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Development of a CNG intensifier for high pressures Aichinger, Christoph 1993

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DEVELOPMENT OF A CNG INTENSIFIER FOR HIGH PRESSURESbyCHRISTOPH AICHINGERCand. Ing., Technische Universitat Graz (Austria), 1989A THESIS SUBMITTED IN PARTIAL FULFILLMENT OFTHE REQUIREMENTS FOR THE DEGREE OFMASTER OF APPLIED SCIENCEinTHE FACULTY OF GRADUATE STUDIESMechanical Engineering DepartmentWe accept this thesis as conformingto the required standardTHE UNIVERSITY OF BRITISH COLUMBIAMarch 1993© Christoph AichingerIn presenting this thesis in partial fulfilment of the requirements for an advanceddegree at the University of British Columbia, I agree that the Library shall make itfreely available for reference and study . I further agree that permission for extensivecopying of this thesis for scholarly purposes may be granted by the head of mydepartment or by his or her representatives . It is understood that copying orpublication of this thesis for financial gain shall not be allowed without my writtenpermission .(Signature)Department of n1=.C.ltmQ eik L	c	 IN	1,11The University of British ColumbiaVancouver, CanadaDate"1'(	 g4I	 f	 1JDE-6 (2/88)llABSTRACTA CNG intensifier has been designed to compress natural gas from a variable, elevated-pressurestorage source (20 - 200 bar) to a constant high discharge pressure (200 bar) . The intensifierdelivers variable amounts of gaseous fuel (5 - 43 kg/hr) at a maximum required pressure ratio of10:1 for fueling a heavy duty bus diesel engine which is converted to a gas-diesel engine.Based on a study of alternative intensifier concepts, a single-stage and a two-stagereciprocating crank-shaft driven intensifier prototype was built . The bore and stroke dimensionsof both intensifiers are 31 .75 mm and 82 .55 mm, respectively . The stage volume ratio of thetwo-stage intensifier is 2.78:1.The performance of the intensifier prototypes was measured at pressure ratios between6:1 and 12:1 while keeping the discharge pressure constant at 200 bar . Taking into accountvolumetric efficiency and the need to limit compression temperature for satisfactory life of theelastomer seals, the performance of the intercooled two-stage intensifier prototype proved to bedecisively superior, providing volumetric efficiencies of 92 per cent, isentropic efficiencies of 77per cent and maximum operating temperatures of 120°C with air cooling at design pressure ratio(10:1) . A bore-to-valve area ratio of 16 :1 or less was found to be adequate to limit the valve flowvelocity to Mach numbers less than 0 .1.The test results of the proposed two-stage intensifier indicate a potential for efficientintensifier operation at pressure ratios exceeding 10 :1 . However, the crank-shaft actuation of theintensifier can lead to friction- and alignment-related problems at higher piston frequencies and itis suggested that a variable displacement, hydraulic actuator could be utilized to further improveefficiency and to offer better controlability .TABLE OF CONTENTSABSTRACT llTABLE OF CONTENTS illLIST OF TABLES ixLIST OF FIGURES xACKNOWLEDGEMENTS xvi1 INTRODUCTION 11 .1 Motivation 11 .2 Natural Gas in Diesel Engines 31 .3 The UBC Intensifier-Injector System 61 .4 Requirements for a CNG Intensifier 91 .4.1 Operating Pressures 91 .4.2 Intensifier Capacity 111 .4.3 Power Consumption 121.4.4 Operating Temperatures 141 .4.5 Off-Design Performance 151 .4.6 General Requirements 161 .4.7 Summary 171.5 Objectives 182 EVALUATION OF ALTERNATIVE INTENSIFIER CONCEPTS 192.1	Introduction 192.2 Reciprocating Compressors 212.2.1 Characteristics of the Compression Cycle 21iv2.2.2 Volumetric Efficiency 262.2 .3 Methods of Capacity Control 292 .2 .3 .1 Speed Control of the Compressor Driver 292.2.3 .2 Control of Clearance Volume 312.2.3.3 Control of Gas Inlet 312.2.3.4 Control of Gas Discharge 322.3 Hydraulic Actuation 332.3.1	General 332.3.2 Circuit Arrangement 332.3.3 Design Specifications 352.3 .4 Summary 382.4 Pneumatic Actuation 392.4.1 General 392.4.2 Circuit Arrangement 392.4 .3 Alternative Air Sources 402.4.4 Haskel Gas Booster 422.4.5 Alternative Air-Driven Intensifier Designs 452.4 .6 Summary 452.5 Mechanical Actuation 462.5 .1 General 462.5.2 Camshaft Actuation 462.5.3 Crankshaft Actuation 482 .5 .4 Rotating Pistons 502 .5 .5 Summary 522.6 Summary and Conclusion 533 THE MECHANICAL DRIVE CONFIGURATION 593 .1	Design Details 59v3.1 .1 Intensifier Dimensions 593.1 .2 Valves 653.1 .3 Sealing 683.1 .4 Cooling 693.1 .5 Capacity Control 693 .2 Intensifier Components (Version 1 .1) 713.3 Improvements (Version 1 .2) 763 .3 .1 Alignment 763 .3 .2 Sealing 783 .3 .3 Lubrication 803 .3 .4 Valves 813 .3 .5 Assembly Drawing 823.4 Two Stage Intensifier Specifications (Version 2.1) 843 .4.1 General 843 .4.2 Sizing of Stages 853 .4 .3 Valves 883 .4.4 Assembly Drawing 904 EXPERIMENTAL ARRANGEMENT 924.1 Test Apparatus 924.2 Piping Plan of Single-stage Intensifier (Version 1 .2) 944.3 Piping Plan of Two-Stage Intensifier (Version 2 .1) 974.4 Data Acquisition System 1004.5 Data Acquisition Software 1035 RESULTS 1055 .1	General 1055 .2 Test Limitations 106vi1071101101111121.141151161181201251251261261291291301311331361371381421451451461475 .3 Accuracy and Repeatability5 .4 Single-stage Intensifier Perfomance5 .4.1 Intensifier Capacity5.4.2 Volumetric Efficiency5.4.3 Power Consumption5.4.4 Isentropic Efficiency5 .4.5 Operating Pressures5 .4.6 Operating Temperatures5 .4.7 Cylinder Pressure Data5 .4.8 The Pressure-Volume Diagram5 .4.9 The Effect of Intensifier Speed5 .4.10 The Effect of Changes in Discharge Pressure5 .4.11 The Effect of Cooling5 .4.12 Capacity Control5 .5 Two-Stage Intensifier Performance5.5.1 Intensifier Capacity5 .5 .2 Stage Pressure Ratios5 .5.3 Volumetric Efficiency5 .5 .4 Power Consumption5 .5 .5 Isentropic Efficiency5 .5.6 Operating Temperatures5 .5 .7 The Pressure-Volume Diagram5 .5 .8 The Polytropic Coefficient5 .5 .9 The Effect of Intensifier Speed5 .5 .10 The Effect of Changes in Discharge Pressure5 .5 .11 The Effect of Intercooling5 .6 Comparison of Single-stage and Two-Stage Intensifiervu6 CONCLUSIONS AND SUGGESTIONS 1506.1 Conclusions 1506 .2 Suggestions 1547 REFERENCES 1578 APPENDICES 159Appendix AAppendix Al Thermodynamic Properties of CNG 160Appendix A2 Engine Specifications 163Appendix A3 Sample Calculation: Centrifugal Compressor 168Appendix A4 Performance Charts of Commercially Available 171Appendix ASCompressorsHydraulic Driver Price Breakdown 177Appendix A6 Sample Calculation : Hydraulic Driver 178Appendix A7 Specification of Existing Air Compressor 182Appendix A8 Sample Calculation: "Skip-Fire" Compressor 183Appendix A9 Sample Calculation: Haskel Gas Booster 185Appendix Al Injector Dimensions and Modifications 192Appendix Al Sample Calculation: Conversion of a 50ccm 194Appendix Al2Combustion EngineSample Calculation: Capacity of a Hydraulic Pump 196Appendix A13 Piston Rod Calculation 198Appendix A14 Flow Calculation of Check Valve 200Appendix BSingle-stage Intensifier Parts (Version 1 .2) 202Appendix B1Appendix B2 Two Stage Intensifier Parts (Version 2 .1) 219viiiAppendix B3 Determination of Clearance Volume (Version 1 .2) 229Appendix B4 Determination of Clearance Volume (Version 2 .1) 233Appendix CAppendix Cl Equipment Specification : Test Apparatus 237Appendix C2 Equipment Specification : Instrumentation 239Appendix C3 Equipment Specification : Data Acquisition 253Appendix C4 Test Procedure 255Appendix C5 Performance Calculations : Single-stage Intensifier 257Appendix C6 Performance Calculations : Two-Stage Intensifier 261Appendix DThe Effect of Cooling 266Appendix D1Appendix D2 Single-stage Intensifier Capacity 268Appendix D3 The Effect of Changes in Discharge Pressure 269Appendix ETwo-Stage Intensifier Capacity 271Appendix ElAppendix E2 P-V Diagrams of the Two-Stage Intensifier 272ixLIST OF TABLESTable 1.1 EPA Emission Standards 2Table 2.1 Comparison between k and n 24Table 2.2 Comparison of Alternative Drive Configurations 57Table 3.1 Single-stage Intensifier Dimensions 64Table 3.2 Two-Stage Intensifier Dimensions 87Table 4.1 Legend for Figure 4 .2 .1 96Table 4.2 Legend for Figure 4 .3 .1 100Table 5.1 Comparison between Expected and Measured Single-stageIntensifier Capacity111Table 5.2 Comparison between Expected and Measured Two-StageIntensifier Capacity131Table 5.3 Polytropic Coefficients of the Two-Stage Intensifier 143Table A1.1 Composition of Natural Gas 160Table A1.2 Composition of Natural Gas 160Table A13 Mass of Natural Gas 161Table A1.4 Gravitation Factor of Natural Gas 162Table A1.5 Lower Heating Value of Natural Gas 162Table A2.1 Performance Data of Detroit Diesel 6V-92 TA Engine 164Table B1 Parts List for Single-stage Intensifier (Version 1 .2) 204Table B2 Parts List for Two-Stage Intensifier (Version 2 .1) 221Table B3 List of Clearance Spaces of Single-stage Intensifier 231Table B4 List of Clearance Spaces of Two-Stage Intensifier 236Table C2 Pressure Transducer Calibration Intervals 243xLIST OF FIGURESFigure 1.2 .1 Methods of using Natural Gas in Diesel Engines 4Figure 1 .3.1 Schematic of the UBC Intensifier-Injector System 7Figure 1 .3.2 Intensifier Injector Research Outline 8Figure 1 .4 .1 Change of Relative Fuel Consumption, Pressure Ratio andCompressibility with Tank Pressure10Figure 1 .4 .2 Fuel Consumption of Detroit Diesel Engine at Full LoadConverted to CNG12Figure 1.4.3 Power Consumption versus Pressure Ratio forVarious Mass Flow Rates13Figure 1.4.4 Relative Power Loss Due to Gas Compression 14Figure 1 .4 .5 CNG Compression Temperatures for Various Inlet