MFB No. of Degrees cr cr Time Runs BTDC (ms) (ms) (ms) (ms) (ms) 20 5.660 .186 4.371 .422 10.031 2 30 5.568 .134 3.786 .113 9.354 2 40 5.455 .088 3.613 .126 9.068 10 "Forced Squish-Jet" Design 4J with 8 millimet er Spark Probe Depth Spark Timing 0-5% MFB 5-90% MFB MFB No. of Degrees cr cr Time Runs BTDC (ms) (ms) (ms) (ms) (ms) 20 5.791 .065 4.147 .379 9.938 2 30 5.588 .003 3.578 .017 9.166 2 40 5.563 .132 3.498 .223 9.061 10 "Forced Squish-Jet" Design 4J with 16 millimeter Spark Probe Depth Spark Timing 0-5% MFB 5-90% MFB MFB No. of Degrees V- cr r1 cr Time Runs BTDC (ms) (ms) (ms) (ms) (ms) 20 30 40 6.199 5.788 .145 .398 2.936 2.849 .321 .395 9.135 8.637 2 10 Table 4.7: Combustion Durations with Piston 4J Tables 73 "Forced Squish-Jet" Design NJ with 2 millimeter Spark Probe Depth Spark Timing 0-5% MFB 5-90% MFB MFB No. of Degrees cr Time Runs BTDC (ms) (ms) (ms) (ms) (ms) 20 30 3.309 .238 4.346 .564 7.655 4 40 4.766 .185 5.148 .615 9.914 10 "Forced Squish-Jet" Design NJ with 8 millimeter Spark Probe Depth Spark Timing 0-5% MFB 5-90% MFB MFB No. of Degrees /* cr Time Runs BTDC (ms) (ms) (ms) (ms) (ms) 20 30 40 5.337 .142 2.371 .129 7.708 10 "Forced Squish-Jet" Design NJ with 16 millimeter Spark Probe Depth Spark Timing 0-5% MFB 5-90% MFB MFB No. of Degrees Time Runs BTDC (ms) (ms) (ms) (ms) (ms) 20 30 40 5.560 .066 2.679 .177 8.239 8 Table 4.8: Combustion Durations with Piston NJ Figures 74 Figure 2.1: Schematic of Rapid Intake and Compression Machine Figures 75 Figure 2.2: Schematic of RICM Intake Valve and Linkage Mechanism Figures 76 "Forced Squish-Jet" Piston 2J with two 2 mm jet slots "Forced Squish-Jet" Piston 4J with four 2 mm jet slots "Forced Squish-Jet" Piston 2LJ with two 4 mm jet slots "Forced Squish-Jet" Piston NJ with No jet slots Figure 2.3: "Forced Squish-Jet" Piston Crowns for the RICM Figures 77 F O R C E D S Q U I S H - J E T HEAD A S S E M B L Y R A P I D C O M P R E S S I O N M A C H I N E M E C H A N I C A L E N G I N E E R I N G D E P A R T M E N T THE UNIVERSITY OF BRITISH COLUMBIA tCMC: 1:2 HAWN IY C . M A W L E 1 WUWING NUMBER DATE: 2 2 . 0 6 . 8 9 AfftOVED »Y R L E - 0 9 5 Figure 2.4: Cross-Section of the "Forced Squish-Jet" Head Assembly TIME (milliseconds) TIME (milliseconds) Figure 3.1: Ensamble-Averaged Mean Velocity and Turbulence Intensity for N 10 and 15 Cycles = 2, 5, Figures 79 Figure 3.2: Comparisons of Ignition Delay (FIP) and Combustion Duration Computed with Krieger-Borman and Rassweiler-Withrow Analysis on Identical Pressure Files (from reference [48]) C R A N K A N G L E (d«g) Figure 3.3: Comparisons of Estimates of Apparent MFB to Results from Comprehensive Computer Model (from reference [50]) Figures Figure 4.1: Mean Velocity and Turbulence Intensity measured at the centre and near edge of the combustion chamber lined up with a hole in the 20 mm Perforated Plate Figures 81 TIME (milliseconds) TIME (milliseconds) Figure 4.2: Mean Velocity and Turbulence Intensity measured at the centre and near the edge of the combustion chamber lined up with a hole in the 10 mm Perforated Plate Figures 82 - «t centre " near edge 10 2 0 30 4 0 5 0 TIME (milliseconds) 6 0 7 0 Figure 4.3: Mean Velocity and Turbulence Intensity measured at the centre and near the edge of the combustion chamber lined up with a hole in the 5 mm Perforated Plate Figures 83 TIME (milliseconds) TIME (milliseconds) Figure 4.4: Mean Velocity and Turbulence Intensity measured at the centre of a hole and between two holes with the wire parallel and perpendicular to the 20 mm Perforated Plate Figures 84 TIME (milliseconds) Figure 4.5: Mean Velocity and Turbulence Intensity measured at the centre of a hole and between two holes with the