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Performance of a partially stratified-charge natural gas engine Reynolds, Conor 2002

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PERFORMANCE OF A PARTIALLY STRATIFIED-CHARGE NATURAL GAS ENGINE By Conor Reynolds B.A., B.A.I., University of Dublin, Trinity College, Ireland, 1998  A THESIS SUBMITTED IN PARTIAL FULFILLMENT OF THE REQUIREMENTS FOR THE DEGREE OF MASTER OF APPLIED SCIENCE in THE FACULTY OF GRADUATE STUDIES DEPARTMENT OF MECHANICAL ENGINEERING  We accept this thesis as conforming to the required standard  THE UNIVERSITY OF BRITISH COLUMBIA November 2001 © Conor Reynolds, 2001  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  hU^HfirUtcfa- &U(r(\iei~J^lNQ~  The University of British Columbia Vancouver, Canada Date  DE-6 (2/88)  ABSTRACT  The Partially Stratified-Charge (PSC) concept is a novel way to improve the combustion characteristics of an otherwise standard spark-ignition engine, operating at a lean air-fuel ratio.  PSC involves providing a relatively rich air-fuel mixture in the vicinity of the  spark plug, while maintaining an ultra-lean homogeneous charge in the main combustion chamber area. This is achieved by injecting a small quantity of natural gas, comprising less than 5 % of the overall fuel mass, in the region of the spark plug just prior to ignition. A stable flame kernel is formed in this region, which then rapidly propagates through the combustion chamber. PSC requires no modification to basic engine design, but only to the method of introducing fuel to the combustion chamber. Natural gas has been used, both for injection through the spark-plug and for the main homogeneous fuelling, because PSC is suited to gaseous fuelling and also because natural gas is an abundant, clean burning and relatively low cost fuel.  Tests have been performed using a fully instrumented single cylinder research engine. The experiments were undertaken at selected engine speeds and MBT (minimum advance for best torque) spark timing and wide-open throttle for both homogenous and PSC engine configurations. Engine load was varied without use of a throttle by leaning the air-fuel mixture; the minimum achievable load therefore occurred at the lean misfire limit. The effect of changing PSC injection timing and quantity has been investigated.  The performance characteristics of PSC have been compared to the general homogeneous fuelling case.  The results show that the improved engine performance with PSC,  particularly during ultra-lean operation, reduced levels of both brake specific fuel consumption and exhaust emissions by approximately 8% and there was an increase in brake mean effective pressure of approximately 7%. An analysis of in-cylinder pressure data revealed that PSC reduces ignition delay and combustion duration, and increases peak cylinder pressure.  An extension of the lean misfire limit was achieved due to  improved combustion initiation and stability with PSC.  ii  TABLE OF CONTENTS  Abstract  ii  Table of Contents  iii  List of Tables  v  List of Figures  vi  Nomenclature  ix  Acknowledgements  xii  Chapter 1  INTRODUCTION  1  1.1  Overview and Objectives  1  1.2  Literature Review  3  1.2.1  Natural Gas Composition and Properties  3  1.2.2  Natural Gas Fuelling Strategies  6  Chapter 2  EXPERIMENTAL FACILITY AND DATA ANALYSIS  10  2.1  Introduction  10  2.2  Research Engine  10  2.3  Partially Stratified-Charge System  12  2.4  Instrumentation  16  2.5  Data Acquisition, User Interface and Data Analysis  2.6  Uncertainties and Repeatability  18 20  iii  Chapter 3  RESULTS AND DISCUSSION  24  3.1  Introduction  24  3.2  An Optimisation Technique based on BSFC  24  3.3  Spark Timing Results  25  3.4  Performance Results - Efficiency and Power  27  3.5  Extension of the Lean Misfire Limit  29  3.6  Emissions Results  30  3.6.1  Carbon Monoxide  30  3.6.2  Total Hydrocarbons and Methane  31  3.6.3  Oxides of Nitrogen  31  3.7  Combustion Analysis  32  3.8  Summary  34  CONCLUSIONS  56  4.1  Conclusions  56  4.2  Recommendations for Future Work  57  Chapter 4  REFERENCES  59  Appendix A  Engine Operating Procedure  62  Appendix B  Emissions Bench: Operating and Calibration Procedure  65  Appendix C  Data Acquisition System Operating Procedure  69  Appendix D  Data Processing Instructions  73  Appendix E  List of Acquired Data and Calculated Parameters  75  Appendix F  Engine Dimensions/Geometry  81  Appendix G  Approximate Stoichiometric Air-fuel Ratio for B.C. Natural Gas  87  iv  LIST OF TABLES  Table 1.1  Average Composition of Natural Gas (Tilbury Gate Station, B.C.)  Table 1.2  Comparison of Properties of B.C. Natural Gas, Pure Methane  3  and Gasoline  5  Table 2.1  UBC Ricardo Hydra Engine Specifications  12  Table 2.2  Calculated PSC Control Settings (Based on Solenoid Delay), 2500rpm  15  Table 2.3  Ricardo Hydra Engine - Main Instrumentation Specifications  17  Table 2.4  Emissions Bench - Analyser Specifications  18  Table 2.5  Estimated Uncertainties for Performance Parameters  21  Table 3.1  Comparison of MBT Spark Timings in Degrees Crank Angle (BTDC) for Homogeneous and PSC Fuelling CO, tHC and NO (Grams of Emission per Kilogram of CNG), ^=1.65 Heating Values of the Components of B.C. Natural Gas  Table 3.2 Table G.l  26  x  32 87  v  LIST OF FIGURES  Figure 2.1  Schematic of Ricardo Hydra Test Engine with PSC Fuelling System  11  Figure 2.2  PSC Spark-plug Injector  13  Figure 2.3  Experimental Apparatus to Determine Injection Delay Characteristics  14  Figure 2.4  Empirical Data showing Injector Delay vs. Signal Duration  15  Figure 2.5  Repeatability Test: BSFC vs. Relative Air-Fuel Ratio  22  Figure 2.6  Repeatability Test: BMEP vs. Relative Air-Fuel Ratio  22  Figure 3.1  BSFC with Changing PSC Flow-Rate (EOI at MBT-10), 2500rpm BSFC with Changing PSC Timing, 2500rpm  35  (Injection flow-rate = 15 g/h)  35  Figure 3.3  Spark timing (MBT) Retard due to PSC, 2500rpm  36  Figure 3.4  Thermal Efficiency with Changing PSC Flow-Rate, 2500rpm  36  Figure 3.5  Thermal Efficiency with Changing PSC Timing, 2500rpm  37  Figure 3.6  Power Control Using Lean Combustion, 2500rpm  37  Figure 3.7  BMEP with Changing PSC Flow-Rate, 2500rpm  38  Figure 3.8 Figure 3.9 Figure 3.10  BMEP with Changing PSC Timing, 2500rpm Gross IMEP from In-cylinder Pressure Data, 2500rpm Sample P-V Diagram at A= 1.10, Homogeneous Fuelling, 2500rpm  38 39 39  Sample logP-logV Diagram at A=l. 10, Homogeneous Fuelling, 2500rpm  40  Coefficient of Variation of Indicated Mean Effective Pressure, 2500rpm  40  Figure 3.2  Figure 3.11 Figure 3.12  vi  Figure 3.13 Figure 3.14 Figure 3.15 Figure 3.16  Carbon Monoxide Emissions with changing PSC Flow-Rate, 2500rpm  41  Carbon Monoxide Emissions with changing PSC Timing, 2500rpm  41  Total Hydrocarbons Emissions with Changing PSC Flow-Rate, 2500rpm  42  Total Hydrocarbons Emissions with Changing PSC Timing, 2500rpm  42  Figure 3.17  Comparison between tHC and CH4 Emissions, 2500rpm  43  Figure 3.18  Oxides of Nitrogen with Changing PSC Flow-Rate (Siemens Analyser), 2500rpm Relative Magnitude of NOx Emissions (Siemens Analyser), 2500rpm  43  Figure 3.19 Figure 3.20  44  Maximum In-cylinder Pressure for Homogeneous and PSC Fuelling, 2500rpm  44  Figure 3.21  BSFC with Changing PSC Flow-Rate, 2000rpm  45  Figure 3.22  Thermal Efficiency with Changing PSC Flow-Rate, 2000rpm  45  Figure 3.23  BMEP with Changing PSC Flow-Rate, 2000rpm  46  Figure 3.24  CO with Changing PSC Flow-Rate, 2000rpm  46  Figure 3.25  tHC with Changing PSC Flow-Rate, 2000rpm  47  Figure 3.26  N O with Changing PSC Flow-Rate (API Analyser), 2000rpm  47  Figure 3.27  BSFC with Changing PSC Flow-Rate, 15OOrpm  48  Figure 3.28  CO with Changing PSC Flow-Rate, 1500rpm  48  Figure 3.29  tHC with Changing PSC Flow-Rate, 1500rpm  49  Figure 3.30  N O with Changing PSC Flow-Rate (API Analyser), 1500rpm  49  Figure 3.31  BSFC with Highest Possible PSC Flow-Rate, 3000rpm  50  x  x  vii  Figure 3.32  CO with Highest Possible PSC Flow-Rate, 3OOOrpm  50  Figure 3.33  tHC with Highest Possible PSC Flow-Rate, 3000rpm  51  Figure 3.34  N O with Highest Possible PSC Flow-Rate (API Analyser), 3 OOOrpm  51  Coefficient of Variation of Maximum Cylinder Pressure, 2500rpm  52  Crank Angle at which Maximum Cylinder Pressure Occurs, 2500rpm  52  Figure 3.35 Figure 3.36 Figure 3.37  x  Coefficient of Variation of Crank Angle at Which P  max  Occurs,  2500rpm  53  Figure 3.38  In-cylinder Pressure Data; comparison at X=l .65, 2500rpm  53  Figure 3.39  Heat Release Rate vs. Crank Angle; comparison at A,=T .65, 2500rpm Mass Fraction Burned vs. Crank Angle; comparison at A.=1.65,2500rpm  54  Figure 3.40 Figure 3.41  54  Combustion Duration (5-95% of Mass Burned Fraction), 2500rpm  55  Figure 3.42  Ignition Delay vs. Relative Air-Fuel Ratio, 2500rpm  55  Figured  Sample Test Sheet  72  Figure F.l  Schematic of Engine Intake and Exhaust Geometry  83  Figure F.2  Ford Festiva (1998) Piston Geometry  84  viii  NOMENCLATURE  SYMBOLS  m  Mass flow-rate of fuel  N  Engine Speed  P  Pressure  Pb  Brake Power  QLHV  Lower heating value  R  Gas constant  T  Temperature  Tb  Brake torque  V  Volume  f  V  a  Displaced cylinder volume  r|f  Fuel conversion efficiency (also called thermal efficiency)  X  Relative air-fuel ratio  o  Standard deviation  COR  Uncertainty in a given result R  ABBREVIATIONS AFR  Air-fuel ratio  ABDC  After bottom dead centre  ATDC  After top dead centre  B.C.  British Columbia  BBDC  Before bottom dead centre  BDC  Bottom dead centre  BTDC  Before top dead centre  BMEP  Brake mean effective pressure  BOI  Beginning of injection  BOS  Beginning of signal  BSFC  Brake specific fuel consumption  ca.  Crank angle  CH  4  Methane  CI  Compression ignition  CNG  Compressed natural gas  CO  Carbon monoxide  C0  Carbon dioxide  2  COV  Coefficient of variation  CR  Compression ratio  EOI  End of injection  EOS  End of signal  FID  Flame ionisation detector  FS  Full scale  H  Hydrogen  2  HC  Hydrocarbons  H/C  Hydrogen to carbon ratio  HHV  Higher heating value  IC  Internal combustion  IMEP  Indicated mean effective pressure  LFE  Laminar flow element  LHV  Lower heating value  LML  Lean misfire limit  MBT  Minimum (spark advance) for best torque  MFB  Mass fraction burned  NDIR  Non-dispersive infra red  nmHC  Non-methane Hydrocarbons  NOx  Oxides of Nitrogen (mainly NO and NO2)  0  2  Oxygen  PSC  Partially Stratified Charge  RAFR  Relative air-fuel ratio  SI  Spark ignition  TDC  Top dead centre  tHC  Total hydrocarbons  UBC  University of British Columbia  SUBSCRIPTS a  air  b  brake  e  exhaust  f  fuel  i  intake  max  maximum  min  minimum  ACKNOWLEDGEMENTS First and foremost I would like to thank my supervisor, Dr. Robert L. Evans, for his guidance and encouragement throughout my studies at UBC. Dr. Fernando de Castro assisted me with a wealth of practical knowledge and Gordon Wright was invaluable in designing and implementing both hardware and software for the data acquisition system, as well as helping me with seemingly endless troubleshooting. I greatly appreciate Greg Brown's work developing software to speed the data-processing stage. I would like to extend my gratitude to all the faculty staff who contributed to the skilled and material requirements of my research, including Jeff Kohne, Richard Vandolder, Doug Yuen, Perry Yabuno and all the workshop technicians. I am very grateful for the support given to the project by Westport Innovations Incorporated, and in particular to Dr. Sandeep Munshi for his clear explanations and advice. Thanks to all my colleagues and friends who made studying at UBC such a rewarding experience, and especially to Cam Shute and all my friends that I met through the Varsity Outdoor Club. I feel fortunate to have spent time with such great people. I would like to acknowledge my parents, who have always encouraged me and supported my decisions. Finally, thanks to Rebecca Goulding. You're the best!  xii  Chapter 1 INTRODUCTION  1.1  Overview and Objectives  The world's transportation system is dependent on petroleum based fuels, a resource that is finite and whose reserves are diminishing. The oil crisis of the 1970s highlighted this dependence and provided the initial motivation to research various alternative fuels. More recently, emissions regulations that are becoming increasingly stringent have maintained this incentive. Interest in compressed natural gas (CNG) as an automotive fuel has grown during this time, prompted by its relative cleanliness and low cost as well as the desire, in some countries, to reduce the dependence on imported oil (Nichols, 1993, Unichetal., 1993). Previous CNG-fuelled internal combustion (IC) engines have mostly been conversions from a base diesel compression-ignition engine or a gasoline spark-ignition engine. Compression ignition (CI) engines have been adapted by fitting a fuel injection system to the intake manifold and replacing the diesel injector in the cylinder head with a spark plug. This makes use of the higher compression ratio in a CI engine, which suits CNG combustion. Quite a large number of 'two-tank' spark-ignition vehicles are currently in use - in the US postal service, for example - that can run on either gasoline or CNG (Dinh, 1994). They can take advantage of the cheaper cost of CNG where refuelling is possible, and the driver can choose which fuel to use with a flick of a switch. But these engines are usually optimised for gasoline and as a result run with much lower efficiency and power than it is possible to achieve when fuelled with CNG alone. It is apparent that while natural gas is suitable for fuelling a modern IC engine, simple conversion from a base diesel or gasoline configuration is not an ideal solution. It is better to develop a high-efficiency, low-emission engine technology that is dedicated to natural gas fuelling (Weaver, 1989, Ingersoll, 1996).  Homogeneous-charge spark  1  ignition (SI) engines fuelled with CNG have shown promise in improving thermal efficiency and reducing exhaust emissions compared to gasoline-fuelled engines. This is especially the case when the air-fuel ratio (AFR) is 'lean', that is, when more air is present than is required for stoichiometric combustion (Gupta and Bell, 1994). The limit to how lean we can make the air-fuel mixture is called the 'lean misfire limit' (LML), which is defined as  COVIMEP  = 10%.  