Temperaturesand Pressure Ratios15Figure 2.1 .1 Classification of Compressor Types 20Figure 2.2 .1 Typical Reciprocating Compressor Cycle 22Figure 2.2 .2 Reciprocating Compressor Modes 25Figure 2.2 .3 Single Acting versus Double Acting Mode 25Figure 2.2 .4 Impact of Clearance Volume on Volumetric Efficiency 28Figure 2 .2.5 The Effect of Add-On Clearance Spaces 31Figure 2 .2.6 Progressive Indicator Card Illustrating the Operation ofInfinite-Step Control32Figure 2 .3 .1 Hydraulic Driver Configuration 34Figure 2.3 .2 Piston Rod - Stroke Selection Chart 35xiFigure 2 .33 Hydraulic Cylinder Bores versus Operating Pressure andCorresponding Pump Size36Figure 2 .3 .4 Hydraulic Circuit Arrangement 38Figure 2.4.1 Air-Driven Intensifier Arrangement 40Figure 2.4.1 Haskel Gas Booster 43Figure 2 .5.1 Camshaft-Actuated Plunger 47Figure 2 .5.2 Direct Transmission versus Crosshead Transmission 49Figure 2 .5.3 HYDURA PVWH Hydraulic Open Loop Pump 51Figure 2 .6.1 Hydraulic Pump Efficiency Chart 55Figure 3 .1 .1 Volumetric Efficiency and Displacement Factor as aFunction of Relative Clearance Volume61Figure 3.1 .2 Basic Intensifier Displacement versus Intensifier Speed 62Figure 3.1 .3 The Effect of Bore Size on Stroke Length and Thrust Force 63Figure 3.1.4 Expected Intensifier Capacity of Single-stage Intensifier 65Figure 3.1 .5 NUPRO Check Valve ("CH" Series) 66Figure 3 .1.6 Axial Cut Through Check Valve ("CH" Series) 67Figure 3 .1.7 The PolyPak Sealing System 68Figure 3 .1 .8 Proposed Flow Control System 70Figure 3 .2 .1 Intensifier Prototype Version 1 .1 73Figure 3.2.2 Intensifier Piston Configuration 75Figure 3.3.1 Crank Case Piston versus Crosshead 76Figure 33 .2 INA Permaglide ® Bushing 77Figure 3.3 .3 Floating Rod Attachement 78Figure 3.3.4 Typical V-Packing Arrangement 80Figure 3 .3.5 Valve Poppet 82Figure 3 .3.6 Intensifier Prototype Version 1 .2 83Figure 3 .4 .1 Two-Stage Intensifier Concept 85Figure 3.4.2 Stage Pressure Ratios of Two-Stage Intensifier 86Figure 3.4.3 Expected Volumetric Efficiency of Two-Stage Intensifier 87Figure 3 .4.4 Expected Intensifier Capacity of Two-Stage Intensifier 88Figure 3 .4.5 Valve Configuration of the First Stage 89Figure 3 .4.6 Two-Stage Intensifier Prototype Version 2 .1 91Figure 4.1.1 Testrig 93Figure 4.2.1 Piping Plan for Single-stage Intensifier 95Figure 4.3.1 Piping Plan for Two-Stage Intensifier 99Figure 4.4.1 Intensifier Data Acquisition System 101Figure 5 .3.1 Sample Data Regression for Volumetric Efficiency 108Figure 5 .3.2 Frictional Torque Due to Bearings and Flywheel 109Figure 5 .4 .1 Measured CNG Massflow of Single-stage Intensifier 110Figure 5 .4.2 Volumetric Efficiency of Single-stage Intensifier 112Figure 5 .4.3 Intensifier Power Consumption 113Figure 5 .4.4 Intensifier Power Consumption Per Unit Mass 113Figure 5 .4 .5 Isentropic Efficiency of Single-stage Intensifier 114Figure 5.4.6 Operating Pressures at 200 RPM Intensifier Speed 116Figure 5.4.7 Operating Temperatures at 200 RPM Intensifier Speed 117Figure 5.4.8 Cylinder Pressure Data at 129 RPM Intensifier Speed and200 Bar Discharge Pressure119Figure 5 .4.9 Cylinder Pressure Data at 193 RPM Intensifier Speed and200 Bar Discharge Pressure119Figure 5 .4.10 The p-V Diagram at 129 RPM Intensifier Speed 121Figure 5 .4 .11 The p-V Diagram at 193 RPM Intensifier Speed 122Figure 5.4.12 The ln(p)-ln(V) Diagram at 129 RPM Intensifier Speed 124Figure 5.4.13 The ln(p)-ln(V) Diagram at 193 RPM Intensifier Speed 124xiiiFigure 5.4.14 Back Pressure Due to Bypassing 126Figure 5.4.15 Cylinder Pressure Due to Valve Unloading 127Figure 5.4.16 Power Consumption Due to Flow Control Methods 128Figure 5 .5.1	Two-Stage Intensifier Capacity 129Figure 5 .5.2	Stage Pressure Ratios of the Two-Stage Intensifier 131Figure 5 .5.3	Volumetric Efficiency of the First Stage 132Figure 5.5.4	Volumetric Efficiency of the Second Stage 132Figure 5.5.5	The Effect of Operating Temperature on the Measured Torque 133Figure 5.5.6	Power Consumption of the Two-Stage Intensifier 134Figure 5.5.7	Power Consumption per Mass Flow Rate of the Two-Stage Intensifier 135Figure 5.5.8	Isentropic Efficiency of the Two-Stage Intensifier 136Figure 5.5.9	Operating Temperatures of the Two-Stage Intensifier 137Figure 5 .5.10 The p-V Diagram of the First Stage at 100 rpm and 1384.02 : 1 Stage Pressure RatioFigure 5.5.11 The ln(p)-ln(V) Diagram of the First Stage at 100 rpm and 1394 .02 : 1 Stage Pressure RatioFigure 5.5.12 The p-V Diagram of the Second Stage at 100 rpm and 1402 .62 : 1 Stage Pressure RatioFigure 5.5.13 The ln(p)-ln(V) Diagram of the Second Stage at 100 rpm and 1402.62 : 1 Stage Pressure RatioFigure 5.5.14 The p-V Diagram of Both Stages at 100 rpm and 14110.5 : 1 Overall Pressure RatioFigure 5.5.15 The p-V Diagram of Both Stages at 100 rpm and 1422.9 : 1 Overall Pressure RatioFigure 5.5.16 