wire parallel and perpendicular to the 10 mm Perforated Plate Figures 85 TIME (milliseconds) Figure 4.6: Mean Velocity and Turbulence Intensity measured at the centre of a hole and between two holes with the wire parallel and perpendicular to the 5 mm Perforated Plate Figures 86 TIME (miUiiecond») Figure 4.7: Mean Velocity and Turbulence Intensity measured at the centre of a hole with the wire parallel to the 20, 10 and 5 mm Perforated Plates Figures 87 H U o a > z a 2 2 — 20 mm holes 10 mm holes 5 mm holes 3.5 4.5 LN(TIME) - 2 cn Z a H z w a D a a: D — 20 mm holes 10 mm holes 5 mm holes -8 3.5 4.5 LN(TIME) Figure 4.8: Log-Log Plot of Mean Velocity and Turbulence Intensity measured at the centre of a hole, with the wire parallel to the 20, 10 and 5 mm Perforated Plates Figures 88 a « z A D CD Z o H O < a: s o . e . 0.6 0.4 0.2 / / TDC ' • / // 7 • • 10 deg BTDC II: II: 20 deg BTDC 30 deg BTDC III II: II /•/ 40 deg BTDC / / 'J '•'/ 1 j 1 1 V n l :i in K.l fin fin fm /// 6 6 10 12 14 16 18 20 TIME FROM IGNITION (millinueconds) 22 24 Figure 4.9: MFB Curves with ignition timing at TDC as well as 10, 20, 30 and 40 degrees BTDC for the 20 mm Perforated Plate o a z CC E> Z o H O < h. V> cn < 2 0.8 . 0.6 0.4 0.2 '/•••' / TDC ' / • • / 10 deg BTDC ih: / 20 deg BTDC .'/// / 30 deg BTDC //// / 40 deg BTDC IH: / i it: / It! II! II! II! II 11 / I a / /// / lilf iiy 10 12 14 16 18 20 22 24 TIME FROM IGNITION (millimseconds) Figure 4.10: MFB Curves with ignition timing at TDC as well as 10, 20, 30 and 40 degrees BTDC for the 10 mm Perforated Plate Figures 89 0.8 J D a z a D CO z o r* o <: cc u. in cn <; S 0.4 0.2 ' 7 '/ / '/ / / / / TDC / 10 deg / 20 deg / 30 deg 40 deg BTDC BTDC BTDC BTDC //' /' / / //// / //// / //// / //// / !// 1 1 } "~\ 1 1 ' " T " i r 8 10 12 14 16 18 20 22 24 2 4 6 TIME FROM IGNITION (millimseconds) Figure 4.11: MFB Curves with ignition timing at TDC as well as 10, 20, 30 and 40 degrees BTDC for the 5 mm Perforated Plate 0.8 _| a a z CD 0.6 Z o H O < A 6-VI CO < 0.2 J 0.4 — 20 mm holes I •' ' 10 mm ho 1 es 5 mm holes 1 : / / ' / / '' / / ' / •'  / / •' / / •' ' / / / / ' / / 1 1 •'  1 1 •' / / / / - / / / / ' / / •' - y / / y' / ''' 1 1 1 — — — 1 1 1 J 1 — i i r 10 12 14 16 18 20 22 24 0 2 4 6 8 TIME FROM IGNITION (millimseconds) Figure 4.12: MFB Curves with ignition timing at TDC for the 20, 10 and 5 mm Perforated Plates Figures 90 Q a 2 A 9 CO 2 O H O < Efa CO CO •«! 2 0.8 0.6 0.4 0.2 J 20 mm holes 1 ' 1 10 mm holes I •' 1 5 mm holes 1 ;' i I •' ' / •'' ' / •'  ' / •' /' / ' / / •' '' / •' / / ' ' / '' / '' / / •'' / / •' / •'' / /•' / ' / ' / /' / -/ / / / 20 22 24 0 2 4 6 8 10 12 14 16 18 TIME FROM IGNITION (millimseconds) Figure 4.13: MFB Curves with ignition timing at 10 degrees BTDC for the 20, 10 and 5 mm Perforated Plates D a 2 A 9 a 2 O l -M H O < A a CO cn < 0.8 J 0.6 0.4 0.2 J y 20 mm holes 10 mm holes 1 ' ^ 5 mm holes 1 • i 1 •' ' 1 ' / •' ' / / / / / / / ' /' / •' / / ; ' / •' ' / •'" / / •'  ' / ! I '• ' I / / •''/ / • / J •/ / / ' i i i i i i i 20 22 24 0 2 4 6 8 10 12 14 16 18 TIME FROM IGNITION (millimseconds) Figure 4.14: MFB Curves with ignition timing at 20 degrees BTDC for the 20, 10 and 5 mm Perforated Plates Figures 91 Q a z es m z o H O < OS fa. cn cn < s 0.8 J 0.6 0.4 J 0.2 J 20 mm holes / •'' 10 mm holes / •' ' 5 mm holes / •' ' / •' / / •'  ' / •' ' / '•' ' / •' ' / •' ' / •' /' / ' ' / .'' / . ; / / / It //' 1 1 1 1 — • — 1 1 8 10 12 14 16 18 TIME FROM IGNITION (millimseconds) 20 22 24 Figure 4.15: MFB Curves with ignition timing at 30 degrees BTDC for the 20, 10 and 5 mm Perforated Plates D a z cc m z o H O < St b. cn cn •< £ 0.8 J 0.6 0.4 J 0.2 J 1 20 mm holes 1 ;' I 10 mm holes • // i i i i i i 5 mm holes / •' ' / ' / ' / •' ' / •' / / •' / / .' ' / /'/ / •' 1 1 / //: 1 1 1 — — r — t — — i 22 24 0 2 4 6 8 10 12 14 16 18 20 TIME FROM IGNITION (millimseconds) Figure 4.16: MFB Curves with ignition timing at 40 degrees BTDC for the 20, 10 and 5 mm Perforated Plates Figures 92 10 2 mm depth 6 mm depth 16 mm depth 8 . 