This is the excess-air ratio beyond which  combustion quality - and hence engine performance - is adversely affected. An extension of the LML is desirable as it allows the engine to be run at lower loads without use of a throttle, hence reducing pumping losses. One method of achieving such an extension is to provide a relatively rich air-fuel mixture in the vicinity of the spark plug, while maintaining an ultra-lean homogeneous charge in the main chamber area (Green and Zavier, 1992, Arcoumanis et al., 1997, Evans, 1999b,). This is known as the 'Partially Stratified-Charge' (PSC) technique. The concept involves injecting a small percentage of the overall fuel mass in the region of the spark plug just prior to ignition. It requires no modification to basic engine design, but only to the method of introducing fuel to the combustion chamber. Previous attempts to examine the effects of the PSC technique have indicated that this is a promising concept that may improve engine performance and reduce exhaust emissions for a given air-fuel ratio. The objective of this study is to extend current knowledge on the subject and quantify the effects of the PSC technique on a SI engine fuelled with natural gas. Specifically, we addressed the following research questions: (a) does using the Partially Stratified-Charge technique in a spark-ignition engine fuelled with natural gas result in any of the following performance improvements: (i) an extension of the lean misfire limit, (ii) an increase in thermal efficiency and engine power, and/or (iii) a reduction in exhaust emissions? (b) if so, is there evidence of an improvement in combustion quality as determined by incylinder pressure measurement?  2  1.2  Literature Review  1.2.1  Natural Gas Composition and Properties  Natural gas is composed mainly of methane, (typically 84 to 99% composition by mole fraction) and also contains ethane, propane, carbon dioxide and nitrogen. In British Columbia, it contains approximately 96% methane. Variation in composition can occur in a single source or from different fuelling sources if a vehicle ranges geographically, and may be large enough to produce significant changes in the stoichiometry of the fuel and its octane number. The natural gas supply composition in B.C. is quite consistent, however, as can be seen if samples from 1993 and 2001 are compared, (Table 1.1). Table 1.1 Average Composition of Natural Gas (Tilbury Gate Station, B.C.) Component  Symbol  Mole %  Mole %  (May 1993)  (May 2001)  95.5  96.36  Methane  CH  Ethane  C2H-6  3.0  1.91  Propane  CH  0.5  0.41  Butane  C4H10  0.15  0.15  Pentane  C5H12  0.04  0.04  Carbon Dioxide  C0  0.2  0.20  Nitrogen  N  0.6  0.89  Other  -  0.01  0.04  4  3  2  2  8  The stoichiometric air-fuel ratio (by mass) for combustion of dry B.C. natural gas in dry air is 16.8:1 (See Appendix G). In comparison, a typical value for the stoichiometric airfuel ratio of gasoline is 14.6:1. This means that a given mass of CNG requires at least 15% more air for complete combustion. Consequently, there will be proportionally less CNG inducted into the combustion chamber and the heat release from combustion will be less.  3  An important property of natural gas is that it is in a gaseous state at atmospheric temperature and pressure, unlike gasoline or diesel, which are liquids under the same conditions. There are a number of advantages to using gaseous fuels. Gaseous fuels generally offer cleaner combustion due to improved fuel-air mixing and a higher hydrogen to carbon ratio than is found in conventional liquid fuels, (less carbon in the fuel implies less carbon dioxide emissions).  An engine running on CNG emits a  negligible amount of soot and about 25% less carbon dioxide ( C O 2 ) than a diesel engine of the same efficiency (Kapus and Chmela, 1994). Cold-start enrichment and fuel vaporisation prior to combustion are not required. CNG emits only 35% more total hydrocarbons (tHC) atfirststart than when warmed-up, compared with 160% more for gasoline (Raine et al., 1997). This is an important characteristic of an engine running on CNG, because the exhaust emissions that are released during start-up are mostly unburned methane. Methane is considered quite unreactive. It is a 'greenhouse gas', however, and is believed to contribute to global warming. Finally, engine safety is increased. CNG is lighter than air and will dissipate rapidly in the event of leakage. The principle disadvantage of a gaseous fuel is that when it mixes with air in the intake manifold of a spark-ignition engine, a significant volume of air is displaced. For a given engine displacement the amount of air-fuel mixture that can be inducted and burnt per cycle is reduced by about 10% (Weaver, 1989). The volumetric efficiency of the engine is therefore about 10% less, with a corresponding reduction in engine power output. In dedicated CNG engines, power loss due to this reduction in volumetric efficiency can be overcome by increasing the compression ratio. Methane is a particularly stable molecule and so has a very high octane number, 120, by far the highest of any commonly used fuel. A fuel with a high octane number has good resistance to auto-ignition (knock) under conditions of high temperature and pressure. As a consequence, when fuelling a vehicle on CNG the engine compression ratio may be increased from that of a gasoline engine usually around 10:1 - to as much as 18:1 (Fleming and O'Neal, 1985). This increase gives higher torque and thermal efficiency, but it can have the undesirable effect of causing an increase in the formation of oxides of nitrogen (NOx). The high temperatures and pressures associated with such a high compression ratio can also put undesirable stresses on engine components. Testing indicates that a compression ratio of about 15:1  4  is the optimum for methane (Stone and Ladommatos, 1991). However, a balance must be reached such that power and efficiency are maximised for acceptable levels of NOx (Takagaki and Raine, 1997). For this reason the UBC single cylinder research engine has been designed to have a compression ratio of approximately 12:1. Table 1.2 Comparison of Properties of B.C. Natural Gas, Pure Methane and Gasoline Fuel Property  Natural Gas  Methane  Gasoline  Molecular Weight  16.61  16.04  -110  H/C Ratio (mol/mol)  3.924  4.0  1.87  H H V (MJ/kg)  54.44  55.5  47.3  L H V (MJ/kg)  49.11  50.0  44.0  Stoichiometric A / F  16.81  17.23  14.6  Research Octane Number  -120  120  91-99  Motor Octane Number  -120  120  82-89  Natural gas exhibits lower laminar flame speeds than conventional hydrocarbon fuels. As an illustration of this property, a stoichiometric mixture of methane and air at atmospheric pressure will have a laminar flame speed ( S L ) of approximately 35 cm/s, compared with S > 50 cm/s for conventional hydrocarbon fuels, (Strahle, 1993). Slow L  combustion inhibits engine performance because the fuel that bums before piston top dead centre (TDC) increases the work required for compression, while that burning too late in the cycle performs less work on the piston during expansion. Methane flame speed is also strongly affected by both pressure and air-fuel ratio. A stoichiometric mixture at one bar has S L = 35 cm/s, but at 20 bar S drops to 15 cm/s. Similarly, the L  flame speed at 1 bar is reduced to 20 cm/s when the air-fuel ratio is leaned out to A,=1.4. In a lean-burn engine, combustion obviously takes place at high pressures in lean mixtures. In order to deal with the slow flame speed the spark timing may be advanced, but a better solution is to design the combustion chamber such that the burn time is minimised. Fast-burn combustion chambers make use of intense small-scale turbulence to increase the rate of combustion. Careful design is necessary to achieve this turbulence without increasing heat-loss to the combustion chamber walls or increasing the mean flow velocity near the spark (Gambino et al., 1993, Das and Watson, 1997, Evans, 5  1999a). Compact combustion chambers are important because they decrease the distance the flame has to travel. If hydrogen - a fast-burning fuel - is added to natural gas the combustion speed of the resulting fuel is increased. This gaseous mixture has been called 'hythane', and typically contains about 15-25% hydrogen gas by volume. Use of this fuel can increase engine efficiency and reduce the exhaust concentrations of regulated pollutants (Larsen and Wallace, 1994). N O was significantly reduced by 13-32% and x  there was a reduction of 5-13% in total HC emissions. It was found that there was little increase in pre-ignition activity (which can lead to knock), due to the hydrogen addition.  1.2.2  Natural Gas Fuelling Strategies  Lean fuelling - that is, combustion with excess air - of SI engines has a number of advantages over stoichiometric fuelling. CNG has a relatively poor lean flammability limit, (5.3% by volume in air compared with 0.6% for gasoline). Nonetheless, a leanburn strategy with CNG can work well - provided the air-fuel mixture is not too lean and improve engine performance and exhaust emissions characteristics. Lean operation can increase thermal efficiency by reducing throttling (pumping) losses during part-load operation. Near throttle-less operation can be achieved if the air-fuel ratio is changed to match the load. Other advantages of lean fuelling are that it decreases the likelihood of knock, reduces emissions (especially NOx), and may permit the use of higher compression ratios while reducing heat transfer to the cylinder walls.  However there are a number of difficulties associated with lean-burn operation, which arise from both the unfavourable lean flammability limit and the slower flame propagation speed of natural gas. As the mixture is leaned out beyond a certain point, (relative air-fuel ratio (k) of approximately 1.6), further efficiency gain of lean operation cannot be realised. Less complete combustion, increased cyclic variations and eventually misfiring will occur, as well as increased levels of carbon monoxide (CO) and total hydrocarbons. Therefore careful engine design to address these issues is vital. Some examples include improvement of fuel-air mixing, changing the combustion chamber configuration and generation of turbulence.  High turbulence levels can increase  6  efficiency and reduce emissions as the mixture is leaned out. The Nebula combustion chamber, which creates high swirl and then converts it to intense turbulence as the piston approaches TDC, addresses this problem (Kingston Jones and Heaton, 1989). This design has a fast-burn capability that allows the mixture to be leaned out to X of approximately 1.7. Previous work in the department of Mechanical Engineering at UBC (Evans and Blaszczyk, 1997) has resulted in the design of a family of fast-burn "squishjet" combustion chambers, which are specifically designed to increase turbulence generation prior to ignition, and during the combustion period. These fast-burn chambers were found to increase thermal efficiency by as much as 5%. Ignition quality, which influences ignition delay and cyclic variability, can be improved in general by providing a quiescent area at the location of the spark plug (Kapus and Chmela, 1994). This is an added complication to combustion chamber design, as it has been shown that a high level of turbulence is required in the rest of the cylinder for fast, complete combustion. Lean operation may also eliminate the need for expensive exhaust catalytic treatment in some applications. Current SI transportation engines generally use a three-way catalyst in the exhaust stream to decrease NOx, CO and tHC emissions to acceptable levels. Three-way catalytic converters are very effective, (with conversion efficiencies of 90% or more), but they require the air-fuel ratio to be very close to stoichiometric. With lean fuelling of natural gas, CO and HC emissions are reduced due to the excess oxygen available for combustion. The excess oxygen can also provide further oxidation in the later portions of the cycle and in the exhaust system. Use of oxidation-only catalysts, which are less expensive than three-way catalysts, could further reduce emissions. Measurements of in-cylinder HC concentrations by gas sampling prior to exhaust valve opening show levels of about 1.5 to 2 times the average tailpipe exhaust level (Heywood, 1988). NOx emissions are reduced with lean-burning as a result of the lower in-cylinder gas temperatures due to the excess air. Lean burning of CNG, in conjunction with a high turbulence combustion chamber, has been shown to meet stringent European NOx limits without penalties in thermal efficiency through excessive ignition retard (Kingston Jones and Heaton, 1989). However, the stability of combustion deteriorates considerably (i.e. cyclic variation and misfiring), as the lean misfire limit is approached.  