The Steady-State Polytropic Coefficient of the First Stage 144Figure 5.5.17 The Effect of Changes in Discharge Pressure 146Figure 5 .5.18 Intercooling Temperatures 146Figure 5 .5.19 The Effect of Intercooling 147xivFigure 5 .6 .1 Power Consumption per Unit Mass for Single Stage andTwo-Stage Intensifier148Figure 6 .2 .1 : Proposed variable displacement, hydraulically-driven reciprocating 156intensifier prototype, variably operated as single stage and two-stageintensifier.Figure A2 .1 Power Output of Detroit Diesel 6V-92 TA Engine 165Figure A2.2 Break Specific Fuel Consumption of Detroit Diesel 6V-92 TA Engine 165Figure A2 .3 CNG Fuel Consumption (Mass) of Detroit Diesel 6V-92 TA Engine 166Figure A2 .1 CNG Fuel Consumption (Volume) of Detroit Diesel 6V-92 TA Engine 167Figure A3 .1 Comparison of Calculated Efficiency Contours with Test Dataon Centrifugal Pumps168Figure A4 .1 General Capacity Chart of Sulzer "Labyrintpiston" - Compressors 172Figure A4.2 General Capacity Chart of Neuman & Esser Compressors 1.73Figure A4 .3 General Capacity Chart of Atlas Copco Compressors 174Figure A4.4 General Capacity Chart of LMF Compressors 175Figure A4.5 Ingersoll-Rand High Pressure, Air-Cooled Piston Compressors 176Figure A7.1 Technical Data Sheet of Tu-Flo 700 Compressor 182Figure A9.1 Haskel Model Selection Chart 185Figure A9.2 Haskel Performance Data 186Figure A9 .3 Haskel Performance Calculations 187Figure A13 .1 Forces and Stresses for Various Rod Sizes 198Figure B1 .1 Assembly Drawing of Single Stage Intensifier Version 1 .2 203Figure B2 .1 Assembly Drawing of Two-Stage Intensifier Version 2 .1 220Figure B3 .1 Expected Volumetric Efficiency of Single Stage Intensifier Version 1 .2 232xv240242242244246246Figure C2.2Figure C23Figure C2.4Figure C2.5Figure C2.6Figure C2.7Figure C2.8Figure C2.9Calibration of Strain Gage Type LoadcellCalibration Curve of Inlet PressureTransducerCalibration Curve of Outlet PressureTransducerCalibration Curve of Interstage Pressure TransducerCalibration Curve of Inlet Temperature ThermocoupleCalibration Curve of Wall, Outlet and InterstageTemperature ThermocouplesCalibration Curve of Tachometer used on the Electric MotorOriginal Calibration Chart for PCB PressureTransducer # 10671248250Figure C2.10 Original Calibration Chart for PCB PressureTransducer # 10672	251Figure D1.1Figure D1.2Figure D13Figure D2.1Figure D2.2Figure D3.1:Figure D3.2:Figure D33:Figure E1.1Figure E1.2Figure E2.1Figure E2.2Figure E23Figure E2.4Figure E2.5Figure E2.6The 1n(P)-ln(V) Diagram at Design Pressure Ratio (no cooling).The ln(P)-ln(V) Diagram at Design Pressure Ratio (air cooling)The ln(P)-ln(V) Diagram at Design Pressure Ratio (water cooling)Single Stage Intensifier Capacity in Normal Cubic Meters per HourSingle Stage Intensifier Capacity in Standard Cubic Feet per MinuteVolumetric Efficiency at 200 bar Discharge Pressure and 150 rpmVolumetric Efficiency at 175 bar Discharge Pressure and 150 rpmVolumetric Efficiency at 150 bar Discharge Pressure and 150 rpmTwo-Stage Intensifier Capacity in Normal Cubic Meters per HourTwo-Stage Intensifier Capacity in Standard Cubic Feet per MinuteThe p-V Diagram at 100 rpm and 4 .6 : 1 Overall Pressure RatioThe p-V Diagram at 100 rpm and 6 .7 : 1 Overall Pressure RatioThe p-V Diagram at 150 rpm and 10 .4 : 1 Overall Pressure RatioThe p-V Diagram at 150 rpm and 8 .9 : 1 Overall Pressure RatioThe p-V Diagram at 150 rpm and 6 .4 : 1 Overall Pressure RatioThe p-V Diagram at 200 rpm and 12 .1 : 1 Overall Pressure Ratio266266267268268269269270271271272272273273274274xviACKNOWLEDGEMENTSI wish to express my sincerest gratitude to Professor Dr . Philip G. Hill for accepting me to dograduate studies at UBC under his supervision and for his help to mature on a challengingresearch project.Thanks are due to K. Bruce Hodgins, research engineer and project manager, whoprovided support and helpful advice throughout the entire project, and taught me the word"double-check" . Thanks to Dale Nagata for solving software problems and for introducing me tothe miracles of applied electronics.The staff of the mechanical engineering machine shop has done a great job in turning myideas into an intensifier . Thanks in particular to Len Drakes, Tony Besic, Doug Yuen and"Meister" Anton Schreinders for their support, their creative input to my project and the manymachining lessons.I would like to thank my collegues Paul Walsh, Hardi Gunawan, Yinchu Tao, DehongZhang, Alexander Chepakovic, Brad Douville, Art McDonald, Alain Touchette and Peter Mtuifor making the combusiton lab an awesome place to study . Special thanks to Patric Ouellette foradvice and friendship.The kind welcome by the staff of International House and its support throughout mystudies are gratefully acknowledged.I want to offer very special thanks to my wife, Susi, and