1 1 1 1 1 — ' " " i " " | | T 180 200 220 240 260 280 300 320 340 360 CRANK ANGLE (degree.) Figure 4.17: Mean Velocity and Turbulence Intensity measured at the 2, 8 and 16 mm depths near the bowl edge for Piston BIP with respect to crank angle 10 2 mm depth 6 mm depth 16 «m depth T 8 H cn 2 a 63 2 a D CD S I 10 I 60 70 TIME (milliseconds) Figure 4.18: Mean Velocity and Turbulence Intensity measured at the 2, 8 and 16 mm depths near the bowl edge for Piston BIP with respect to time Figures 94 20 ~ 15 2 mm depth 8 mm depth 16 mm depth 10 . 1 1 r 180 200 220 240 260 280 300 CRANK ANGLE (degrees) 320 340 360 2 mm depth 8 mm depth 16 mm depth 180 200 220 240 260 280 300 320 340 360 CRANK ANGLE (degrees) Figure 4.19: Mean Velocity and Turbulence Intensity measured at the 2, 8 and 16 mm depths at the centre of the bowl for Piston BIP with respect to crank angle Figures 95 20 2 mm depth 8 mm depth 16 mm depth „ 15 10 . 60 70 TIME (milliseconds) 10 2 mm depth 8 mm depth 16 mm depth 4 . f 1 ^ 10 20 30 40 50 TIME (milliseconds) 60 70 Figure 4.20: Mean Velocity and Turbulence Intensity measured at the 2, 8 and 16 mm depths at the centre of the bowl for Piston BIP with respect to time Figures 96 10° 6 T D C T 20* BTDC f i 30" B T D C O SPARK LOCATION + HWA L O C A T / O N F L O W M O T I O N S K E T C H FOR B O W L - I N - P I S T O N D E 5 / G N RAPID C O M P R E S S I O N / M A C H I N E MECHANICAL ENGINEERING DEPARTMENT THE UNIVERSITY OF BRITISH COLUMBIA t o u t F U L L 5 I Z E wuwn iy C. M A W L E DATfe 2 7 . 0 9 . 8 9 A m r o v n ) IT M A W I N O NOMBtft Figure 4.21: Sketch of Flow Pattern Generated with the Bowl-in-Piston Design Figures 97 Figure 4.22: Mean Velocity and Turbulence Intensity measured at the 2, 8 and 16 mm depths near the bowl edge lined up with a jet slot for Piston 2J Figures 98 10 — e 2 mm depth 6 mm depth 16 mm depth >• H cn x Z 6 a H Z H s -<a • J m g 2 T 180 200 220 240 260 280 300 CRANK ANGLE (degrees) 320 340 360 Figure 4.23: Mean Velocity and Turbulence Intensity measured at the 2, 8 and 16 mm depths near the bowl edge between the jet slots for Piston 2J Figures 99 ID BTDC 20' BTDC H 30° BTDC O SPhRK L O C A T I O N + HWA L O C A T I O N FLOW M O T I O N S K E T C H FOR F O R C E D S Q U I S H - J E T D E S I G N RAPID C O M P R E S S I O N M A C H I N E MECHANICAL ENGINEERING DEPARTMENT THE UNIVERSITY OF BRITISH COLUMBIA %CA±l, FULL SIZE H U W N »Y C M A W L E MAWINS NUMBER DATt 2 7 . 0 9 . 8 9 APftOVlD IT Figure 4.24: Sketch of Flow Pattern Generated with the "Forced Squish-Jet" Designs Figure 4.25: Mean Velocity and Turbulence Intensity measured at the 2, 8 and 16 mm depths at the centre of the bowl for Piston 2J Figures 101 20 _ 15 W If >• H *-i U O H > 10 z S P i s t o n 2J P i s t o n 2LJ P i s t o n 4J •7 / / S •i •< :'i,v.! V. V. 'i 5 5 v; .-\ f ' V". - " T 180 200 220 240 260 260 300 CRANK ANGLE (degrees) 320 340 360 10 6 . cn Z H H Z D CO cc H 2 P i s t o n 23 P i s t o n 2LJ P i s t o n 4J / t I > I l i.