7  An eventual goal of lean fuelling is to reduce emissions to a level where a catalyst may be dispensed with and to obtain throttle-less operation - at least for a wide range of operating loads. Very lean combustion is required in order to attain this, which is why much research is directed towards extending the lean misfire limit. The technique of local charge stratification was tested in a constant volume combustion chamber, and was found to greatly stabilise combustion for ultra-lean air-fuel ratios (Arcoumanis, Hull and Whitelaw, 1994). This was then applied to a lean-burn spark ignition test engine fuelled with propane (Arcoumanis, Hull and Whitelaw, 1997). It was found that a stable and consistent flame kernel formed after ignition due to the small quantity of rich mixture injected near the spark gap, allowing flame propagation through a very lean mixture, (X of greater than 2.5 in the constant volume chamber, and 1.8 in the test engine). Another partially stratified-charge concept uses injection of methane or natural gas through the spark plug (Green and Zavier, 1992, Evans, 1999a, Evans, 1999b). The advantage of this technique is that no modification to the basic engine design is required, only to the fuel distribution system. This is the basis for this M.A.Sc. research project. In general, the stratified-charge concept allows easy ignition and good flame growth characteristics with ultra lean fuelling. The previous work on this topic has indicated that it provides an improvement in engine performance, a reduction in exhaust emissions and extension of the lean misfire limit. In conclusion, use of natural gas as an IC engine fuel has many proven advantages. There is currently more interest in designing engines for dedicated fuelling with natural gas, although large scale engine re-design is not desirable, as it would be a significant barrier to production. Fortunately this is not necessary. There are a number of ways existing spark-ignition engine types can be optimised for CNG fuelling in order to reduce emissions without impairing efficiency or power output. These include compression ratio, lean-burn technology, appropriate combustion chamber design and an engine management system that uses closed loop control of fuel injection (Patrick et al., 1991,  8  Graboski et al., 1997). Further improvements to engine performance would certainly be achieved by implementing turbocharging with intercooling. On-board fuel storage, vehicle range, refuelling infrastructure and the economics of vehicle and fuel costs are clearly important factors that would influence CNG fuelled vehicle production. Although they are beyond the scope of this study, it is obvious that private- ownership of CNG fuelled vehicles is not an immediate practical goal. It is realistic, however, to target this engine technology as a solution for medium-duty vehicle fleets within a city boundary. With careful thought about appropriate applications, CNG fuelled vehicle technology can become a realistic transport solution for the near future.  9  Chapter 2 EXPERIMENTAL FACILITY AND DATA ANALYSIS  2.1  Introduction  The experimental work reported here was conducted in the Department of Mechanical Engineering at the University of British Columbia. The research undertaken used a spark-ignited Ricardo Hydra single cylinder research engine, fuelled by natural gas. The test engine was fully instrumented, including a piezoelectric pressure transducer for highspeed measurements of in-cylinder pressures. A comprehensive emissions bench enabled measurement of exhaust levels of oxygen, carbon dioxide, carbon monoxide, total hydrocarbons and oxides of nitrogen. Experimental data were acquired using a National Instruments multi-channel system. Measurements were then displayed, monitored and recorded by a Labview interface developed at UBC for the Ricardo Hydra. The data were processed and analysed using Microsoft Excel spreadsheets developed by Westport Innovations Inc. and adapted for the UBC Ricardo Hydra. 2.2  Research Engine  The engine used for the reported research was a Ricardo Hydra single cylinder research engine (serial number 30) coupled to a 30kW dynamometer. The dynamometer system provided for both power adsorption and motoring, and a controller was used to maintain a constant test speed. Fig. 2.1 is a schematic of the Ricardo Hydra test engine in its current configuration. General engine specifications are given in Table 2.1. The test engine was rebuilt in-house to prepare for natural gas fuelling. This included reboring the cylinder lining so that a standard Ford Festiva piston could be used, albeit with modifications to the piston bowl. The combustion chamber was a bowl-in-piston type with compression ratio of 11.9 and a squish ratio of 47.4%, (see Appendix F - Engine Geometry - for further details).  10  Two Stage Pressure Regulator ( P (-170 bar to 27 bar)  Bladder-type Accumulator Laminar Flow != Element  Main CNG Fuelling RAFR sensor and exhaust emissions sample  In-cylinder pressure transducer  Spark-Plug/ Injector  Crankshaft  (Optical Shaft Encoder on crankshaft measures speed and position)  Figure 2.1 Schematic of Ricardo Hydra Test Engine with PSC Fuelling System  11  Table 2.1 UBC Ricardo Hydra Engine Specifications Number of cylinders  1  Bore  81.12 mm  Stroke  88.90 mm  Connecting rod length  158.01 mm  Swept volume  459.46 cc  Clearance volume  41.988 cc  Compression ratio  11.9:1  Squish ratio  47.4%  Rated speed  5400rpm  Rated power  15 kW  Valve timing:  2.3  Inlet opens  12° BTDC  Inlet closes .  56° ABDC  Exhaust opens  56° BBDC  Exhaust closes  12° ATDC  Partially Stratified-Charge System  The Partially Stratified-Charge (PSC) System used a high-pressure natural gas line to supply the spark-plug injector assembly mounted in the research engine. The highpressure injection gas passed through a thermal mass flow-meter and an accumulator to damp injection oscillations. The plug-injector assembly was comprised of a fast-response solenoid valve, a check valve to prevent back-flow of hot combustion gases and the spark plug, which contained a capillary tube to direct the injected fuel to the region of the spark (Fig. 2.2). A controller built in-house allowed the characteristics of both the spark (dwell and discharge duration, timing and number of sparks) and the injection (duration and endof-injection timing) to be set by the operator.  12  The timing characteristics of the PSC injection were not measured in real time, but were pre-determined using an experimental set-up shown in Fig. 2.3. A good understanding of the injection control system was necessary because there was a delay between when the voltage signal was sent to the solenoid, and the time when the flow actually began. This delay was made up of two components: first, there was a delay while the current rose to a value sufficient to open the solenoid, and secondly there was an inertial delay related to the mass of the moving parts in the solenoid. Although these delays are on the order of milliseconds, PSC injection duration is only a couple of milliseconds and at 2500rpm the piston takes 12 milliseconds to go from BDC to TDC. In practice, the delay times were found to be very significant indeed, and a limiting factor when it came to the quantity of fuel that could be injected through the spark-plug.  r  13  Two Stage Pressure Regulator (CNG pressure set at 10 bar - representative differential pressure experienced by solenoid during engine operation. Solenoid Valve  Voltage control signal  Check Valve Fast response pressure transducer (piezoelectric type)  Oscilloscope Vent to Atmosphere  Pressure signal (indication of flow)  Figure 2.3 Experimental Apparatus to Determine Injection Delay Characteristics  The results of the timing tests were as follows. The delay from beginning of signal (BOS) to beginning of injection (BOI) was found to be independent of speed, differential pressure and injection duration. This constant delay was 5.4 milliseconds. The delay from end of signal (EOS) to end of injection (EOI), however, was found to be independent of speed and differential pressure but was a function of signal duration. A fourth order polynomial made the best fit with the experimental data (R = 0.992), and 2  describes well how the delay approaches a constant for long injection duration, (Fig. 2.4).  14  20  0-1  ,  ,  ,  ,  1  3  4  5  6  7  8  Signal Duration (mS)  Figure 2.4 Empirical Data showing Injector Delay vs. Signal Duration The table below (Table 2.2) shows settings required to control injection timing values in the PSC control system at 2500rpm. It takes into account the delays that were determined at both the beginning and the end of injection. For example, if an injection duration of approximately 15 degrees (1.0 millisecond) which ends at TDC was required, it was necessary to set the signal duration to 3.84 milliseconds and the end-of-signal timing to 121 degrees after BDC (59 degrees before TDC). Table 2.2 Calculated PSC Control Settings (Based on Solenoid Delay), 2500rpm Speed: 2500rpm  mS/rev: 24 mS  deg/mS: 15 deg  Approx. Injection Duration (degrees) Approx. Injection Duration (mS) End of Injection (deg. before spark)  5 0.33 0  10 0.67 0  15 1 0  20 1.33 0  Enter in Control Box: Signal Duration (mS) End of Signal (deg. ABDC)  3.61 128  3.73 125  3.84 121  3.95 118  15  Extensive running with PSC injection using these settings indicated good EOI accuracy. In general, this calibration of PSC injection was found to be satisfactory for the testing in this study. For future PSC testing it might be preferable to have a way of determining in real time when injection starts and ends. This could be implemented by permanently installing a second piezoelectric-type pressure transducer in the injector fuel line after the solenoid valve, as described in Fig. 2.3. Observation of pressure changes in the fuel line would enable real-time identification of injection start- and end-points. This could be linked to the high-speed data acquisition system and be displayed in Labview in a similar way to the in-cylinder pressure. 2.4  Instrumentation  The test engine was fully instrumented for research purposes. A quartz piezoelectric pressure transducer was mounted in the head for high-speed, high-resolution in-cylinder pressure measurements. The transducer had a heat shield to reduce the effects of thermal shock. In order to reference the rapidly changing pressure to crankshaft position, a shaft encoder was installed and set to a resolution of 0.5 crank angle degrees. Hence the cylinder pressure was measured 1440 times each engine cycle (two crankshaft revolutions).  The shaft encoder also measured engine speed.  Engine torque was  measured using a calibrated strain-gauge on the dynamometer. Both the high-pressure injection flow of CNG and the main "homogeneous" fuel flow were measured using thermal mass flow-meters. A laminar flow element (LFE) was used to measure the intake flow rate, with a six-litre surge tank to dampen intake oscillations. Two fine wire-mesh flame traps were positioned above and below the surge tank to prevent backfire from damaging the LFE. The differential pressure across the LFE is proportional to the volumetric flow of air through the intake, and was measured using a differential-pressure transducer.  Secondary measurement of the air-fuel ratio near  stoichiometric conditions was made with an AFR sensor, which has a sensor in the exhaust stream close to the exhaust manifold. This instrument made it possible to crosscheck the accuracy of our air and fuel flow-rate measurements near stoichiometric  16  conditions, as the AFR sensor is not very accurate when the engine is operating near the lean misfire limit. A full list of all parameters measured and recorded during testing is presented in Appendix E, including the temperatures, pressures and ambient conditions that are not mentioned here. Full details of the engine instrumentation are presented below in Table 2.3.  Table 2.3  Ricardo Hydra Engine - Main Instrumentation Specifications  Instrument  Make  Model  Range  Uncertainty  In-Cylinder Pressure Transducer  PCB Piezotronics  112B10  0 - 200 bar  + 1 bar (0.5%ofFS)  Optical Shaft Encoder (Engine Speed and Crankshaft Position)  US Digital  H1-360-IE  0 - 10,000 rpm  ± 5 rpm (speed) ± 1 deg (position)  Strain Gauge (Engine Torque)  Strain Gauge  n/a  0 - 50 Nm  ± 0.5 Nm (1.0%ofFS)  Relative Air-Fuel Ratio from Exhaust  ECM  AFRecorder 2400G  0.7-2.0 (X)  ±0.01 (0.8<\<\.2) ± 0.05 (elsewhere)  Main NG Mass Flow Meter/Controller  MKS Instruments  1559A-100C-SV  0 - 100 slm  + 0.6 slm (0.6%ofFS)  PSC NG Mass Flow Meter  MKS Instruments  179A-24C-S3BM  0-20 slm  ± 0 . 1 2 slm (0.6%ofFS)  Intake Air Flow  Meriam  50MW 20-1.5  0-30 scfm  ± 0.3 scfm (1.0%ofFS)  Differential Pressure Transducer  AutoTran  600 D-014  0 - 20" water  ± 1.0% (FS)  Gauge Pressure Transducer (Int./Exh.)  AutoTran  600 D-018  ± 1.0 bar  ± 1.0% (FS)  PSC Solenoid Valve  Omega  SV122 .  n/a  n/a  Accumulator on HP Natural gas Line  Hydac  SB330  0 - 3000 psi  n/a  Note: "1.0% of FS" refers to the uncertainty as a percentage of Full Scale output.  Exhaust emissions were measured using a comprehensive emissions bench. The emissions bench had the capability to measure oxygen, carbon dioxide, carbon monoxide, 17  total hydrocarbons and oxides of nitrogen. The main exhaust gases of interest for this study were carbon monoxide, total hydrocarbons and oxides of nitrogen. The analyser specifications are listed in Table 2.4. Table 2.4 Emissions Bench - Analyser Specifications Exhaust Component  Make  Model  Operating  Range  Uncertainty  Principle  Carbon Monoxide (CO)  Siemens  Ultramat 21P  NDIR  0 - 10,000ppm  ±25ppm (0.25% of FS)  Total Hydrocarbons (tHC)  Ratfisch  RS-55  FID  0 - 10,000ppm  ±100ppm (1.0%ofFS)  Oxides of Nitrogen (NOx)  API  200AH  Chemiluminescent  0-3,000ppm  ± 15ppm (0.5%ofFS)  Oxides of Nitrogen (NOx)  Siemens Ultramat (old instr.) 22P  NDIR  0 - 3,000ppm  ± 50ppm (1.7%ofFS)  Oxygen (O )  Siemens  Oxymat5E  Paramagnetic Alternating Pressure  0-21%  ±0.2% (1.0%ofFS)  Beckman  880  NDIR  0 - 20%  ± 0.2% (1.0%ofFS)  z  Carbon Dioxide (C0 ) 2  Note: "1.0% of FS" refers to the uncertainty as a percentage of Full Scale output.  2.5  Data Acquisition, User Interface and Data Analysis  The data acquisition system was a 47-channel National Instruments system. Data could be acquired at two speeds; a "low" speed, (approximately 80 samples per minute), for the fuel and air flow-rates, temperatures, etc. and a "high" speed, (1440 measurements per cycle), for the in-cylinder pressure. The data acquisition system was linked to a computer running Labview software, and there the signals were processed to convert them to meaningful values. A Labview interface on the control room computer allowed the user to record the data in text file format when a required engine running condition had stabilised. One hundred 18  consecutive cycles o f in-cylinder pressure data were acquired when the user specified. For the "low-speed" data, (also referred to as the "performance" data), the user sampled data for a predetermined time. The low-speed data for this study was all recorded for a test time o f two minutes, which corresponds to approximately 160 samples o f data per test point.  The raw performance data was first visually inspected and then averaged using a Microsoft Excel "macro". This macro also generated standard deviations and coefficients of variation for each test point, which could be used as an indication o f the quality and reliability o f each measurement.  The averaged results were inserted into a spreadsheet  developed by Westport Innovations Inc. and adapted for use with the Ricardo Hydra at UBC.  