my daughter, Reingard, who werea constant source of moral support and motivation and who made Vancouver a home so far awayfrom home . Thanks, also, to my mother-in-law, Irmtraud, who contributed to the success ofthese studies in her own way.Schlief1ich mochte ich meine Eltern, Fritz and Christiane, erwahnen, denen ich viel mehrzu verdanken habe, als die Freude, dieses Werk in Handen zu halten ."Der Weg ist das Ziel"(Goethe)1Chapter 1IntroductionA CNG 1 Intensifier is a natural gas compressor which is located on-board an urban bus. Itspurpose is to compress and supply gaseous fuel to a heavy duty diesel engine which has beenconverted to direct injection of natural gas into the engine cylinder . The newly developed gasdiesel injector requires constant high gas pressure . Thus, the intensifier compresses (i .e.intensifies) natural gas from variable tank pressures up to a fixed injection pressure matching itscapacity with the constantly changing fuel requirement of the diesel engine . The term intensiferhas been chosen because the compressor operates at elevated inlet pressures rather thanatmospheric inlet pressure.1.1 MotivationThe importance of using the earth's natural resources efficiently and responsibly has becomeevident. Alarming levels of pollution in urban areas, global atmospheric pollution and thegreenhouse effect are direct consequences of the massive use of fossil fuels for transportationand industrial processes . In an effort to ameliorate urban air quality and to reduce generalatmospheric pollution, the Environmental Protection Agency (EPA) in the United Statesestablished the Clean Air Act Amendments (CAAA) . The CAAA set regulations that target,among other sources of pollution, passenger cars, buses and trucks.Buses and trucks have been traditionally powered by diesel engines because of theirgreater durability and higher thermal efficiencies compared to gasoline engines . Diesel exhaust1 Compressed Natural Gas2is characterized by concentrations of nitrogen oxides (NO R) similar to those of gasoline engines,slightly lower concentrations of unburned hydrocarbons (HC) and carbon dioxide (CO 2) andcarbon monoxide (CO). However, diesel engines are a major source of particulate matter (PM)emissions [7].Recognizing the impact of diesel exhaust in urban areas, the new EPA regulations callfor a reduction of 90% in PM and 15% in NO R for urban buses between 1990 and 1994 . Theproposed emission standards are summarized in Table 1.1.URBAN BUSES HEAVY-DUTYENGINE EMISSION STANDNOx HC CO PM NOx HC CO PM1990 6 .0 1 .3 155 0.60 6.0 1 .3 153 0.601991 5.0 1.3 15.5 0.25 5.0 1 .3 15.5 0.251992 5 .0 1 .3 155 0.10 5.0 1 .3 155 0.101993 5 .0 1 .3 155 0.05 -	5 .0 1 .3 155 0.051994 4.0 1.3 15.5 0.05 4.0 1 .3 15.5 0.05Table 1.1: EPA Emissions Standards (Source : 1991 Detroit Diesel Information Update)Satisfying the stringent NOR and PM emissions simultaneously becomes increasingly difficulteven with latest diesel technology. Improved diesel engine design, electronic control andparticulate traps are among the strategies to solve the problem . Electronic control of theinjection has significantly improved the emission characteristics, but further improvements arenecessary to meet the 1993 standards . Particulate traps are being designed and testedsuccessfully, but reliability and cost are concerns.Another strategy is to use an alternative fuel . At present, methanol and natural gas are the mostpromising alternative fuels for diesel engines.3Methanol engines have been field tested and have met both NOR and PM emissionstandards . However, since methanol is industrially derived from natural gas (or coal) throughchemical processes, its cost as engine fuel is considerably higher than that of natural gas . Thecorrosive characteristics of methanol require the replacement of engine parts by more resistantmaterials . Methanol bums with an invisible flame and the exhaust emissions contain aldehydeswhich are pollutants not yet regulated by the EPA . These factors render methanol as a lessattractive alternative fuel.The history of natural gas used as engine fuel goes back to the early 1920's . Problemsassociated with handling and storage of pressurized gases limited the usage of natural gas asengine fuel . Today, natural gas can be stored in gaseous form (CNG) and in liquified form(LNG2). Both methods are save and reliable . The main advantage of natural gas as alternativeengine fuel is its availability (particularly in North America), its low cost, and its potential forclean burning. Although, these factors are in favour of natural gas, its application to dieselengines faces some challenges.1.2 Natural Gas in Diesel EnginesSince the auto-ignition temperature of natural gas is higher than that of liquid diesel fuel, auto-ignition cannot be achieved by the