-- i iV '•  j ;.; /.i •• > I' /: / • i i " ' '. ' i 1 t i • 11 •'I', /; n; i 1 V I ft 11 1' 1 '/ V /' Wl */ *'| ' 1 180 200 220 240 260 280 300 320 340 360 CRANK ANGLE (degrees) Figure 4.26: Mean Velocity and Turbulence Intensity measured at the 8 mm depth near the bowl edge lined up with a jet slot for Pistons 2J, 2LJ and 4J Figures 1 0 2 10 e . >-H vi Z H H Z a o z a J D m cc D 2 . P i s t o n BIP P i s t o n 2J P i s t o n 4J P i s t o n NJ / / 1. <:l \ I U \ \ i \. I \\ i V- 1 •i :i /O y 1B0 200 220 240 260 260 300 C R A N K A N G L E (degree.) i r 320 340 360 Figure 4.27: Mean Velocity and Turbulence Intensity measured at the 2 mm depth near the bowl edge away from jet slots for Pistons BIP, 2J, 4J and NJ Figures 103 20 T 15 Pis t o n BIP Pi s t o n 2J Pi s t o n 2LJ Pi s t o n 4J Pi s t o n NJ 10 . 5 . 1 I I I 180 200 220 240 260 280 CRANK ANGLE (degree.) 360 P i s t o n BIP P i s t o n 2J Pi s t o n 2LJ P i s t o n 40 180 200 220 240 260 260 300 320 340 360 CRANK ANGLE (degree.) Figure 4.28: Mean Velocity and Turbulence Intensity measured at the 8 mm depth near the bowl edge away from jet slots for Pistons BIP, 2J, 2LJ, 4J and NJ Figures 104 20 15 P i s t o n BIP Pis t o n 23 Pist o n 2LJ Pis t o n 40 Pist o n NJ H U O a > z < a 2 10 . 1B0 200 220 240 260 280 300 CRANK ANGLE (degrees) 320 340 360 10 P i s t o n BIP P i s t o n 2J Pi s t o n 2LJ Pi s t o n 4J Pi s t o n NJ 4 . 2 . 1 1 1 1 180 200 220 240 260 280 300 CRANK ANGLE (degree.) 320 340 360 Figure 4.29: Mean Velocity and Turbulence Intensity measured at the 2 mm depth at the centre of the bowl for Pistons BIP, 2J, 2LJ, 4J and NJ 10 _ 6 . P i s t o n BIP P i s t o n 2J P i s t o n 2L J P i s t o n 4J P i s t o n NJ cn 2 a H 2 a u 2 a D CP OS D 6 . T i e o 200 220 240 260 280 300 CRANK ANGLE (degrees) 320 340 360 Figure 4.30: Mean Velocity and Turbulence Intensity measured at the 8 mm depth at the centre of the bowl for Pistons BIP, 2J, 2LJ, 4J and NJ Figures 106 10 P i s t o n BIP Pi s t o n 2J Pi s t o n 2LJ P i s t o n 4J P i s t o n NJ > H i—i tn Z ta H z I—( a o z H to c£ D H 4 . 160 —I— 200 220 240 260 280 300 CRANK ANGLE (degrees) 320 340 360 Figure 4.31: Mean Velocity and Turbulence Intensity measured at the 16 mm depth at the centre of the bowl for Pistons BIP, 2J, 2LJ, 4J and NJ Figures 107 Q a z A D n z o H O •< cs a < S 0.8 0.6 0.4 J 0.2 J 2 mm depth 8 mm depth 16 mm depth - J - T -6 I 10 I 12 6   14 TIME FROM IGNITION (millim» eco nds) 16 18 20 Figure 4.32: MFB Curves with ignition timing at 20 degrees BTDC for a spark probe depth of 2, 8 and 16 mm for Piston BIP Q a z A D 00 z o *—I H O < A a 1/3 cn < 2 o . e 0.6 J 0.4 0.2 2 mm depth 8 mm depth 10 12 16 18 20 0 2 ^ 6 8 10 12 14 TIME FROM IGNITION (millimseconds) Figure 4.33: MFB Curves with ignition timing at 30 degrees BTDC for a spark probe depth of 2, 8 and 16 mm for Piston BIP Figures 1 108 0.8 J Q a 2 « 0.6 2 O H U •< OS 0.4 b-CO co < 0.2 2 mm depth 8 mm depth 16 mm depth 10 12 16 6 8   14 TIME FROM IGNITION (millimseconds) 18 20 Figure 4.34: MFB Curves with ignition timing at 40 degrees BTDC for a spark probe depth of 2, 8 and 16 mm for Piston BIP 0.8 a 2 BS D „ ffl 0.6 2 O H < a. CO CO <: 2 0.4 0.2 2 mm depth 8 mm depth 16 mm depth 10 12 14 I 16 18 20 TIME FROM IGNITION (millimseconds) Figure 4.35: MFB Curves with ignition timing at 20 degrees BTDC for a spark probe depth of 2, 8 and 16 mm for Piston 2J Figures 109 D Z A D CD Z o H O < a: b. cn cn <: 2 0.8 J 0.6 0.4 J 0.2 2 mm depth 8 mm depth 16 mm depth ~1— 12 14 6 8 10 TIME FROM IGNITION (miUimseconds) " I -16 18 20 Figure 4.36: MFB Curves with ignition timing at 30 degrees BTDC for a spark probe depth of 2, 8 and 16 mm for Piston 2J Q ta z A S CO z o H < cn cn <: 0.8 J 0.6 J 0.4 0.2 J 2 mm depth 8 mm depth 16 mm depth 10 12 T" 14 16 —r~ 18 20 TIME FROM IGNITION (miIlim»econd») Figure 4.37: MFB Curves with ignition timing at 40 degrees BTDC for a spark probe depth of 2, 8 and 16 mm for Piston 2J Figures 110 1 1 1 1 1 1 1 1 1 1 1 0 2 4 6 8 10 12 14 16 18 20 TIME FROM IGNITION (millimseconds) Figure 4.38: MFB Curves with ignition timing at 