This spreadsheet calculated a comprehensive range o f parameters under the  headings " A i r and Fuel F l o w " , "Volumetric Efficiency", "Performance and Efficiency", "Combustion Analysis", "Volumetric (Wet-Basis) Emissions" and "Exhaust Emission Flows". The main performance parameters o f interest for this study were: Brake Specific Fuel Consumption ( B S F C ) - a measure o f engine efficiency, Brake M e a n Effective Pressure ( B M E P ) - a measure o f engine torque, and the exhaust emissions C O , t H C and NOx-  They are all presented with respect to the relative air-fuel ratio (k), which is  defined by E q . 2.1.  x  =  _J^J\c ua _ l  (Eq. 2.1)  L  (A IF) stoichiometric  One hundred consecutive cycles o f in-cylinder pressure data - "high-speed" data - were acquired for each test point. This data was processed using another Westport Innovations Inc. program, this time in Fortran code. The program outputs two results files i n text format. The first is the averaged cylinder pressure values at half crank-angle resolution, which is used to produce plots o f P vs. V , P vs. crank angle, log-P vs. log-V and heat release rate. A s the properties o f the fuel are known, it was also possible to calculate and plot heat release and heat transfer vs. crank angle. The second results file is a summary of some statistical properties o f the data.  These properties are the maximum cylinder  pressure (P x), the crank angle o f P x , the polytropic exponents for the compression and ma  ma  19  expansion processes and the indicated mean effective pressure ( I M E P ) . A l s o calculated are the coefficients o f variation ( C o V ) o f P  m a x  , o f the crank angle o f P  m a x  and o f I M E P .  These data are presented i n summary plots that show how they change as we extend the relative air-fuel ratio to the lean misfire limit. A l l calculated data are corrected according to S A E standards J1349, J1312 and J177.  2.6  Uncertainties and Repeatability  A study was performed to evaluate the uncertainties due to instrument and experimental errors. Uncertainties are expressed i n terms o f a plus or minus range - within which we are confident that the true value lies - and have been determined for the principal performance and emissions data to be presented.  The analysis takes into account the  estimated uncertainties in individual measurements. A n overall uncertainty estimate for a given parameter is calculated as follows:  If the parameter R is the result o f a given function o f the independent variables xi,X2, ...x„, then the uncertainty i n the result is given by:  (Eq. 2.2)  oo =< R  where OOR is the uncertainty i n the result and ooi, O02,...oo„ are the uncertainties i n the independent variables (Holman, 2000).  F o r example, the engine brake power, PA, is  defined as: 2n x N x T  u  60x1000  (kW)  (Eq. 2.3)  where N is the engine speed i n rpm and Tb is the brake torque in N m . Using E q . (2.2) we get the uncertainty for Pi, (Eq. 2.4).  2n  oo  t  60000  iNy(oo y (T y(oo yf  2  n  +  h  N  (Eq. 2.4)  20  It would be possible using this method to calculate an uncertainty value for each test point, (i.e. for each speed and torque combination). For the purposes of this analysis, however, it was deemed appropriate simply to calculate the "worst case uncertainty" for each speed tested. These could then be applied with confidence over the whole range of results for that speed. For example, it was found that at 2500rpm the largest brake power uncertainty value was ± 0.132 kW, and it occurred with the maximum motor torque. Interestingly, when the brake power uncertainties were carried forward to be used in the calculation of uncertainties for BMEP, BSFC and thermal efficiency, there was little difference in uncertainties between speeds, (Table 2.5). Note that all of the uncertainties quoted in Tables 2.3 and 2.4 are conservative. Therefore the accuracy of the performance parameters may be somewhat better than the estimated uncertainties in Table 2.5. Error bars have been included on the repeatability plots discussed below to give an indication of their magnitude.  Table 2.5 Estimated Uncertainties for Performance Parameters Engine Speed:  1500rpm  2000rpm  Parameters:  Units:  Uncertainties:  Brake Power BMEP BSFC Thermal Efficiency  (kW) (bar) (g/kW-h) (%)  +0.080 ±0.142 ±7.8 ±0.91  ±0.106 ±0.140 ±8.0 ±0.85  2500rpm  ±0.132 ±0.139 ±7.1 ±0.82  The repeatability of data collected at any given test point was examined by performing the same engine tests three times each on two different days. The engine speed was set at 2000rpm, and the points examined were with homogeneous fuelling at two different relative air-fuel ratios, A=1.37 and X=l.57. The results of this analysis are displayed below (Figs. 2.5 and 2.6).  21  300 -,  • Homogeneous, RAFR=1.4 (16 Oct) • Homogeneous, RAFR=1.4 (17 Oct) A Homogeneous, RAFR=1.6 (16 Oct)  2 g 0  A Homogeneous, RAFR=1.6 (17 Oct)  260  240  -I  220  :  1.3  ,  ,  ,  1.4  1.5  1.6  1 1.7  Relative Air-Fuel Ratio  Figure 2.5 Repeatability Test: BSFC vs. Relative Air-Fuel Ratio  if  m  6  5  m Homogeneous, RAFR=1.4 (16 Oct)  \ I  • Homogeneous, RAFR=1.4 (17 Oct) A Homogeneous, RAFR=1.6 (16 Oct) A Homogeneous, RAFR=1.6 (17 Oct) 4. 1.3  1.4  1.5  1.6  Relative Air-Fuel Ratio  Figure 2.6 Repeatability Test: BMEP vs. Relative Air-Fuel Ratio  1.7  A qualitative inspection of these results suggests reasonable repeatability between data from tests at the same engine running conditions, even from the tests performed on different days. At A,=1.37 the BSFC data (Fig. 2.5) shows a scatter of less than ±0.7%, and the BMEP data is also reasonable, at ±2.5%. At the leaner point, A,=1.57, both BSFC and BMEP have poorer repeatability (approximately ±3% and 4.5% respectively). This might be explained by the fact that here the engine is running very close to the lean misfire limit, which causes variable combustion quality. This study provides only an initial qualitative analysis of the repeatability and uncertainty of our results. However, it suggests that the improvements in performance observed with the implementation of PSC technology are meaningful.  23  Chapter 3 RESULTS AND DISCUSSION  3.1  Introduction  Testing of the Partially Stratified-Charge (PSC) technology was undertaken with wideopen throttle over a range of four speeds, 1500rpm, 2000rpm, 2500rpm and 3000rpm. Full-load to minimum load testing at each speed was achieved by leaning out the air-fuel ratio to the lean misfire limit, as discussed in Chapter 1. By never throttling the engine, pumping losses at part-load were minimised. The results discussed in detail below are for 2500rpm, as this is a representative speed at which the technology would most likely be applied.  Similar results from 2000rpm were obtained, but at 1500rpm no  improvement with PSC was observed. At 3000rpm, PSC control limitations meant that it was not possible to inject enough fuel to get significant results. A complete set of performance results for each test speed is included at the end of this chapter. 3.2  An Optimisation Technique Based on BSFC  The two PSC injection settings that were optimised in this experimental work were the injection flow-rate and the injection timing. Changing the "pulse-width", (that is, the duration for which the solenoid was open), varied the flow-rate of fuel through the injector. The end-of-injection (EOI) timing was varied with respect to the spark timing, which was always set at minimum advance for best torque (MBT). This is discussed in more detail in section 3.3. While the engine was being run with different injection conditions, the resulting effect on brake specific fuel consumption (BSFC) was examined. BSFC is defined as the fuel consumed per unit power output, (Eq. 3.1), and so it is effectively a measure of the engine's efficiency. mAglh)  BSFC(g/kW.h)  =- ^ l  (Eq.3.1)  24  Pb is the brake power developed by the engine and m is the total mass flow-rate of f  homogeneous plus injected fuel. First, in order to determine the optimum injection flow-rates for each speed, the EOI timing was held constant with respect to the spark. Then a number of different injection pulse-widths were tested over a range of excess-air ratios, from X=\A to the lean misfire limit (LML). It was decided to run the initialflow-rateoptimisation tests (for 2500rpm) with EOI timing set to ten degrees before the spark (Fig. 3.1). This was based on the assumption that it would be best for the injection to end a short time before ignition, allowing time for some mixing of the pure fuel with the homogeneous in-cylinder mixture. The EOI timing was then varied while keeping theflowrate at the value shown to give the best possible improvement in performance (Fig. 3.2).  These 2500rpm  optimisation results indicate that the greatest reduction in BSFC at a given relative airfuel ratio occurs with the highest injectionflowrate that is was possible to generate with the current experimental set up, that is 14g/hr at this speed. Likewise, the best EOI timing appears to be the most advanced case, at MBT-15 degrees (that is, 15 degrees before the MBT spark timing). Here again it would have been preferable to be able to continue to advance the EOI timing with respect to the spark timing. With the current control system, however, there was a limit to how advanced the beginning of injection could be. This limitation was the solenoid delay between the control signal and start of injection. In turn this affected how much fuel could be injected and the maximum EOI advance with respect to ignition timing. Nonetheless, the optimisation process revealed that PSC can improve engine BSFC - and hence efficiency - at relative air-fuel ratios greater than 1.5.  3.3  Spark  Timing Results  For all engine running conditions that were tested in this research, the spark timing was set to "minimum advance for best torque" (MBT) timing. In the homogeneous fuelling cases, MBT was determined by carefully observing the torque as the spark was retarded (moved closer to TDC). MBT was defined as the point where the torque was just about  25  to drop again from a maximum value. When running with PSC fuelling the same procedure to find MBT was followed, except that the EOI timing was always kept at the same advance before the spark. The timings determined during testing are presented below (Table 3.1). Table 3.1 Comparison of MBT Spark Timings in Degrees Crank Angle (BTDC) for Homogeneous and PSC Fuelling Speed:  1500 rpm  2000 rpm  2500 rpm  3000 rpm  RAFR  Homog.  1.00  24  28  31  34  1.10  26  29  32  36  1.20  28  31  34  38  1.30  30  33  36  41  1.40  32  32  35  34  40  39  44  41  1.50  36  36  41  37  43  42  45  44  1.55  38  46  44  47  45  1.60  42  PSC  Homog.  PSC  Homog.  PSC  Homog.  PSC  1.45  1.65  40  46  40  49  47  49  45  43  47  45  54  50  51  47  1.70  51  53  A reduction in MBT spark advance, at lean air-fuel ratios, was observed whenever the PSC technology was used. This was even true at 1500rpm, where the advantages of PSC were not necessarily obvious in other performance results. The retardation of MBT timing at 2500rpm is shown for different relative air-fuel ratios (Fig. 3.3); it is apparent beyond X=IA, and the difference increases as the air-fuel mixture is made leaner. With PSC technology at A,=1.65, it was possible to retard the spark by 5 degrees without loss of torque. In other words, the duration between initiation of combustion (the spark) and peak cylinder pressure is shorter with PSC. This was the most obvious initial indication that PSC resulted in faster and "better" combustion at lean air-fuel ratios, when compared with homogeneous fuelling.  26  3.4  Performance Results - BSFC, Efficiency and Power  There was an improvement in BSFC, efficiency and power when PSC technology was implemented. The maximum flow-rate of fuel achievable through the spark-plug at 2500rpm - approximately fifteen grams per hour - was equivalent to about 1.2% of the total fuel mass flow-rate at X-l.6.  Green and Zavier, (who conducted experiments with a  similar PSC technology), found that the best improvement in efficiency occurred with flow rates of three to four times higher than it was possible to achieve with the current PSC set-up at UBC.  At A.=1.65, the BSFC for the best PSC case at 2500rpm is  approximately 7% lower than the homogeneous charge case, (Figs. 3.1 and 3.2). The 2000rpm testing at the same relative air-fuel ratio showed an improvement of nearly 9%, (Fig. 3.21).  The results for 1500rpm and 3000rpm, however, showed no BSFC  improvement with PSC, (Figs. 3.27 and 3.31). At 3000rpm it is clear that there was not enough injected fuel - and probably not enough EOI advance - to glean any improvement with PSC. At 1500rpm this was perhaps due to an incorrect choice of EOI timing, as sufficient flow-rates were attained at this speed. Results at both 2000rpm and 2500rpm showed a reduction in BSFC beyond A=1.5 with the PSC concept.  Thermal efficiency, (also called the "fuel conversion efficiency"), is inversely proportional to BSFC. Therefore, at 1=1.65, it was improved by the same proportion as BSFC is reduced - approximately 7% at 2500rpm. The fuel conversion efficiency is defined as:  7/(%) = f  —  (Eq. 3.2)  BSFC(kg/kW-h)xQ (U/kg) IHV  QLHV  is the lower heating value of the natural gas used, (49.11 MJ/kg). The highest  efficiency for this engine (around 31.5%) occurs at A=1.4 and thereafter drops off until the lean misfire limit (Figs. 3.4 and 3.5). With the PSC cases at X=1.65, the efficiency is still over 30%.  27  Brake mean effective pressure (BMEP) is a useful measure of the relative performance of an engine - it does not depend on engine size and so can be used to compare two different engines. It is obtained by dividing the work per cycle by the displaced cylinder volume per cycle (VJ), and can also be expressed in terms of torque, (Eqs. 3.3a and 3.3b respectively).  BMEP(bar) =  2P (W)xl0 \00xV (dm )xN(rev/s) 3  A  3  d  BMEP(bar) =  ( ' \ lOOx V (dm ) 47rT  N  m  ( E q  .3  3 b )  d  As BMEP is a measure of the engine's ability to do work, it is appropriate to use a plot of BMEP vs. relative air-fuel ratio to visualise the "throttling" effect of leaning the air-fuel mixture. Even with the purely homogeneous fuelling case (Fig. 3.6), it is clear that a significant reduction in power is possible. At the LML, BMEP is 40% less than the maximum power at stoichiometric operation. Extending our ability to "throttle" the engine by leaning the air-fuel mixture is one of the primary objectives of this research, and is discussed in more detail in section 3.5. Figures 3.7 and 3.8 show the change in BMEP with PSC at 2500rpm, for 1=1.3 to the lean misfire limit. The BMEP gain with PSC is nearly as large as the improvement in thermal efficiency, that is 6-7% at A,=1.65.  