compression ratio of conventional diesel engines . Use ofsmall amounts of diesel fuel (i .e. pilot diesel) can preheat the cylinder content sufficiently toprovide auto-ignition of the diesel fuel.The amount of pilot diesel can be either fixed or variable . Dual-fuel engines can operate ondiesel fuel only under certain conditions, whereas gas-diesel engines always operate with bothfuels.2 Liquified Natural Gas4Figure 1.2.1 : Methods of using natural gas in diesel engines [2].HIGH —PRESSURE	DIESEL FUELNATURAL GASCC) DIRECT INJECTION5Three principal methods of using natural gas in diesel engines are shown in Figure 1.2.1:• Manifod injection of natural gas (MING)• Timed port injection of natural gas (PING)• Direct injection of natural gas (DING)In MING engines (also known as naturally fumigated engines) natural gas is injected into theinlet manifold (Fig . 1.2.1 A) forming a naturally pre-mixed fuel-air mixture . The advantage ofthis method is associated with its simplicity . It can be easily adapted to existing diesel engines.However, this method lacks in part-load performance due to poor flammability of lean mixturesand is susceptible to knock and incomplete combustion . High quantities of pilot diesel arerequired . The application of this method is limited to four-stroke engines which operate atrelatively constant speed and load . In two-stroke engines, fuel is lost due to the scavengingprocess.As an attempt to solve the problems imposed by the MING engine, the timed port injection ofnatural gas (PING) has been developed where natural gas is injected into the inlet manifold closeto the inlet port (Fig.1.2 .1 B) or into the cylinder at moderate pressure close to the time of inletport closure . Although better stratification can be achieved due to timing of the injection, knockand incomplete combustion are still a concern . Reduction in compression ratio, throttling andrelatively high piot diesel quantities are required with the overall effect of reduced efficiency.In a DING engine, high-pressure gas is injected directly into the combustion chamber near theend of the compression stroke (Fig . 1.2.1 C). Thus, a full stratification of fuel-air mixture canbe obtained resulting in good flammability over the entire load range with only a small pilotdiesel quantity. Successful operation has been demonstrated in applying this method to amedium speed railway engine 3 . Previous work on marine diesel engines has shown that direct3 1n this case seperate injectors for diesel and natural gas were used with 2% pilot diesel quantity .6injection of natural gas coupled with diesel pilot ignition can provide high efficiencies and lowemissions.The advantages of this method can be summarized as follows [11]:• Use of the basic diesel cycle with compression ignition of pilot diesel followed by highpressure gas injection.• No detonation limit if gas injection is simultaneous with liqid fuel pilot injection• No throttling• Lean burning, requiring no mixture ratio control• Diesel cycle efficiency• Negligible unburned fuel in the exhaustHowever, there is one major disadvantage associated with this method if CNG is used as fuel:An on-board compressor (i .e . intensifier) is needed to generate the required gas pressure.1.3 The UBC Intensifier-Injector SystemA research team at the University of British Columbia (UBC) has conducted studies in naturalgas fuelling of IC engines for several years and is currently engaged in the development of aCNG direct injection system for heavy-duty bus engines . A prototype injector has been designedand patented by P .G. Hill, K.B . Hodgins from the Department of Mechanical Engineering atUBC and R.J. Pierik a former graduate student and research engineer at UBC . The prototypeinjector is intended to replace the existing diesel injector in the series 60 and 71 engines withoutmodifications of the engine, itself. Together with the intensifier and the on-board CNG storagetanks this conversion kit presents itself as an efficient, economically viable alternative to existingengine modifications .7Figure 1.3.1 : Schematic of the UBC Intensifier-Injector System4 .A schematic of the intensifier-injector system is presented in Fig. 1.3.1. An existing DetroitDiesel Electronic Control system (DDEC) has been modified to control timing of injection andfuel quantity. The cam-actuated plunger pressurizes the diesel liquid which in turn forces apoppet valve to open so that both diesel and natural gas enter the combustion chamber,simultaneously. When a sufficient amount of fuel has been admitted to the combustion chamber,the diesel pressure is released by means of opening the diesel supply/return valve. The poppet issubsequently closed by a counter-acting spring . A metering valve controls the amount of pilotdiesel which is gas-blast atomized by