40 degrees BTDC for a spark probe depth of 2 mm for Pistons BIP, 2J, 2LJ, 4J and NJ P i s t o n BIP Pis t o n 2J Pis t o n 2LJ Pi s t o n 4J Pis t o n NJ 0 2 4 6 8 10 12 14 16 18 20 TIME FROM IGNITION (millimseconds) Figure 4.39: MFB Curves with ignition timing at 40 degrees BTDC for a spark probe depth of 8 mm for Pistons BIP, 2J, 2LJ, 4J and NJ Figures 111 P i s t o n BIP Pis t o n 2J P i s t o n 2LJ P i s t o n 4J P i s t o n NJ 1 1 1 1 1 1 1 1 I 1 1 0 2 4 6 6 10 12 14 16 18 20 TIME FROM IGNITION (millimseconds) Figure 4.40: MFB Curves with ignition timing at 40 degrees BTDC for a spark probe depth of 16 mm for Pistons BIP, 2J, 2LJ, 4J and NJ D a z EC D m z o »—(-H V < os u. < 0.8 J 0.6 0.4 J 0.2 P i s t o n BIP Pis t o n 2J Pis t o n 2LJ Pi s t o n NJ ~ i 1 1 1 1 1 1 r -2 4 6 8 10 12 14 16 TIME FROM IGNITION (millimseconds) 18 20 Figure 4.41: MFB Curves with ignition timing at 30 degrees BTDC for a spark probe depth of 2 mm for Pistons BIP, 2J, 2LJ and NJ Figures 112 a a z D m z o O a: tL. cn cn <: o . e 0.6 . 0.4 0.2 . Pi s t o n BIP P i s t o n 2J Pis t o n 2LJ "T 1 1 1 1 1 1 1 1 2 4 6 8 10 12 14 16 16 20 TIME FROM IGNITION (millimseconds) Figure 4.42: MFB Curves with ignition timing at 30 degrees BTDC for a spark probe depth of 8 mm for Pistons BIP, 2J and 2LJ P i s t o n BIP Pi s t o n 2J P i s t o n 2L0 0 2 4 6 6 10 12 14 16 18 20 TIME FROM IGNITION (miUimseconds) Figure 4.43: depth of 16 MFB Curves with ignition timing at 30 degrees BTDC for a spark probe mm for Pistons BIP, 2J and 2LJ Figures 113 1 1 1 1 1 1 1 1 1 1 1 1 0 2 4 6 6 10 12 14 16 18 20 TIME FROM IGNITION (millimseconds) Figure 4.44: MFB Curves with ignition timing at 20 degrees BTDC for a spark probe depth of 2 mm for Pistons BIP and 23 l 1 1 1 1 1 1 1 1 1 1 I 0 2 4 6 8 10 12 14 16 18 20 TIME FROM IGNITION (millimseconds) Figure 4.45: MFB Curves with ignition timing at 20 degrees BTDC for a spark probe depth of 8 mm for Pistons BIP and 23 Figures 114 Figure 4.46: MFB Curves with ignition timing at 20 degrees BTDC for a spark probe depth of 16 mm for Pistons BIP and 2J Appendix A Description of M F B Computer Program The listing of the MFB computer program based on the method developed by Rassweiler and Withrow [46] is given at the end of this appendix. The number of data points read into the program were first reduced by a factor of ten to decrease the calculation and file transfer time required. This increased the time between data points to 0.16 milliseconds from 0.016. The effect of the reduced step size can be seen in figure A . l where the MFB curves produced with the two step sizes are practically indistinguishable. The point at the start of the combustion period was found from the crank angle data and the end of combustion was selected to be the point at which there was no pressure rise due to combustion. Since the end of combustion was well after the point where the piston had stopped at TDC, the point of no pressure rise due to combustion was also the point of peak pressure. During the early stages of combustion the pressure rise due to the piston motion (change in volume) was calculated based on a polytropic compression of the fuel/air mixture. The volume change was determined from the piston position calculated from the crank angle data: Vol{9) = [2RC + Lev - X{6)]A X{9) = Rc cos 9 + y/L2CR - (Rc sin 8f - {LCR - Rc) where X(9) is the distance from BDC to the top of the piston, Rc is the crank radius, LCR is the connecting rod length, Lev is the height of the clearance volume with the disc piston and A is the cross-sectional area of the cylinder. 