The gross indicated mean effective pressure (IMEP) is also presented (Fig. 3.9). IMEP is different from BMEP in that it is obtained from in-cylinder pressure data, by integrating around a P-V curve. The area enclosed by the "power loop" of the diagram (see Fig. 3.10) represents the work delivered to the piston by the combustion gases over the compression and expansion strokes per cycle.  This is called the "gross" IMEP, to  distinguish from when we include the pumping loop losses, ("net" IMEP). Gross IMEP does not take into account the pumping and friction losses experienced by the engine, and therefore it is always higher than BMEP. It is obvious at this stage that there is a serious problem with this pressure data, as the IMEP calculated here is consistently at least one  28  bar lower than the equivalent BMEP value, and the plots have different slopes. This is most likely due to severe thermal shock, as the piezoelectric transducer is not watercooled. Thermal shock is defined as the thermally induced shift in transducer output resulting from rapid change in heat flux at the transducer face. An example of how thermal shock affects pressure data by causing a sudden decrease in transducer output is given in Fig. 3.11. A second possible reason for pressure data inaccuracies may be due to problems with "pegging" the data. Signal pegging - referencing the output to a known value - is necessary because the signal offset from the transducer changes constantly and must be calculated each cycle. However, it is still possible to learn a number of things about the combustion process from this pressure data. The calculated parameters that should be relatively unaffected by the thermal shock are the maximum cylinder pressure coefficient of variation (COV) of  (PMAX),  COVc.aPmax  PMAX, COVIMEP,  crank angle (ca.) of  PMAX,  and -fromthe heat release calculations - relative combustion duration and  ignition delay. These are all discussed in more detail in the next sections. 3.5  Extension of the Lean Misfire Limit  The BMEP results are the best way to understand the benefits of an extension to the lean misfire limit. As well as an improvement in efficiency, PSC technology permits such an extension with respect to the homogeneous case. The lean misfire limit is defined in this research as the point at which the coefficient of variation in indicated mean effective pressure  (COVIMEP)  becomes greater than ten percent.  Beyond this point cyclic  variability becomes a problem and starts to affect vehicle driveability (Heywood, 1988, p. 417). In agreement with Heywood, the Ricardo Hydra runs poorly if at all when the COVIMEP  is greater than 10%.  COVIMEP  is the standard deviation in IMEP divided by the  mean IMEP and is usually expressed in percent, (Eq. 3.4).  COV (%) imP  =^ - x l 0 0 IMEP  (Eq. 3.4)  29  The lean misfire limit was increased when the optimised P S C technique was used at 2500rpm. The magnitude of this extension was determined by looking at a plot of COVIMEP  with respect to 1 (Fig. 3.12), and is confirmed by engine stability when running  - it was approximately 5% for 2500rpm. Although it is difficult to determine trends from these data, beyond 1=1.6  P S C  appears to reduce  COVIMEP-  At 1=1.7, with  P S C  the  coefficient of variation is still within the acceptable range. At 1=1.65, P S C may reduce COVIMEP  by up to 40%,from12% to 7%. The advantage obtained from an extension of  the LML with P S C is that a given engine power can be realised with a higher excess-air ratio than is otherwise possible.  In other words, an engine using P S C technology  consumes less fuel for a desired power output than a conventional engine at the same airfuel ratio.  3.6  Emissions Results  The principal emissions examined in this study were the brake specific exhaust emissions CO, tHC and N0 . By "brake specific" we refer to the mass of the exhaust component X  per hour per kilowatt of brake power produced by the engine, and therefore the emissions results are independent of engine size and may be compared to other test results. For example: CO(glkW  h) =  C  P (kW)  O  i  g  l  K  )  (Eq. 3.5)  b  3.6.1  Carbon Monoxide  Carbon monoxide increases as a homogeneous air-fuel mixture is leaned out. However there is a progressive reduction of carbon monoxide emissions with PSC, as EOI timing is advanced relative to the MBT spark (Figs. 3.13, 3.14 and 3.24). This reduction is evident beyond 1=1.4. The greatest comparison is possible at 1=1.65, where there is a very significant 25% reduction in CO. At 1500rpm (Fig. 3.28) there is no change in CO with the PSC settings tested, but at 3000rpm there is a small improvement (Fig. 3.32). This is of interest because it is evidence of "more complete" combustion, even though the amount of PSC injection attainable at this speed was minuscule.  30  3.6.2  Total Hydrocarbons and Methane  There was also a reduction in total hydrocarbon (tHC) emissions with PSC, but this improvement is not evident until close to the lean misfire limit, (Figs. 3.15, 3.16 and 3.25). At 3000rpm (Fig. 3.33), there is a small reduction in tHC that may be interpreted in a similar way to the reduction in CO. The important point to note here is that the tHC emissions are mostly composed of unburned natural gas, which in this case is 96% methane (CH4). In fact, when the tHC and CH4 emissions are compared for the same running conditions, it is found that only about 5% of the tHC is composed of nonmethane hydrocarbons (nmHC). The nmHC emissions can be from the small quantity of oil that leaks into the combustion chamber, usually past the piston rings. While methane is recognised as a "greenhouse gas" that may contribute to global warming, it is far less reactive than nmHC, which contributes to problematic photochemical smog. It was of some concern that the relatively fuel-rich injection near the spark-plug might cause an increase in unburned natural gas and carbon monoxide, as described by Green and Zavier (1992). This was not found to be the case, however. It is likely that this was because of the smaller injection flow-rates used in this research. Also, the improved combustion quality with PSC meant that cyclic variation was reduced and more of the mixture in the cylinder is burned. The quantity of injected fuel was very small with respect to the total fuel mass burned per cycle and there was plenty of excess oxygen in the combustion products to oxidise any extra carbon monoxide or unburned hydrocarbons before the exhaust was emitted.  3.6.3  Oxides of Nitrogen  There was some increase in emissions of oxides of nitrogen (NOx) at a given X when PSC technology was implemented. This increase was close to constant at approximately 2.25 g/kW-h over the range of air-fuel ratios tested, or greater than 50% at 1=1.6, (Fig. 3.18). This was not as discouraging as it initially seemed, however, when it was considered that the production of NOx at stoichiometric fuelling was approximately 22 g/kW-h, (Fig. 3.19). Close to the lean misfire limit - whether we are using PSC technology or not - we produced less than 15% of that value. A further promising development occurred when a more accurate NOx analyser was installed, operating on  31  the principle of chemi-luminescence. Testing at 1500rpm, 2000rpm and 3000rpm with the new instrument all produced results showing that the NOx emissions near the LML were even lower (Figs. 3.26, 3.30 and 3.34, and Table 3.2). At 1=1.6 there was only 0.5 g/kW-h for each case, that is less than 3% of the stoichiometric N 0 production. In X  general, NOx was dramatically reduced as the air-fuel mixture was made leaner. This was due to the strong temperature dependence of NOx formation - and the combustion was "cooler" with more excess air present in the gas mixture. Any increase in NOx formation during PSC combustion was most likely due to in-cylinder temperature change. In order to understand this increase one must examine the in-cylinder pressure. PSC caused higher peak cylinder pressures (Fig. 3.20), which obviously means increased incylinder temperatures and hence more NOx-  It is also possible that the near-  stoichiometric combustion that takes place at ignition may be a source of additional N0  X  production. If it were decided to optimise the PSC technology for minimum NOx, the spark could be retarded from MBT until peak cylinder pressures and temperatures were lowered sufficiently to achieve this end. Engine exhaust emissions results for automotive applications are often reported in terms of grams of emission per kilometre. With a research engine it is obviously not possible to reference data to road kilometres, but emissions can be reported as "grams of emission per kilogram of fuel". These results are presented in Table 3.2.  Table 3.2 CO, tHC and N 0 (Grams of Emission per Kilogram of CNG), 1=1.65 X  2500rpm Homogeneous Emission: CO (g/kg)  2500rpm PSC Fuelling  2000rpm Homogeneous  2000rpm PSC Fuelling  21.2  16.6  24.1  19.0  tHC (g/kg)  67.5  54.4  79.7  69.0  NOx (g/kg)  8.7 (Siemens)  14.1 (Siemens)  1.4 (API)  0.1 (API)  3.7  Combustion Analysis  In-cylinder pressure data at each relative air-fuel ratio for 2500rpm was processed and analysed so that a comparison between homogeneous fuelling and the best PSC case  32  could be made. Although there were problems with the pressure data (as discussed in section 3.4), it was possible to gain information about the combustion process that it was not possible to infer from any of the performance data. For example, coefficients of variation for the magnitude and position of maximum cylinder pressure allowed cyclic variation for the engine to be quantified (Figs. 3.35 and 3.37). The COV of maximum cylinder pressure ( P M A X ) almost doubled as the air-fuel mixture was leaned out, as shown in Fig. 3.35.  PMAX  was not stable at stoichiometric fuelling - it had a COV of 8%. The  COV of crank angle of  PMAX,  however, was very low, (about 0.5%), indicating good  stability for this parameter for all air-fuel ratios, (Fig. 3.37). It was also interesting to consider where the maximum cylinder pressure occurred, as illustrated by Fig. 3.36. As the mixture was leaned out,  PMAX  moved closer to TDC, but in general it was near the  optimum of between 10-15 degrees after TDC. The shift of P M A X closer to TDC with increasing X was due mainly to the advance of MBT spark timing, (refer to Table 3.1 and Fig. 3.3). The slow flame speed of natural gas in lean mixtures is probably the reason for this advance. Heat release over a longer duration meant that best torque occurred when PMAX  was closer to TDC. The implementation of PSC moved P M A X closer again to TDC  than the homogeneous case - but only by about one degree. It may be possible to further retard the PSC spark timing without loss of torque. This would result in a slight lowering of peak cylinder pressure and temperature, and consequently would further reduce the production of NOx-  Heat release data were used to determine ignition delay and combustion duration. The very low value for COV . . c  a  Prnax  (Fig. 3.37) gives confidence that the influence of thermal  shock was consistent and repeatable for measurements with respect to crank angle. Therefore it was assumed that it would not affect a comparative evaluation of either combustion duration or ignition delay. An example of the averaged in-cylinder pressure data at 1=1.65, 2500rpm - from which the heat release data was calculated - is given in Fig. 3.38. Heat release rate increased with PSC (Fig 3.39), explaining why the position of  PMAX  shifted closer to TDC when PSC was implemented. Comparative values for  ignition delay and combustion duration were determined using plots of mass fraction burned (MFB) vs. crank angle, (Fig. 3.40). The 5%-95% MFB combustion duration (Fig.  33  3.41) increased steadily from stoichiometric combustion (30 degrees duration) to the LML (45 degrees duration). PSC fuelling reduced the burn duration beyond 1=1.5 by about 4 degrees. Ignition delay is expressed in terms of the duration between the spark and 5% of mass fraction burned. 5% MFB was chosen because it is difficult to accurately identify where combustion starts, (refer to heat release rate curves, Fig. 3.39), and this was considered to be a more reliable method for comparative purposes. The comparative plot of ignition delay showed a reasonably consistent delay until 1=1.4 (Fig. 3.42). Thereafter the delay increased rapidly, rising to a maximum duration for homogeneous fuelling that was almost 20 degrees longer than at stoichiometric.  Once again, PSC results in an  improvement to this parameter: at 1=1.65 the duration is reduced by 7 degrees. This is a strong indication that well-timed PSC injection improves the quality of combustion initiation. 3.8  Summary  Partially Stratified-Charge technology was found to improve engine performance at 2000rpm and 2500rpm when compared with simple homogeneous fuelling at lean air-fuel ratios (in general greater than 1=1.5). At 1500rpm no improvement in performance with PSC was observed, and this is thought to be due to un-optimised injection settings, in particular end-of-injection timing with respect to the spark. At 3 OOOrpm it was not possible to inject the desired quantities of fuel through the spark plug, and hence no conclusions could be drawn regarding performance improvement with PSC. However a small reduction in CO and tHC emissions close to the lean misfire limit suggested a positive effect of PSC.  