the natural gas flow at the injector nozzle . High pressurenatural gas is provided by the intensifier.4lmplementation of the "Intensifier-Injector Technology" [12) . Courtesy of K .B . Hodgins.ACCESSORY SHAFTDRIVEN ACTUATORDDECSOLENOIDPOPPET SEAT ANGLE/	U .S. Patent k 5,067,467 November 19918The overall research project is conducted in four contributing areas as outlined in Fig. 1 .3.2.Design and improvements of the injector prototype are supported by flow visualization of thefuel injection and numerical simulation of the injection.ENGINE CONVERSIONINJECTORDEVELOPMENTINTENSIFIERDEVELOPMENTINJECTOR7 PERFORMANCEDetroit-Diesel 71Single CylinderFLOWVISUALIZATIONSchlieren Photograph)Atmospheric PressureCOMPUTERSIMULATIONTEACHCODESINGLE STAGEECIPROCATINGI TWO-STAGEECIPROCATINGI Detroit-Diesel7 6V-92 TASchlieren + LaserPressurized ChamberKIVACODETWO-STAGEROTARYINJECTORIMPROVEMENTINTENSIFIERIMPROVEMENTENGINEFIELD TESTINGFigure 1.3 .2: Intensifier-Injector Research Outline.The objective of the numerical simulation of the gas injection is to predict the effect of changesin injector geometry and injection conditions and to get an understanding of the interaction ofthe various physical and chemical processes . Thus, it is intended as a design tool to optimize the9injector performance and to suggest design changes . The numerical work has originally beendone using the TEACH code and is currently being conducted on the KIVA II code whichincludes multidimensional mathematical models for spray dynamics, combustion and turbulence.The calculations are based on a moving piston.The performance of each injector prototype is investigated on a specially designed testapparatus. With a high speed video camera a picture of the gas injection is taken in small timeintervals . The natural gas jet is made visible using schlieren and laser technology . Jetpropagation and penetration are measured and evaluated . The objective of these optical studiesis to closely simulate real engine injection to evaluate injector performance.With information obtained from flow visualization and numerical simulation and with theperformance data of a single-cylinder and a 6-cylinder Detroit Diesel engine, the injector can beoptimized and a bus engine prepared for field testing.1 .4 Reauirements for a CNG Intensifier1 .4.1 Operating Pressures,The intensifier is located between the fuel tanks and the injectors (Fig . 1 .3.1). Hence, the inletpressure is the current CNG tank pressure and the discharge pressure is determined by thepressure requirement of the injector.To ensure efficient fuel mixing and burning in the combustion chamber of the diesel engine, thegas must be injected at approximately 200 bar (3000 psi) . Experiments are still beingundertaken to determine optimum injection pressure . A pressure variation of less than ±10% isexpected to be acceptable . This means that the discharge pressure of the compressor has to be200 bar . The maximum allowable tank pressure in a CNG storage tank is 200 bar. Therefore,10the maximum intensifier inlet pressure is 200 bar . The gas pressure decreases due to fuelconsumption . The important considerations in selecting a minimum allowable tank pressure arecompression work and space and weight of the tanks (corresponding to a given mass of gasconsumed between fillings).Figure 1.4.1 shows the change of pressure ratio versus existing tank pressure, as well as therelative fuel consumption taking the compressibility factor of methane into account . Thecompressibility factor has been obtained from a general compressibility chart [6] for a constantinlet temperature of 288 K (Tcr = 220 K, per = 4 .46 MPa) . The mass in a constant volume canthen be calculated for various tank pressures:m	 PTank	V R • TTank - Z(1.4.1)where Z indicates the compressibility factor for the corresponding tank pressure.RELATIVE FUEL CONSUMPTION1'= 288 K, GASCONSI'ANT R= 500.9 J/(kg*K)150100Compressibility Factor Z50 —Relative Fuel Consumption (%)Pressure Ratio (*10)o -0	20	50	100	150	200	250TANK PRESSURE (bar)Figure 1.4.1 : Dependence of relative fuel consumption, pressure ratio (p 2/pi) andcompressibility factor (Z) on the current tank pressure.The relative fuel consumption is the ratio of the mass at a specific tank pressure to the mass atmaximum tank pressure (i .e . full tanks).11m	 V/Tank ,1-m	100V )FullThe pressure ratio increases infinitely towards low tank pressures . The goal is to choose aminimum tank pressure where the pressure ratio is within a reasonable range, yet, most of thefuel has been consumed.This consideration lead to the compromise of adjusting the minimum tank pressure at 20 bar . Atthis point, 92.2% of the fuel has been consumed and the pressure ratio is 10 :1 . Hence, theminimum inlet pressure is 20 bar and the maximum pressure ratio is 10 :1.The