115 Appendix A. Description of MFB Computer Program 116 The change in pressure due to the movement of the piston was determined from: = P(i){Y±zllr _ P { l _ i) V{i) where P(i — 1) is the pressure before the volume change , V(i — 1) and V(i) are the two volumes and 7 is the polytropic coefficient of compression. The polytropic coefficient of compression was selected to be 1.35 based on calculations from the experimental data. The effects of changing 7 from 1.25 to 1.40 are shown in figure A.2 where the changes can be seen to have only a small effect on the overall results of the analysis. During each interval the pressure rise due to piston motion was subtracted from the total pressure rise leaving the pressure rise due to combustion. The pressure rise was assumed to take place at a constant volume (since the volume change was small) at the end of the interval (V(i)). Since the pressure rise produced at a constant volume from the combustion of a given mass of fuel/air mixture is inversely proportional to the volume [46], the pressure rise due to combustion in each interval was scaled to an equivalent pressure rise due to the combustion of the same mass of fuel/air mixture at one consistent volume. The volume chosen was the volume at TDC since most of the combustion takes place after the piston had stopped at TDC. Thus, the equivalent pressure rise due to the combustion is: APvTDC(i) = A P v ( 0 ^ VTDC where V(i) is the volume in interval i and and VTDC is the volume at TDC. The MFB is the sum of the pressure rise due to combustion up to the end of each interval divided by the total pressure rise due to combustion. U E f = 1 A P V x c c U ) where i is the interval and N is the total number of intervals. Appendix A. Description of MFB Computer Program 117 2 3 4 5 6 7 8 9 TIME FROM IGNITION (milbmseconds) Figure A. l : Comparison between the MFB curves calculated with a time step of 0.016 and 0.16 milliseconds 0 1 2 3 4 5 6 7 8 9 TIME FROM IGNITION (millimseconds) Figure A.2: Comparison between the MFB curves calculated using a polytropic coefficient of compression of 1.25, 1.30, 1.35 and 1.40 Appendix A. Description of MFB Computer Program 118 VAX FORTRAN V4. 4-177 Page 1 DUBO:(MAWLE.PROGRAMS)MB_RANDW.THESIS,-2 0001 C - - - - - - - - - - - - - - - - - - - - . — — . - . — . - . - — . . — . . . . . . . . . . . . . . . . . „ . . „ _ 0002 C - MB •** MASS BURN PROGRAM *** 13.12.1985 0003 C 0004 C 0005 C THIS PROGRAM USES THE METHOD DELELOPED BY RASSWEILER fc WITHROW 0006 C BUT TAKES CONSTANT TIME INCREMENTS INSTEAD OF CONSTANT DCA 0007 C INCREMENT. THE PISTON STOPS AT TDC SO THERE IS NO EXPANSION 0008 C STROKE. 0009 C 0010 IMPLICIT REAL*8(A-H,0-Z) 0011 0012 REAL TIME(5,1000), PRES(5,1000),DEGREE)5,1000),MFB(1000) 0013 REAL ZUENDUNG(5,1).DP(1000),DPP(1000),DPCV(1000),VOL(1000) 0014 0015 CHARACTER*40 INFILE)5),OUTFILE,EZFILE.TELLEFILE 0016 0017 C 0018 IL-1 0019 C READ IN INPUT AND OUTPUT FILE NAMES 0020 990 CONTINUE 0021 WRITE(6,1020) I L 0022 1020 FORMAT)/,' ENTER INPUT FILE NAME (eg MBIN*.DAT) No:',12) 0023 READ(5,1040) I N F I L E ( I L ) 0024 1040 FORMAT)A4 0) 0025 0026 WRITE(6,1060) 0027 1060 FORMAT)/,' ENTER OUTPUT FILE NAME (eg MBOUT*.DAT):',/,$) 0028 READ(5,1040) OUTFILE 0029 0030 0031 1100 FORMAT(12) 0032 1140 WRITE(6,1160) 0033 1160 FORMAT(/,' ENTER IF TELLEGRAPH ON (1) OR OFF (0):*,/,$) 0034 READ(5,1100) NTELLEGRAPH 0035 0036 I F (NTELLEGRAPH -EQ. 0) GOTO 1200 0037 WRITE(6,11B0) 0038 1180 FORMATf/,' ENTER TELLEGRAPH OUTPUT FILE NAME (eg TELLMB*.DAT):'S) 0039 READ(5,1040) TELLEFILE 0040 1200 CONTINUE 0041 C FINISHED READING IN FILE NAMES 0042 0043 IF ( I L .GT. 1) GOTO 1290 0044 0045 C READ IN SPARK ADVANCE 0046 WRITE(6,1220) 0047 1220 FORMATf/,' ENTER SPARK ADVANCE IN DCA BTDC:',/,$) 0048 READ(5,1240) SPKAD 0049 1240 FORMAT(F6.2) 0050 C 