Analysis of pressure data at 2500rpm revealed that PSC  technology increased peak in-cylinder pressure, reduced combustion duration and ignition delay times and in general increased combustion stability at lean air-fuel ratios for this speed. The overall combustion duration, as measuredfromspark to 95% MFB, was reduced by up to 15% when using PSC.  34  300 •Homogeneous charge - no injection -Inj. Rate: 7 g/h (0.6% of fuel at RAFR=1.6) •Inj. Rate: 14 g/h (1.1% of fuel at RAFR=1.6)  280  260  240  220 1.5  1.6  1.7  Relative Air-Fuel Ratio  Figure 3.1 BSFC with Changing PSC Flow-Rate (EOI at MBT-10), 2500rpm  300 -Homogeneous fuelling - no injection •EOI at MBT-5 (1.1% of fuel at RAFR=1.6) •EOI at MBT-10 (1.2% of fuel at RAFR=1.6)  280  •EOI at MBT-15 (1.2% of fuel at RAFR=1.6)  260  240  220 1.3  1.4  1.5  1.6  1.7  Relative Air-Fuel Ratio  Figure 3.2 BSFC with Changing PSC Timing, 2500rpm (Injection flow-rate = 15 g  60  55  u o  H aa  u WI  • Homogeneous fuelling - no injection  "  A P S C fuelling  50  m  A  •  V V  45  M S  1  40  « a.  35  5»  30  • H  • II  1.0  1.2  1.4  1  1.6  Relative Air-Fuel Ratio  Figure 3.3 Spark timing ( M B T ) Retard due to P S C , 2500rpm  35% -Homogeneous charge - no injection -Inj. Rate: 7 g/h (0.6% o f fuel at RAFR=1.6)  33%  u c  •Inj. Rate: 14 g/h (1.1% o f fuel at R A F R = 1 . 6 )  31%  '5  S  W  E I.  29%  II  27%  25% 1.3  1.4  1.5  1.6  1.7  Relative Air-Fuel Ratio  Figure 3.4 Thermal Efficiency with Changing P S C Flow-Rate, 2500rpm  1.8  1.3  1.4  1.5  1.6  1.7  Relative Air-Fuel Ratio  Figure 3.5 Thermal Efficiency with Changing PSC Timing, 2500rpm  re  1.0  1.2  1.4  1.6  Relative Air-Fuel Ratio  Figure 3.6 Power Control Using Lean Combustion, 2500rpm  1.8  7  6  4-1  ,  ,  ,  ,  1  1.3  1.4  1.5  1.6  1.7  1.8  Relative Air-Fuel Ratio  Figure 3.7 BMEP with Changing PSC Flow-Rate, 2500rpm  -Homogeneous fuelling - no injection - EOI at MBT-5 (1.1 % of fuel at RAFR= 1.6) -EOI at MBT-10 (1.2% of fuel at RAFR=1.6) -EOI at MBT-15 (1.2% of fuel at RAFR=1.6) 1.3  1.4  1.5  1.6  1.7  Relative Air-Fuel Ratio  Figure 3.8 BMEP with Changing PSC Timing, 2500rpm  38  10.0  B Homogeneous fuelling - no injection A PSC fuelling 1.  es  Xi  PH  W  1.2  1.4  1.6  Relative Air-Fuel Ratio  Figure 3.9 Gross IMEP from In-cylinder Pressure Data, 2500rpm  70  Figure 3.10 Sample P-V Diagram at A=1.10, Homogeneous Fuelling, 2500rpm  39  1000.00  100.00 CS  10.00 u  a.  1.00  a U  0.0|)00  1  U.UI JU1  0.0 pio  0.10  Evidence of Thermal Shock 0.01  Crank Angle (deg)  Figure 3.11 Sample logP-logV Diagram at 1=1.10, Homogeneous Fuelling, 2500rpm  12  10  > • Homogeneous fuelling - no injection A PSC fuelling  1.0  1.2  1.4  1.6  1.8  Relative Air-Fuel Ratio  Figure 3.12 Coefficient of Variation of Indicated Mean Effective Pressure, 2500rpm  40  10 -Homogeneous charge - no injection -Inj. Rate: 7 g/h (0.6% of fuel at RAFR=1.6) -Inj. Rate: 14 g/h (1.1% of fuel at RAFR=1.6)  "Sfc i  o u  1.3  1.4  1.5  1.6  1.7  1.8  Relative Air-Fuel Ratio  Figure 3.13 Carbon Monoxide Emissions with changing PSC Flow-Rate, 2500rpm  10 -Homogeneous fuelling - no injection -EOI at MBT-5 (1.1% of fuel at RAFR=1.6) - EOI at MBT-10(1.2% of fuel at RAFR= 1.6) -EOI at MBT-15 (1.2% of fuel at RAFR=1.6)  | •  O u  1.5  1.6  Relative Air-Fuel Ratio  Figure 3.14 Carbon Monoxide Emissions with changing PSC Timing, 2500rpm  41  40  -Homogeneous charge - no injection -Inj. Rate: 7 g/h (0.6% of fuel at RAFR=1.6) -Inj. Rate: 14 g/h (1.1% of fuel at RAFR=1.6)  30  I cu I  u  s  1.3  1.4  1.5  1.6  1.7  1.8  Relative Air-Fuel Ratio  Figure 3.15 Total Hydrocarbons Emissions with Changing PSC Flow-Rate, 2500rpm  40  Homogeneous fuelling - no injection EOI at MBT-5 (1.1% of fuel at RAFR=1.6) 30 XI  EOI at MBT-10 (1.2% of fuel at RAFR=1.6) EOI at MBT-15 (1.2% of fuel at RAFR=1.6)  I  "5k -4-*  cu I  u pq  1.5  1.6  Relative Air-Fuel Ratio  Figure 3.16 Total Hydrocarbons Emissions with Changing PSC Timing, 2500rpm  42  40  •Homogeneous fuelling - tHC  30  •Homogeneous fuelling - CH4  20  1.0  1.4  1.2  1.6  Relative Air-Fuel Ratio  Figure 3.17 Comparison between tHC and CH Emissions, 2500rpm 4  -Homogeneous fuelling - no injection -EOI at MBT-5 (1.1% of fuel at RAFR=1.6) -EOI at MBT-10 (1.2% of fuel at RAFR=1.6) - EOI at MBT-15(1.2% of fuel at RAFR=1.6)  1.3  1.4  1.5  1.6  1.7  1.8  Relative Air-Fuel Ratio  Figure 3.18 Oxides of Nitrogen with Changing PSC Flow-Rate (Siemens Analyser), 2500rpm  43  30  c  Homogeneous fuelling - no injection  cu  Es u  EOI at MBT-15 (1.2% of fuel at RAFR=1.6)  CA  a  scu E CA  V  55 I  /•~\ XI I  § fr •  XI  o 1.0  1.2  1.4  1.6  Relative Air-Fuel Ratio  Figure 3.19 Relative Magnitude of NO Emissions (Siemens Analyser), 2500rpm x  60  55  II  •  u  cs XI v i.  s CA  •  •  • A  50  •  CA  k. OH cu  u  45  U cs  40  2  A  A  •  A  • • Homogeneous fuelling - no injection  35  A PSC fuelling 30 1.0  1.2  1.4  1.6  1  Relative Air-Fuel Ratio  Figure 3.20 Maximum In-cylinder Pressure for Homogeneous and PSC Fuelling, 2500rpm 44  Figure 3.21 BSFC with Changing PSC Flow-Rate, 2000rpm  35%  •Homogeneous charge - no injection •Inj. Rate: 21 g/h (2.1% of fuel at RAFR=1.6) 33%  •Inj. Rate: 28 g/h (2.6% of fuel at RAFR=1.6) -Inj. Rate: 38 g/h (3.7% of fuel at RAFR=1.6)  a  .2  31%,  £ u  13  29%  E  27%  25% 1.5  1.6  Relative Air-Fuel Ratio  Figure 3.22 Thermal Efficiency with Changing PSC Flow-Rate, 2000rpm  •Homogeneous charge - no injection •Inj. Rate: 21 g/h (2.1% of fuel at RAFR=1.6) •Inj. Rate: 28 g/h (2.6% of fuel at RAFR=1.6) •Inj. Rate: 38 g/h (3.7% of fuel at RAFR=1.6) 4 4 1.3  1.4  1.5  1.6  1.7  Relative Air-Fuel Ratio  Figure 3.23 BMEP with Changing PSC Flow-Rate, 2000rpm  1.5  1.6  Relative Air-Fuel Ratio  Figure 3.24 CO with Changing PSC Flow-Rate, 2000rpm  40 -Homogeneous charge - no injection -Inj. Rate: 21 g/h (2.1% of fuel at RAFR=1.6) -Inj. Rate: 28 g/h (2.6% of fuel at RAFR=1.6) -Inj. Rate: 38 g/h (3.7% of fuel at RAFR=1.6)  1.5  1.6  1.7  Relative Air-Fuel Ratio  Figure 3.25 tHC with Changing PSC Flow-Rate, 2000rpm  16  -Homogeneous charge - no injection 12  -Inj. Rate: 21 g/h (2.1% of fuel at RAFR=1.6)  1.3  1.4  1.5  1.6  1.7  Relative Air-Fuel Ratio  Figure 3.26 NO with Changing PSC Flow-Rate (API Analyser), 2000rpm x  300  280  260  240  220 1.5  1.6  Relative Air-Fuel Ratio  Figure 3.27 BSFC with Changing PSC Flow-Rate, 1500rpm  T3 i  o u  1.3  1.4  1.5  1.6  1.7  Relative Air-Fuel Ratio  Figure 3.28 CO with Changing PSC Flow-Rate, 1500rpm  30  1.3  1.4  1.5  1.6  1.7  Relative Air-Fuel Ratio  Figure 3.29 tHC with Changing PSC Flow-Rate, 1500rpm  16  -Homogeneous charge - no injection -Inj. Rate: 11 g/h (1.5% of fuel at RAFR=1.6) 12  -Inj. Rate: 17 g/h (2.2% of fuel at RAFR=1.6) -Inj. Rate: 30 g/hr (4.0%of Stoich Flow)  fr •a • x O  z  1.5  1.6  Relative Air-Fuel Ratio  Figure 3.30 NO with Changing PSC Flow-Rate (API Analyser), 1500rpm x  300  -•-Homogeneous charge - no injection 280  -•-Inj. Rate: 3 g/h (0.2% of fuel at RAFR=1.6)  220 -) 1.3  1.4  1.5  1.6  1.7  Relative Air-Fuel Ratio  Figure 3.31 BSFC with Highest Possible PSC Flow-Rate, 3000rpm  •Homogeneous charge - no injection •Inj. Rate: 3 g/h (0.2% of fuel at RAFR=1.6)  JS I  1 6D  8  4  1.3  1.4  1.5  1.6  1.7  Relative Air-Fuel Ratio  Figure 3.32 CO with Highest Possible PSC Flow-Rate, 3000rpm  50  30  -Homogeneous charge - no injection -Inj. Rate: 3 g/h (0.2% of fuel at RAFR=1.6)  20  u a I  1.3  1.4  1.5  1.6  1.7  Relative Air-Fuel Ratio  Figure 3.33 tHC with Highest Possible PSC Flow-Rate, 3000rpm  -Homogeneous charge - no injection -Inj. Rate: 3 g/h (0.2% of fuel at RAFR=1.6)  BI)  •  O Z  1.3  1.4  1.5  1.6  1.7  Relative Air-Fuel Ratio  Figure 3.34 NO with Highest Possible PSC Flow-Rate (API Analyser), 3000rpm x  51  20  c>uS cuu u  0.  16  A  1  A  •  12  A  cu  •O  u  •  x es  o > o U  • Homogeneous fuelling - no injection A PSC fuelling  1.0  1.2  1.4  1.6  Relative Air-Fuel Ratio  Figure 3.35 Coefficient of Variation of Maximum Cylinder Pressure, 2500rpm  20  • Homogeneous fuelling - no injection A PSC fuelling w  15  « E  P* C M  O — c  CS  •J5  10  A  •  • A  1.0  1.2  1.4  A  1.6  Relative Air-Fuel Ratio  Figure 3.36 Crank Angle at which Maximum Cylinder Pressure Occurs, 2500rpm  52  2.0  B Homogeneous fuelling - no injection A PSC fuelling 1.5  E PN  6D  1.0  a  es  s  es o > o  0.5  U  0.0 1.0  1.2  1.4  1.6  Relative Air-Fuel Ratio  Figure 3.37 Coefficient of Variation of Crank Angle at Which P  max  Occurs, 2500rpm  6©-  — Homogeneous fuelling, RAFR= 1.65 — P S C fuelling, RAFR = 1.65  50  CcJ XI  s  CA CA CU L.  D-  cu •a e  IT  1@ 10  1  -60  -40  ,  0-  -20  20  40  60  Crank Angle Degrees  Figure 3.38 In-cylinder Pressure Data; comparison at 1=1.65, 2500rpm  53  5 E  3  Crank Angle Degrees  Figure 3.39 Heat Release Rate vs. Crank Angle; comparison at 1=1.65, 2500rpm  -400%  - Homogeneous fuelling, RAFR= 1.65 (MBT Spark = -54) V  e u s PQ c  -PSC fuelling, RAFR= 1.65 (MBT Spark = -50)  o  Crank Angle Degrees  Figure 3.40 Mass Fraction Burned vs. Crank Angle; comparison at 1=1.65, 2500rpm  54  50.0  ^  45.0  M s> w  a  .2 s  40.0  B  35.0  a  s CO  An  H'  30.0  in  • Homogeneous fuelling - no injection  25.0  A PSC fuelling 20.0 1.0  1.2  1.4  1.6  Relative Air-Fuel Ratio  Figure 3.41 Combustion Duration (5-95% of Mass Burned Fraction), 2500rpm  45.0 -r  OD  13  • Homogeneous fuelling - no injection  40.0  A PSC fuelling ^  35.0  A  i. es  S«  30.0  cs •o  a  25.0  a 20.0 1.0  1.2  1.4  1.6  Relative Air-Fuel Ratio  Figure 3.42 Ignition Delay vs. Relative Air-Fuel Ratio, 2500rpm  55  Chapter 4 CONCLUSIONS  4.1  Conclusions  The Partially Stratified-Charge (PSC) concept is a method for improving the combustion initiation and stability of an otherwise standard spark-ignition engine, lean-fuelled with natural gas. A small quantity of natural gas, comprising less than 5% of the overall fuel mass, is injected in the region of the spark plug just prior to ignition, while in the main combustion chamber area an ultra-lean homogeneous charge is maintained. In this way a relatively rich air-fuel mixture is created in the vicinity of the spark plug. When the spark fires, a stable flame kernel is initiated which then rapidly propagates through the remainder of the lean charge. The purpose of this research was to investigate whether an engine running with PSC would exhibit better performance, namely increased efficiency and reduced emissions, at lean relative air-fuel ratios. The effect of changing PSC injection timing and quantity was investigated, and the performance characteristics of PSC were compared to the general homogeneous fuelling case. Partially Stratified-Charge technology was found to significantly improve engine performance at 2000rpm and 2500rpm when compared with simple homogeneous fuelling at lean air-fuel ratios (in general greater than A=1.5). At 1500rpm there was no improvement in performance with the chosen PSC injection settings. At 3OOOrpm it was not possible to inject the desired quantities of fuel through the spark plug, and no efficiency or power improvements were observed. However a small reduction in CO and tHC emissions close to the lean misfire limit suggested a positive effect of PSC.  Analysis of pressure data at 2500rpm revealed that PSC  technology increased peak in-cylinder pressure, reduced combustion duration and ignition delay times and in general increased combustion stability at lean air-fuel ratios.  56  With consideration of the information presented in this thesis, the following key conclusions have been reached: 1.  With the current Partially Stratified-Charge system, the lean misfire limit for the Ricardo Hydra engine was extended by approximately 5% at speeds of 2000rpm and 2500rpm. At all speeds tested, application of PSC at lean air-fuel ratios resulted in retardation of MBT spark timing.  2.  Use of PSC at 2000rpm and 2500rpm resulted in a reduction of approximately 8% in brake specific fuel consumption at a relative air-fuel ratio of 1.65, and there was a corresponding increase in thermal efficiency. Engine power was increased by up to 7%.  3.  At these speeds significant reductions in carbon monoxide and total hydrocarbon emissions were observed.  Oxides of nitrogen increased slightly due to the  increase in maximum cylinder pressure with PSC, but were found to be very low near the LML when compared with stoichiometric fuelling. 4.  A general improvement in combustion quality and stability occurred with PSC at lean air-fuel ratios at 2500rpm. The coefficient of variation of indicated mean effective pressure was reduced, maximum cylinder pressure increased and the overall combustion duration, from spark to 95% MFB, was reduced by up to 15% due to the implementation of PSC.  4.2  Recommendations for Future Work  1.  It would be interesting to quantify the PSC injection timing in real time, and further investigate variation of end-of-injection timing.  Also, with some  relatively straightforward improvements to the control system, it would be possible to examine a full range of injection flow-rates such that the optimum values of flow-rate could be determined. On a larger scale, it will be necessary to  57  determine optimum PSC injection flow and EOI timing for each air-fuel ratio at each speed. The next stage would be to develop an engine management control system for the PSC engine that could deliver the required power to a user while maximising thermal efficiency and minimising exhaust emissions. Work should continue on the practical aspects of PSC so that eventually an engine can be installed invehicle to demonstrate - and further investigate - the potential of this technology.  58  REFERENCES  Arcoumanis, C , Hull, D. R. and Whitelaw, J. H., 1994, "An Approach to Charge Stratification in Lean-Burn, Spark-Ignition Engines", SAE Paper No. 941878. Arcoumanis, C , Hull, D. R. and Whitelaw, J. H., 1997, "Optimizing Local Charge Stratification in a Lean-burn Spark Ignition Engine", Proc. Inst. Mech. Engrs., Vol. 211, PartD, pp. 145-154. Das, A. and Watson, H. C , 1997, "Development of a Natural Gas Spark Ignition Engine for Optimum Performance", Proc. Inst. Mech. Engrs., Vol. 211, Part D, pp. 361-378. Dinh, H. T., 1994, "Operating, Performance and Emission Characteristics of Compressed Natural Gas Vehicles", ICE-Vol. 21, Natural Gas and Alternative Fuels for Engines, ASME. Evans, R. L., 1999a, "Lean-Burn Natural Gas Engines for High Efficiency and Low Emissions", ICE-Vol. 32-2, 1999 Spring Technical Conference, ASME. Evans, R. L., 1999b, "A Novel Stratified-Charge Engine Design for Low Emissions and High Efficiency", Combustion Inst. Canadian Sect., 1999 Spring Technical Meeting, Paper No. 13. Evans, R. L., and Blaszczyk, J., 1997, "Fast-Burn Combustion Chamber Design for Natural Gas Engines", ICE-Vol. 120, Transactions of the ASME. Fleming, R. D. and O'Neal, G. B., 1985, "Potential for Improving the Efficiency of a Spark Ignition Engine for Natural Gas Fuel", SAE Paper No. 852073. Gambino, M., Corbo, P., Iannaccone, S., Unich, A. and Bata, R., 1993, "High Turbulence Combustion Chamber for Lean-Burn Heavy-Duty CNG Engines", ICE-Vol. 20, Alternate Fuels, Engine Performance and Emissions, ASME. Graboski, M. S., McCormick, R. L., Newlin, A. W., Dunnuck, D. L., Kamel, M. M. and Ingle, W. D., 1997, "Effect of Fuel Composition and Altitude on Regulated Emissions from a Lean-Burn, Closed Loop Controlled Natural Gas Engine", SAE Paper No. 971707. Green, R. K. and Zavier, C. C , 1992, "Charge Stratification in a Spark Ignition Engine", Proc Instn Mech Engrs, Vol 206, Part A, pp 59-64. Gupta, M. and Bell, S. R., 1994, "An Investigation of Lean Combustion in a Natural Gas Fuelled Spark Ignited Engine", ICE-Vol. 21, Natural Gas and Alternative Fuels for Engines, ASME.  59  Heywood, J. B., 1988, Internal Combustion Engine Fundamentals, McGraw-Hill, Inc.,  New York. Holman, J. P., 2000, Experimental Methods For Engineers, 7 ed., McGraw-Hill, Inc., New York. th  Ingersoll, J. G., 1996, Natural gas Vehicles, The Fairmont Press, Inc., GA. Kapus, P. E. and Chmela, F. G., 1994, "The new AVL Gas Engine Combustion System", ICE-Vol. 20, Alternate Fuels, Engine Performance and Emissions, ASME. Kingston Jones, M. G. and Heaton, D. M., 1989, "Nebula System for Lean Burn Spark Ignited Gas Engines", SAE Paper No. 890211. Larsen, J. F. and Wallace, J. S., 1994, "Comparison of Emissions and Efficiency of a Turbocharged Lean-Burn Natural gas and Hythane Fuelled Engine", Natural Gas and Alternative Fuels for Engines, ICE, Vol. 24, pp. 31-40. Lumsden, G. and Watson, H. C , 1995, "Optimum Control of an S.I. Engine with a A=5 Capability", SAE Paper No. 950689: Nichols, R. J., 1993, "The Challenge of Change in the Auto Industry: Why Alternative Fuels?", ICE-Vol. 20, Alternate Fuels, Engine Performance and Emissions, ASME. Patrick, R. S., Eaton, A. R. and Powell, J. D., 1991, "A study of Electronic Engine Control System Structures for Lean-Burn, Natural Gas Engine Operation in a Spark Ignition Engine", ICE-Vol. 15, Fuels, Controls, and Aftertreatment for Low Emissions Engines, ASME. Randolf, A. L.,1994, "Cylinder-Pressure-Based Combustion Analysis in Race Engines", SAE Paper No. 942487. Raine, R. R., Zhang, G. and Pflug, A., 1997, "Comparison of Emissions from Natural Gas and Gasoline Fuelled Engines - Total Hydrocarbon and Methane Emissions and Exhaust Recirculation Effects", SAE Paper No. 970743. SAE Recommended Practice, 1997, "Stoichiometric Air-Fuel Ratios of Automotive Fuels", SAE Jl829 DEC97. SAE Recommended Practice, 1995, "Recommended Practice for Compressed Natural Gas Vehicle Fuel", SAE J1616 FEB94. SAE Recommended Practice, 1995, "Measurement of Carbon Dioxide, Carbon Monoxide and Oxides of Nitrogen in Diesel Exhaust", SAE J177 JUN95.  60  SAE Recommended Practice, 1992, "Measurement of Intake Air or Exhaust Gas Flow of Diesel Engines", SAE J244 AUG92. SAE Standard, 1995, "Engine Power Test Code - Spark Ignition and Compression Ignition - Gross Power Rating", SAE J1995 JUN95. SAE Standard, 1995, "Engine Power Test Code - Spark Ignition and Compression Ignition - Net Power Rating", SAE J1349 JUN95. Stone, C. R. and Ladommatos, N., 1991, "Design and Evaluation of a Fast-Burn SparkIgnition Combustion System for Gaseous Fuels at High Compression Ratios", J. Inst. Energy, Vol. 64, pp. 202-211. Strahle, W. C , 1993, An Introduction to Combustion, Gordon and Breach Science Publishers S.A. Takagaki, S. S. and Raine, R. R., 1997, "The Effects of Compression ratio on Nitric Oxide and Hydrocarbon Emissions from a Spark-Ignition Natural Gas Fuelled Engine", SAE Paper No. 970506. Unich, A., Bata, R. M. and Lyons, D. W., 1993, "Natural gas: A Promising Fuel for I.C. Engines", SAE Paper No. 930929. Weaver, C. S., 1989, "Natural Gas Vehicles - A Review of the State of the Art", SAE Paper No. 892133.  61  Appendix A ENGINE OPERATING PROCEDURE  The following is the start-up procedure for the Ricardo Hydro single-cylinder research engine in the UBC Alternative Fuels Laboratory. In the engine test cell: 1. Check engine oil and coolant level, and check around the engine for leaks. 2. Crank engine by hand once or twice (to ensure there has been no leak into the combustion chamber). 3. Check that there is no condensation in the exhaust (briefly open valve). 4. Turn on both main circuit breakers. 5. Open the two green taps on the low pressure NG line. 6. Turn the cooling water on at the mains tap (several turns). 7. Reset emergency stop button on engine post. 8. Turn ON switch on ignition box. 9. If running with PSC pressurise the injection CNG line to 400psi (27bar). In the control room - preliminaries: 1. Check main control panel is plugged in (should be left on in general). 2. Reset emergency stop button on control panel. 3. Turn on oil and water pumps and heaters. 4. Throttle: switch to "run" (not "idle"), and dial to 100%. 5. Ricardo Ignition: switch always to "off. 6. Turn on multi-spark timing box (also displays accurate RPM). Ignition here off. 7. Check gas flow switch is off, second switch to "flow", dial to zero. 8. Open tap on gas line. 9. Dynamometer: check control switch to "auto". 62  10. Wait for oil to preheat to at least 60deg Celsius, (note: may need to turn heaters off when engine is firing). Do not motor engine cold! 11. Make note of: •  engine hours  •  relative humidity  •  barometric pressure  Engine start-up: 1. Turn on data acquisition system (see Appendix B). 2. Set speed dial to 3.0. 3. Press "reset", and immediately press green start button. 4. Engine should increase to speed of approx. 1500rpm. 5. Turn on ignition using Multi-spark Timing Box, and adjust to 20 deg. BTDC (160 deg. ABDC). Note: should be set to single spark, dwell time 70us, spark time lOOus. 6. Turn switch onflow-meterto "on". 7. Dial upflowuntil steady firing is achieved with lambda = 1.0. Engine is running with stoichiometric air-fuel ratio. 8. After engine has been firing for one minute, turn on Exhaust Lambda Sensor (AFRecorder) - do not leave off while engine is running. Shutting down the engine: 1. Turn off the gas supply at the flow-meter. 2. Cut ignition. 3. Allow engine to motor briefly, until exhaust temperature drops below 100°C. 4. Press red stop button. Engine will come to a complete stop. 5. After about two minutes, disable AFRecorder. 6. Close all gas taps on low and high pressure CNG lines. 7. Turn off switch on ignition box. 8. Depress both emergency stop buttons.  63  9. When engine has cooled (after about fifteen minutes): •  Turn off oil and water pumps.  •  Turn off the two main circuit breakers.  •  Turn off cooling water tap.  WARNING! Do not hit "STOP" button while the engine is firing, or the load will be disabled and the engine speed will increase. If you must perform an emergency shut-down, cut the fuel and ignition at the control panel before stopping the dynamometer.  64  Appendix B EMISSIONS BENCH: OPERATING AND CALIBRATION PROCEDURE  NOTES: 1. Before starting to calibrate, make sure system temperature on oven is approximately 190 C (or 125 C). 2. Do not turn off emissions bench unless a long period without testing is anticipated. 3. Turn off Ratfisch, (Heater Oven, Pump, then Power), when not using - otherwise excessive fuel (H2) and cylinder air will be used. 4. Check that the correct flow rate is being used before each calibration step. 5. As filters get clogged, it may be necessary to increase the gas cylinder pressures to maintain correct flow rates. Procedure - General 1. Open calibration gas cylinders - compressed air, H2, CH4, CO, C O 2 , NO, N2. 2. Select a heated sample stream - turn the selected engine (Ricardo) dial to ON position; (other switches should be OFF). 3. Set valves for C O 2 selection to appropriate engine. 4. Set enclosure temperature on Ogden dial: 190 C for Diesel, 125 C for Spark Ignition Engines. 5. To power the Emissions Bench, turn ON all the breakers, (except the two engines which are not selected and the NOx converter), inside the breaker panel. 6. Select engine, (e.g. Ricardo), by sliding all 6 switches in Cabinet #2 to correct position. 7. Start the heated pump on the Ratfisch RS-55 (heated tHC analyser) by pressing PUMP and HEATER-OVEN buttons ON. (Wait about ten minutes for Ratfisch temp, to reach 190 C). 65  8. Start up Ratfisch by holding "H2 over" button and adjusting Fuel to 0.5 and Air to 0.8, then hold "H2 over" button and press "ignition", and hold until ignition light goes off. 9. If ignition light does not go off, purge H fuel line (remove/replace fuel hose inside 2  cabinet #1 on the back of Ratfisch.) and then repeat step 8. Calibrate CO Analyser (Siemens Ultramat 21P) Range: 0-5V for 0-10,000ppm 1. Turn the CO switch to ZERO. 2. Press ">0<" button to zero the gas and again when zeroing is complete. (Note: during zeroing, flow should read 2.5 1/min. If it doesn't, correct it by adjusting the flow controller inside the cabinet). 3. Turn CO switch to SPAN. Check flow rate is 2.5 1/min, (adjust using "Flow to analysers" knob if necessary). 4. Adjust the CO potentiometer if necessary so that display reads 2077ppm, (i.e. 0.201%). 5. Turn CO switch to RUN. Adjust flow rate to 2.5 1/min again if necessary. Calibrate tHC Analyser (Ratfisch RS-55) Range: 0-10V for 0-10,000ppm 1. Check that Ratfisch temperature is at 190 C before calibrating. 2. Check that ignition light is OFF. If not, hold "H2 over" button and press "ignition", and hold until ignition light goes off. 3. Turn large black knob to "ZeroGas" position. Set sample back-pressure at 200 mbar, maintain at every step. 4. Adjust "Zero" on Gossen display using potentiometer. 5. Turn large back knob to "CalGas" and turn fuel knob back until Gossen display starts decreasing, then return to the maximum value. 6. Adjust "Gain" pot to obtain a reading of 3900 on the Gossen display. 7. Turn large back knob to "Sample" and reset back-pressure to 200 mbar.  66  Calibrate N O Analyser (API) x  Range: 1-5 V for 0-3000ppm 1. Plug in vacuum pump (behind cabinet #2). 2. Turn NO switch to ZERO. 3. Press "calibration", and wait 10 minutes (for readings to stabilize). 4. Press "zero". Press "enter". 5. Turn NO switch to SPAN. 6. Wait 10 minutes (or for readings to stabilize). 7. Press "span". 8. Display should read 1957 ppm NOx9. Press "enter". Press "exit". 10. "Calibration" (yellow) light should go out and "Sample" (green) light should go on. Calibrate CO2 Analyser (Beckman 880)  Range: 0-5V for 0-20% 1. Turn the C O 2 switch to zero (top of cab #2). 2. Check flow rate is 1.5 1/min, (using "Flow to analysers" knob). 3. Press "Zero" then "Enter". Adjust with arrows to read 0% on left of display, and press "Enter" again. 4. Turn C0 switch to SPAN. Check flow rate is 1.5 1/min. 2  5. Press "Span", "Enter". Adjust with arrows to read between 9-18.5%) (around 16% is good) on left of display, then press "Enter". 6. Turn C O 2 switch to RUN. Adjust flow rate to 1.5 1/min again if necessary.  67  Calibrate O2 Analyser (Oxymat 5E) Range: 0-5V for 0-21% 1. Turn the O2 switch to ZERO. (Leave time for, analyser to "settle" at each calibration step). 2. Check flow rate is 0.7 1/min. 3. Enter ".111" to make Code 1 light go out. 4. Set analyser to "Calibration" mode by pressing "Meas/Cal" button. 5. Press "5", then press "Enter" to zero the analyser. Wait until "not ready" light is off. 6. Turn the 0 switch to SPAN. (Check flow rate is 0.7 1/min). 2  7. Press "8", then press "Enter" to span, (wait until "not ready" light is off). 8. Press "Meas/Cal" button to set analyser to Measure mode. 9. Turn O2 switch to RUN. Adjust flow rate to 0.7 1/min again if necessary.  68  Appendix C DATA ACQUISITION SYSTEM OPERATING PROCEDURE  The following is the procedure for using the Data Acquisition System for the Ricardo Hydro single-cylinder test engine. In the engine test cell:  (Note: the following steps are unnecessary if the Ricardo is being tested on a day-to-day basis). 1. Check power bar(s) are on. 2. Turn on AFRecorder (N.B. Do not enable sensor until engine has been running for approximately one minute or it may be damaged). 3. Turn on charge amplifier for the in-cylinder pressure transducer. 4. Switch on the on CPU ("Dynoserver" will start up automatically). 5. After a few seconds (CPU will "beep"), turn on the Data Acquisition System. In the control room:  1. Turn on the second CPU and monitor - must be a recognised user to log on. 2. Open "Dynoclient" executable (icon on desktop). 3. User is prompted to save to a file in drive F:/ - may create a new text file (.txt) or append to a previous one. Example filename: 25mar01_crl 1.9_2000rpm_homog.txt 4. Wait for CPU to make connection to Data Acquisition System (approx. 15 sees). 5. Note: can set "Relative Humidity" and "Barometric Pressure" values now, (from gauges in lab). 6. Open "Pressureclient" (icon on desktop).  69  Saving Test Data:  1. Run engine at required test conditions. 2. Leave engine for 3-5 minutes to stabilise and readjust to the required running conditions if necessary. 3. Set spark timing to MBT and enter this value on "Dynoclient" screen. Ensure "Test Number" is incremented to correct value. (Note: it is very useful to keep a hard copy of the testing for reference when processing data - see sample "test sheet" below). 4. Click "LOG DATA to File" button, wait for required test period and then click again to stop logging. Two minutes of sampling acquires approximately 170 test points. (Note: In order to take a "snapshot" of the data you are currently sampling, you may open Ricardo_Emissions_RunTime.xls (may be found in "Data Processing" folder) and then click "Send Data to EXCEL" button in the Dynoclient window. This data and associated calculations may be viewed in the open spreadsheet). 5. Data is stored as a textfile(.txt). Excel recognises it as "comma delimited"filetype. If you log more data to the samefilename, it is appended to previously saved data. (Note: you must stop "Dynoclient" in order to close thefileyou are writing to. When you restart you can chose a newfilenameor append data to a previous one). 6. To save cylinder pressure data click the "Log Data" button in the "Pressureclient" window. User is prompted to save to afilein drive F:/, and then the high speed Data Acquisition System acquires pressure data for 100 consecutive cycles. Examplefilename:01marl6_crl 1.9_2000rpm_homog_pressL14.txt 7. Set new engine test point and repeat above data acquisition. 