0051 0052 1290 CONTINUE 0053 C READ IN INPUT FILE DATA 0054 OPEN(UNIT-10+IL,FILE-INFILE(IL),STATUS-'OLD',FORM-'UNFORMATTED') 0055 READ(10+IL) NNUMBER 0056 READ)104IL) ZUENDUNG(IL,1) 0057 DO 1300 1-1,NNUMBER endix A. Description of MFB Computer Program 1 KB RANDW$HAIN VAX FORTRAN V4.4-177 Page 2 DUBO:t KAHLE.PROGRAMS]KB_RANDW.TBE5IS;2 005B READ!10+IL) T I M E ( I L , I ) , PRES(IL,I),DEGREE(IL,I) 0059 1300 CONTINUE 0060 WRITE!6,1320) NNUMBER 0061 1320 FORMAT(/, ' FINISHED READING IN ',15,' DATA ARRAYS') 0062 C ...TIME IN MILLISECONDS 0063 C . . .PRES IN KPA 0064 C ...DEGREE IN DCA 0065 CLOSE(UNIT-10+IL) 0066 1322 CONTINUE 0067 C FINISHED EEADING IN INPUT DATA 0068 0069 WRITE(6,1323) 0070 1323 FORMAT(/,' INPUT TBE NUMBER Or DATA POINTS TO BE CHECKED:') 0071 READ{5,1324) NKUM 0072 1324 FORMAT!13) 0073 0074 C OPEN OUTPUT F I L E 0075 C 0076 OPEN(UNIT-2,FILE-OUTFILE,STATUS-'NEW',FORM-'FORMATTED') 0077 WRITE(2,1325) INFILE(IL),SPKAD,ZUENDUNG(IL,1) 0078 1325 FORMAT(/,' INPUT FILENAME:»,A12,' SPKAD:*,F6.2,' ZUENDUNG:',F6.2) 0079 KRITE(2,1327) 0080 1327 FORMAT!/,' TIME DCA PRES BASS F.B. DP DPP DPCV') 0081 0082 I F (NTELLEGRAPH .EQ. 0) GOTO 1331 0083 OPEN ( UNIT-8 , r" ILE-TELLE FILE, STATUS- 'NEW' , FORM- * FORMATTED' ) 0084 WRITE(8,1330) ZUENDUNG(IL, 1) 0085 1330 FORMAT(Fl0.4) 0086 1331 CONTINUE 0087 0088 C 0089 C FIND THE POINT AT SPARK AND AT THE END OF COMBUSTION 0090 0091 DCASPARK - 360.0 - SPKAD 0092 DO 1400 I-l.NNUM 0093 IF (DEGREE!IL,I) .GE. DCASPARK) GOTO 1410 0094 ISTART - I 0095 1400 CONTINUE 0096 1410 PRESMAX - PRES(IL,ISTART) 0097 DO 1450 I-ISTART.NNUM 0098 I F (PRESMAX .GT. P R E S ( I L . I ) ) GOTO 1450 0099 IEND - I 0100 PRESMAX - PRES(IL.I) 0101 1450 CONTINUE 0102 0103 WRITE(6,1459) DCASPARK,ISTART,IEND,PRESMAX 0104 1459 FORMAT!' DCASPARK:',F6.1,• I-',14,' to',14,' PRESMAX:',F8 .1) 0105 0106 C CALCULATE VOLUME AT SPARK AND END OF COMBUSTION 0107 0108 THETA - DEGREE!IL,ISTART)/57.2958D0 0109 X - 50.«DCOS(THETA)+DSQRT(30625.-!50.«DSIN(THETA))**2)-125.0D0 0110 VOL(ISTART) - (114.287D0 - X) / 123.3453D0 0111 0112 THETA - 360.0D0/57.2958D0 0113 X - 50.*DCOS(THETA)+DSQRT(30625.-(60.«DSIN(THETA))*«2)-125.0D0 0114 VOLTDC - (114.2B7D0 - X) / 123.3453D0 Appendix A. Description of MFB Computer Program 120 KB RANDWSMAIN VAX FORTRAN V4.4-177 Page 3 DUBO:(HAWLE.PROGRAMS]KB RANDW.THESIS;2 0115 0116 0117 C WRITE(6,1660) 0118 C 1660 FORMAT!/,'ENTER THE VALUE OF GAMMA (eg 1.25 t o 1.40):') 0119 C READ(5,1670) GAMMA 0120 C 1670 FORMAT!F4.2) 0121 0122 C CALCULATE THE HFB FROM DEGREE AND PRESSURE DATA 0123 0124 GAMMA - 1.35 0125 0126 PSUM - O.OD0 0127 DO 2000 I-ISTART+l.IEND 0128 THETA - DEGREE(IL.I)/57.2958D0 0129 X - 5D.«DCOS(THETA)+DSORT(30625.-(50.*DSIN(THETA))*«2)-125.0D0 0130 VOL(I) - (114.287D0 - X) / 123.3453D0 0131 DP(I) - PRES(IL,I) - P R E S ( l L . l - l ) 0132 DPC - DP(I)-(PRES(IL,I-l)*(VOL(I-l)/VOL(I))*»GAHHA-PRES(IL,I-l)) 0133 DPP(I) - DPCI) - DPC 0134 DPCV(I) - DPC * VOUD/VOLTDC 0135 C I F ( DPCV(I) .GT. 0.0 ) GOTO 1999 0136 C DPCV(I) - 0.0 0137 C 1999 CONTINUE 0138 PSUM - PSUM • DPCV(I) 0139 MFB(I) - PSUM 0140 2000 CONTINUE 0141 0142 C PRINT THE RESULTS INTO THE OUTPUT FILE 0143 0144 DO 3000 I-ISTART*1,1END 0145 HFB(I) - HFB(I)/MFB(IEND) 0146 WRITE(2,3100) TIME!IL,I),DEGREE!IL,I),PRES(IL,I),HFB(I), 0147 • DP(I),DPP(I),DPCV(I) 0148 I F (NTELLEGRAPH .EQ. 0) GOTO 3000 0149 KRITE(8,3200) TIME(IL,I).DEGREE!IL,I),PRES(IL,I).HFB(I) 0150 3000 CONTINUE 0151 0152 I F (NTELLEGRAPH .EQ. 0) GOTO 3001 0153 ENDOFFILE - 5000.0 0154 KRITE(8,3200) ENDOFFILE,DEGREE(IL,I),PRES(IL,I),MFB(I) 0155 0156 3001 CONTINUE 0157 3100 FORMAT(F7.3,FB.2,F9.2,4F9.5) 0158 3200 FORMAT!4F10.4) 0159 0160 STOP 0161 END Appendix B Calculations of Squish Velocity The