8. Note: must stop and close both Dynoclient and Pressureclient before opening data files.  70  Shutting Down:  1. In order to prevent damage to the AFRecorder sensor, it should be disabled about 2 minutes after engine is shut down. 2. Stop "Dynoclient" and close window. 3. Stop "Pressureclient" and close window. 4. Exit Lab view. (Note: the following steps are unnecessary if the Ricardo is being tested on a day-to-day basis). 5. Using "pcAnywhere" (icon on desktop) stop and close the Dynoserver on the DASY CPU, then exit Labview and shut down the CPU remotely. 6. Turn off engine room CPU and Data Acquisition System. 7. Turn off AFRecorder and charge amp. 8. Shut down the control room computer. 9. Turn off control-room computer and monitor.  71  RICARDO TEST SHEET  Test Description:  Engine Hours:  Test Engineer: Date:  Barometric Pressure (mm Hg):  R e l a t i v e H u m i d i t y (%):  S p a r k T i m i n g M B T (y/n): E n g i n e S p e e d (rpm):  N o . of S p a r k s :  Throttle P o s i t i o n (%):  S p a r k D w e l l t i m e (us): S p a r k D i s c h a r g e T i m e (us):  Performance Data File N a m e : P r e s s u r e D a t a File N a m e ( s ) :  Test Exhaust Spark Timing No. Lambda (deg BTDC) From Control Box  Throttle Engine Position Speed (%) (rpm)  Pressure Perf. Test Data (y/n) Start Time  Perf. Test End Time  0 1 2 3 4 5 6 7 8 9 10  Additional  Comments:  Figure C l Sample Test Sheet 72  Appendix D DATA PROCESSING INSTRUCTIONS  The following is the procedure for processing data acquired from the Ricardo Hydro single-cylinder test engine. Performance data: 1. Open Microsoft Excel. Within Excel, open the performancefile(*.txt format) you wish to process. Excel will recognise this as a "delimited" type file. Click "next". Choose "comma" as the delimiter. Click "Finish". 2. Data is in Excel spreadsheet format. Check through data to ensure spark timing, test number, etc. are correct, (refer to test sheet). 3. Divide up data into each sub-test (i.e. lambda settings). 4. Average each column of data for each sub-test. 5. Open the performance data processing tool: "Ric performance spreadsheet 14mar2001).xlt" 6. Fill out "Test Conditions" in this spreadsheet. 7. Insert averaged columns of data into spreadsheet. You will need to use the "paste special" function... choose "value" and "transpose" and then paste the data. 8. Plot required graphs of performance characteristics from calculated values. Pressure data: 1. For processing a pressure datafile(*.txt format), you need the following: •  ricardojprl .exe (this is a Fortran program that averages the pressure data and produces some statistical results)  •  par.txt (afileof parameters needed by the program)  •  The pressurefileyou wish to process (you must change the name so that it is less than 15 characters long). 73  2.  Start " r i c a r d o _ p r l .exe" a n d i f y o u n e e d h e l p w i t h u s i n g it, f o l l o w the p r o m p t s a n d a help.txt f i l e w i l l be generated i n the same folder.  3.  E n t e r the and  filenames  "ricardo_par.txt", "<pressure-file>.txt", "<results  summary>.txt"  "<pr.vs.ca>.txt".  4.  T h e " < p r . v s . c a > . t x t " file c o n t a i n s t h e a v e r a g e d p r e s s u r e d a t a .  5.  O p e n the pressure p r o c e s s i n g E x c e l spreadsheet ("heat release.xls").  6.  Insert the a v e r a g e d p r e s s u r e d a t a i n the sheet " P r e s s D a t a " , p a s t i n g it o v e r the R E D column.  7.  T h e spreadsheet generates the P - V d i a g r a m s , heat release p l o t s , etc.  74  Appendix E LIST OF ACQUIRED DATA AND CALCULATED PARAMETERS  Engine and Fuel Data:  Engine Name  UBC Ricardo  Engine displacement (Litres)  0.459  Number of cylinders  1  2 or 4 stroke engine  4  CNG Lower Heating Value (kJ/kg)  49109  Natural gas atomic H/C ratio  3.924  Estimated molec wt of CNG (kg/kmol)  16.61  Cylinder bore (m)  0.081120  Piston stroke (m)  0.088900  Connecting rod length (m)  0.158013  Cylinder clearance volume (Litres)  0.041988  Dry Air/Fuel Ratio by mass (kg/kg)  16.81  Test Conditions:  Date of the test (dd/mm/yy)  20-Sep-01  Test Carried out By (Name)  C. Reynolds  Type of the test (Homogeneous or PSC Fuelling)  PSC  Engine speed setting (rpm)  2500  Number of Data Points per test (average)  165  Test Data:  (This is the averaged performance data for each relative air-fuel ratio tested) Number of Data Points Test Number Engine speed  (rpm)  Brake torque  (N*m)  Throttle Position  (%)  Spark Timing  (deg BTDC)  Exhaust lambda Intake Manifold Air Flow  (1/min)  Intake Manifold air pressure  (Barg)  Intake Manifold air temperature  (°C)  Intake Manifold NG flow  (kg/hr)  Intake Manifold NG pressure  (Barg)  Intake Manifold NG temperature  (°Q  Stratified NG Flow  (kg/hr)  Stratified NG Pressure  (Barg)  Stratified NG Temperature  (deg C)  Stratified NG EOI Timing  (deg BTDC)  Stratified NG pulse width  (degrees)  Exhaust Emission - C02  (%-dry)  Exhaust Emission - NOx  (ppm-dry)  Exhaust Emission - 02  (%-dry)  Exhaust Emission - CH4  (ppm-dry)  Exhaust Emission - CO  (ppm-dry)  Exhaust Emission - tHC  (ppm-wet)  Exhaust Manifold Pressure  (Barg)  Exhaust Manifold Temperature  (°C)  Engine oil temperature  (°C)  Engine Oil Pressure  (Barg)  76  Engine Coolant Inlet Temperature  (deg C)  Engine Coolant Outlet Temperature (deg C) Barometric Pressure  (mm Hg)  Ambient air temperature  (°C)  Ambient Relative humidity  (%)  Air and Fuel Flow: (These are the calculated parameters) Barometric Pressure  (kPag)  Ambient air density  (kg/m 3)  Manifold air density  (kg/m 3)  Intake Manifold Air Flow  (kg/hr)  Intake Air Velocity  (m/s)  A  A  Air Density ratio Pressure ratio Observed air flow  (kg/hr)  Observed air flow  (litres/sec)  Intake Manifold NG density  (kg/m 3)  Intake Manifold NG Flow  (litres/s)  Stratified NG Density  (kg/m 3)  Stratified NG flow  (litres/sec)  Stratified NG injection mass  (mg/inj.)  Manifold NG mass  (mg/cycle)  Total NG mass/cycle  (mg/cycle)  computed exhaust flow  (kg/hr)  Average exhaust temp  (°C)  A  A  Volumetric Efficiency volumetric eff. based on manifold air density volumetric eff. based on ambient air density Performance and Efficiency Brake Torque  (N*m)  Brake Torque  (ft* lbs)  Brake Power  (kW)  Brake Power  (Bhp)  BSFC - NG equivalent  (g/kW-hr)  BSFC - diesel equivalent  (lb/hp-hr)  Thermal Efficiency  (%)  BMEP  (bar)  Stratified Pilot to Total Ratio (energy basis) Combustion Analysis Partial Pressure of H20 vapor  (kPa)  Specific Humidity  (gH20/kg dry  Mols of intake 02 per mole NG Total mols of exhaust per mole of NG H20 - calculated  (%)  C02 - calculated  (%-wet)  02 - calculated  (%-wet)  C02calc / C02meas (dry-wet) 02calc / 02meas(dry-wet) A/F mass ratio - stoichiometric A/F mass ratio - total Relative A/F ratio - total  Equivalence ratio - total (ie: F/A) Total NG molar flow rate  (kmol/min)  Total mols of exhaust  (kmol/min)  Exhaust Molecular weight  (kg/kmol)  Carbon balance ratio Nitrogen balance ratio Hydrogen balance ratio Oxygen balance ratio  Volumetric (Wet-Basis) Emissions  COwet = COdry * (1-H20)  (ppm)  C02wet = C02dry * (1-H20)  (%)  NOxwet = NOxdry*(l-H20)  (ppm)  K - NOx correction factor NOxwet (ppm)-corrected 02wet = 02dry * (1-H20)  (%)  CH4wet = CH4dry*(l -H20)  (ppm)  tHCwet  (ppm)  nmHCdry  (ppm)  N2wet  (%)  Exhaust molar mass  (kg/kmol)  Total mole fraction Exhaust Emission Flows  COdry  (g/hr)  COdry  (g/kW-hr)  COwet  (g/hr)  COwet  (g/kW-hr)  C02dry  (kg/hr)  C02dry  (kg/kW-hr)  C02wet  (kg/hr)  C02wet [calculated]  (kg/hr)  C02wet [not calculated] (kg/kW-hr) NOxdry  (g/hr)  NOxdry  (g/kW-hr)  NOxwet  (g/hr)  NOxwet  (g/kW-hr)  02dry  (kg/hr)  02dry  (kg/kW-hr)  02wet  (kg/hr)  02wet  (kg/kW-hr)  CH4dry  (g/hr)  CH4dry  (g/kW-hr)  CH4wet  (g/hr)  CH4wet  (g/kW-hr)  tHCwet  (g/hr)  tHCwet  (g/kW-hr)  nmHCwet  (g/hr)  nmHCwet  (g/kW-hr)  Appendix F ENGINE DIMENSIONS/GEOMETRY  These were the dimensions used in a Ricardo WAVE modelling project. Refer to the schematic (Fig. F.l) for details of intake and exhaust sections. Dimensions (top diameter x bottom diameter x length) in mm. If the duct has a bend, the angle is indicated in brackets after the dimensions [deg]. Geometry is described from intake air in to exhaust out. Intake Components 1-1  Air cleaner  105 < 132.5  1-2  Short duct on horn  80 x21.1  1-3  Short duct on horn  69 x 8.5  1-4  Conical part of horn  69 x40x38  1-5  Pipe section before LFE restriction  40 x 530  1-6  Small tubes in LFE (modelled dimension)  0.25 x 15  1-7  Pipe section after LFE restriction  40 x 156  1-8  Fuel injection ring  40 x 10  1-9  Pipe section before surge tank  44.3 x 310  1-10  Surge tank (modelled as cylinder)  75 x340  1-11  Pipe section after surge tank  40.8 x 7.2  1-12  Top cone on throttle body  40.7 x 32.5 x39.2  1-13  Bottom part of throttle body  32.5 x 57  1-14  Intake pipe  32.5 x 31.9 x 138 [90]  1-15  Intake manifold (cylinder head)  31.9 x 32.5 x95.3 [90]  81  Exhaust Components E-l  Exhaust manifold (cylinder head)  28.6 x 30.3 x 51.5 [90]  E-2  Straight section after cylinder head  30.3 x 33.9 x 132.1  E-3  Bend in exhaust pipe  33.9 x 152.4 [90]  E-4  Straight section  33.9 x 279.4  E-5  Exhaust collector (closed end)  69.9 x 203.2  E-6  Exhaust collector (open end)  69.9 x 781.1  E-7  Straight section  60.3 x 482.6  E-8  Slightly curved section  60.3 x 127 [20]  E-9  Straight section  60.3 x 228.6  E-10  Long curved section  60.3 x 787.4 [50]  E-ll  Short curved section  60.3 x 304.8 [60]  E-12  Long straight section  60.3 x 2794  E-13  Elbow  60.3 x 157.1 [90]  E-14  Exhaust pipe through test cell roof  60.3 x 2718  E-15  Muffler  203 x 152  E-16  Elbow  60.3 x 157 [90]  E-17  Short straight section  60.3 x 178  E-18  Elbow  60.3 x 132 [90]  E-19  Short straight section (to atmosphere)  60.3 x 127  82  E-18  E-16  E-19 o O  o  E-15 (muffler)  E-14  (atmosphere)  1-1  E-14  Test Cell Ceiling  1-2 to 1-5  E-12  1-6 (LFE)  E-13  1-7 to 1-9 (including fuel injection-ring)  E-8toE-ll 1-10 E-6 and E-7  E-4.  E-2-  (surge-tank) E-l 1-15  E-3  E-5  Exhaust  1-11 to 1-13 1-14  Intake  Figure F.l Schematic of Engine Intake and Exhaust Geometry Engine Geometry 1. Bore  81.12  2. Stroke  88.9  3. Con-rod length  158.013  4. Wrist pin offset  0.0  5. Compression ratio  11.943  83  Piston Geometry  3.L88 2,283 r  0.407  Inches Figure F.2 Piston Geometry: Piston # 476P by Federal Mogul for Ford Fiesta (1978-80) Combustion Chamber Geometry (Bowl-in-Piston)  1. Cup depth  10.33 mm  2. Lateral offset  0.00 mm  3. Top cup diameter  57.90 mm  4. Bottom cup diameter  51.55 mm  5. Squish ratio, velocity  47.44%, 5.0 m/s  6. Bowl-in-piston volume  27.28 cm  7. Piston land dead volume  0.71 cm  8. Piston crown to cylinder face dead volume  2.27 cm  9. Gasket volume  7.92 cm  10. Combustion chamber volume  41.99 cm  11. Swept volume  459.46 cm  12. Total cylinder volume  501.45 cm  3  3  3  3  3  3  3  84  Cam Profiles  Crank Angle  Valve Lift (mm)  (deg)  Crank Angle (deg)  Valve Lift (mm)  5  0  14  0  23  0  32 41  0 0  41  0 0  50  0  50  0  59  0  59  0  68  0  68  0  77  0  77  0.0508  86 95 104  0 0  86 95 104  0.127 0.20066  113  0  0.35052  122  0  113 122  131  0  131  0.58166  140  0 0  140 149  0.9779 1.9304  149 158  0  5 14  0  23 32  0  0  0.27432 0.42418  0  158  167  0  167  3.05308 4.08432  176 185  0 0  176 185  6.11378  5.1562  194  0  194  6.9088  203 212  0  203 212  7.6454 8.26516  221  221  8.75284  230  0 0  230  9.11606  239  0  239  9.3345  248  0 0  248 257  9.41578  257 266  0  266  9.17448  275  0  275  8.8392  284  0  284  8.36168  293  0  293  7.71652  302  0.08382  302  7.04342  311  0.13208  311  6.12394  320  0.20828  320  5.207  329  0.28194  329  4.12242  338  0.35306  338  3.12166  347  0.4318  347  2.17932  356  0.56134  356  1.07188  365  0.98552  365  0.55372  374  1.90246  374  0.4064  0  9.37006  (Cam profiles continued) Crank Angle (deg) Valve Lift (mm) Crank Angle (deg) Valve Lift (mm) 383 2.9972 383 0.33274 392 4.09448 392 0.2667 401 5.0292 401 0.18288 410 5.99694 410 0.10668 419 6.86308 419 0.0381 428 7.6327 428 0 437 8.1915 437 0 446 8.6995 446 0 455 9.08812 455 0 464 9.30656 464 0 473 9.41324 473 0 482 9.38022 482 0 491 9.20496 491 0 500 8.86968 500 0 509 8.4328 509 0 518 7.8105 518 0 527 7.15772 527 0 536 6.33222 536 0 545 5.44576 545 0 554 4.39928 554 0 563 3.38328 563 0 572 2.30632 572 0 581 1.29286 581 0 590 0.65532 590 0 599 0.45974 599 0 608 0.38608 608 0 617 0.31496 617 0 626 0.23876 626 0 635 0.18034 635 0 644 0.1143 644 0 653 0.04572 653 0 662 0 662 0 671 0 671 0 680 0 680 0 689 0 689 0 698 0 698 0 707 0 707 0 726 0 726 0  Appendix G APPROXIMATE STOICHIOMETRIC AIR-FUEL RATIO FOR B.C. NATURAL GAS  Composition and Heating Values  Table G . l Heating Values of the Components of B.C. Natural Gas Compound  Mole % in Fuel  Molecular Mass of Component (kg/kmol)  Upper Heating Value of Component (kJ/kg)  Lower Heating Value of Component (kJ/kg)  Methane  96.3641  16.043  55517  50030  Ethane  1.9132  30.070  51903  47511  Propane  0.4115  44.097  50325  46333  i-Butane  0.0566  58.123  49347  45560  n-Butane  0.0943  58.123  49505  45719  i-Pentane  0.0233  72.150  48909  45249  n-Pentane  0.0197  72.150  49006  45345  neo-Pentane  0  72.150  48712  45052  Hexane  0.0108  86.177  48678  45103  Heptane  0.008  100.204.  48435  44921  Octane  0.0039  114.231  48251  44783  Carbon Dioxide  0.2001  44.010  0  0  Nitrogen  0.8944  28.013  0  0  TOTAL:  99.9999  Note: Molecular masses based on average molecular mass of constituents. Upper and lower heating values are for gaseous components at STP.  87  Gas Constant  It follows from the composition illustrated in Table G.l that the molecular weight of B.C. Natural gas is 16.61 and the gas constant is: 8 31434 R(kJ/kg • K) = = 0.5006 . 16.61 Stoichiometric Air/Fuel Ratio  The stoichiometric, complete combustion of one mole of dry B.C. natural gas with dry air may be represented by the following: Fuel + Air —> Products  Therefore the air-fuel ratio is calculated by: AFR = (number of moles of air) x (molecular mass of air) (number of moles of fuel) x (molecular mass of fuel) (2.032) x (137.36) = 16.81 (0.9999) x (16.61) Summary:  Hydrogen/Carbon Ratio:  3.924 mol/mol  Upper Heating Value:  54.444  Lower Heating Value:  49.11 MJ/kg  Air/Fuel Ratio by mass:  16.81 kg/kg  MJ/kg  88  

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