combustion chamber volume with the bowl-in-piston design was separated into two parts for this simple analysis. The volume above the squish area was designated Vi and the volume in and above the bowl was designated V2. The squish was assumed to have a uniform velocity profile through the area between the two volumes. The total mass of fluid in the combustion chamber at any one time assuming no leakage is: mtolai = pVtotai (B.l) From continuity of mass for the closed system dmtotai dp dVt so that dp _ P dVtotai dt' vtolal dt  {t5S) Taking the control volume above the squish area (volume Vi) the continuity equation gives: dmi dp,, d\\ /T% . -ar = & v ' +"-flr ( B 4 ) substituting equation B.3 into equation B.4 gives: dmx p dVtotai,r dV\ -dr = - \ ^ : l - d r l i + p ~ d 7 ( B 5 ) Assuming uniform density and uniform velocity normal to the control surface the mass 121 Appendix B. Calculations of Squish Velocity 122 flow out of volume Vi is: drrii —Q^ = pv,qui.hA (B.6) where v,quith is the squish velocity and A is the area between the control surfaces which the fluid moves through. Equating equations B.5 and B.6 and rearranging gives: _ 1 Vi dVtotal dvr The volumes in each region are as follows: Vi = A [ 5 + fcc] (B.8) Vtotai = A2[S + hc + Dh] + A![S + hc] (B.9) where S is the distance from TDC to the top of the piston, he is the clearance height, A-i is the squish area, Db is the bowl depth and A2 is the cross-sectional area of the bowl. Differentiating equations B.8 and B.9 gives: dVtotai dS ,„,.. - Q T ^ + M-QI (B.11) Substituting equations B.8, B.9, B.10 and B . l l back into equation B.7 gives: 1 Atf + hcXAi + A,) dS dS Rearranging equation B.12 gives: dS 1 ALA2Db  dt 'A^MS + hc) + A2(S + hc + Dby Substituting the areas At = TTR2 - irr2 (B.13) Appendix B. Calculations of Squish Velocity 123 A + 2wrS where R is the cylinder radius and r is the bowl radius into equation B.13 and rearranging gives: 8SrDb, R2-r2 f . , m . f c - d t s lR2(S + h c ) + r2Db\ WAV Now all that is required is the piston velocity | | which is found by differentiating the equation of the piston position S = 2RC + hc - X(8) X{8) = Rc cos6+ sjL2CR - {Re sin Of - (LCR - Rc) where X(6) is the distance from BDC to the top of the piston, Rc is the crank radius, LCR is the connecting rod length and 9 is the crank angle. dS dX 86 dt 89 dt (B.15) where |* is the revolution rate of the crank shaft. For this work it was taken to be 1000 rpms. Table B.l lists the piston position, piston velocity and squish velocity for the conditions of the experiments with the bowl-in-piston design. The effects of changing the clearance height {he) from 1 to 5 millimeters is shown in figure B.l which indicates the importance of the clearance height in developing squish. Appendix B. Calculations of Squish Velocity 124 Crank angle Distance from Piston velocity Squish velocity head to piston degrees mm m/s m/s 320 15.67 1.58 3.42 322 14.33 1.52 3.74 324 13.03 1.44 4.11 326 11.80 1.37 4.51 328 10.62 1.30 4.97 330 9.49 1.22 5:48 332 8.43 1.15 6.06 334 7.44 1.07 6.70 336 6.51 0.99 7.41 338 5.65 0.91 8.20 340 4.85 0.83 9.06 342 4.13 0.75 9.99 344 3.48 0.67 10.95 346 2.90 0.59 11.90 348 2.40 0.50 12.73 350 1.98 0.42 13.26 352 1.63 0.34 13.20 354 1.35 0.25 12.12 356 1.16 0.17 9.57 358 1.04 0.08 5.37 360 1.00 0.00 0.00 Table B.l: Piston Position, Piston Velocity and Squish Velocity for the standard bowl-in-piston design with 1 mm clearance gap Appendix B. Calculations of Squish Velocity 125 Figure B.l: Squish Velocity and Piston Velocity based on a simple model for clearance heights of 1, 2 and 5 mm for the standard bowl-in-piston design at 1000 rpm 

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