decision was made without any compressor type in mind and is a goal to aim for, rather thana strict design constraint.1.4.2 Intensifier CauacityThe information about diesel fuel consumption for the diesel engine was provided by the DetroitDiesel Company . It is given in terms of "BSFC" (i .e . "Brake Specific Fuel Consumption" inkglbhp hr) for full load at various engine speeds . To get the actual diesel fuel consumption overa range of engine speeds, the BSFC has been multiplied by the power output for the same speedrange. Power output and fuel consumption were given for 1000 - 2100 rpm which does notrepresent the entire speed range . The idle speed is as low as 600 rpm . Therefore, the values for600 -1000 rpm have been interpolated from the existing curves based on the behaviour ofpreviously investigated diesel engines (for calculations and graphs see Appendix A2).The diesel fuel consumption was then converted to natural gas by replacing 95% of diesel fuelby natural gas, since a small amount of diesel fuel (approximately 5%) is still injected as ignitionsource for the natural gas.Relative Fuel Consumption (%) = (1.4.2)12Figure 1 .4.2: Fuel consumption of Detroit Diesel 6V-92 TA engine at full load converted toCNG.Based on the lower heating values, the mass of natural gas was calculated, equivalent to the totalenergy of the replaced diesel mass . The fuel consumption is plotted versus engine speed in Fig.1.4.2.The total maximum CNG fuel flow for full load is 43 kg/hr @ 2100 rpm . The intensifier isdriven by the crankshaft of the engine and operates at a portion of the engine speed . Hence, it isimportant to know the maximum fuel consumption per engine revolution . It is 460 mg/rev forfull load and occurs at 600 rpm (i .e . idle speed).1.4.3 Power ConsumptionIf the natural gas were considered ideal (Z = 1) and the compression were to be isentropic withk = cp/cv = 1 .3 the power consumption could be estimated fromCNG FUEL CONSUMPTION (MASS)DEI ROIT DIESF 6V-92TA 285 HP1500ENGINE sPH	 ]) [RPM]kg/hr	g/rev (* 100)0.20.10 .50 .40500g/revkg/hrwoo 2000025004020501013k-1k	kP=	-m-R-T1• Pz	1k — i	P1where P indicates the power cosumption in kW, rn the massflow in kg/s, k the isentropiccoefficient of natural gas (k = 1 .3), pi and p2 the CNG inlet and discharge pressures, respectivelyand R the specific gas constant for natural gas (R 500.9 J/(kg*K).The power consumption is mainly governed by two factors, namely pressure ratio and massflow.Hence, the maximum power consumption occures at maximum mass flow rate and maximumpressure ratio . Figure 1.43 shows the power consumption for isentropic compression of naturalgas5 for various mass flow rates.The maximum power consumption at maximum fuel flow (assuming isentropic compression andno losses) is 5 .3 kW.INTENSIFIER POWER CONSUMPTIONFOR VARIOUS MASSFLOW S, CNG INLET 1'P1v PERAI'URE 293 K0	1	2	3	4	5	6	7	8	9	10PRESSURE RATIO (p2/p 1)_._. . 43 kg/hr	35 kg/hr	 — 25 kg/hr	15 kg/hrFigure 1.43: Power consumption versus pressure ratio for various mass flow rates.5The specific composition of natural gas is given in Appendix A2(1.43)14RELATIVE POWER LOSS DUE TO GAS COMPRESSIONk=1 .3,R=500.9JJkgK,T =288Kwoo	150oENGINE SPEW [fpm]PRESSURE RATIO = 2 :1	__.._. PRESSURE RATIO = 5 :1	 PRESSURE RATIO = 10:1	PRESSURE RATIO = 15 :1Figure 1.4 .4: Relative power loss due to gas compression.The power loss due to gas compression is defined as the intensifier power consumption dividedby the engine power output at a specified engine speed . The values for CNG mass flow andengine power output have been taken from Figure A2 .3 and Figure A2.1, respectively(Appendix A2) . Figure 1 .4.4 illustrates the maximum power loss due to gas compression forvarious pressure ratios . It can be seen that the gas compression at 10 :1 pressure ratio causes(ideally) a maximum power loss of about 2 .8 % with isentropic compression.1.4.4 Operating TemueraturesPower consumption and temperature rise due to the isentropic compression of natural gas (Z = 1)are a function of the pressure ratio, the gas properties and of the inlet condition of the gas.Figure 1 .4.5 illustrates the temperature rise for various inlet temperatures assuming single stagecompression . It can be seen that temperatures as high as 240 °C can be expected for high inlet15temperatures (on hot days) due to compression. In addition, frictional heat increases theoperating temperatures even more.Figure 1.4.5: CNG compression temperatures for various inlet temperatures andpressure ratios.The fact that the operating temperatures can exceed 200°C requires special materials for valvesand seals .It also raises the question of how to cool the system, since high operating temperaturesgreatly reduce the lifetime of parts and materials.1 .4.5 Off-Design PerformanceThe intensifier is powered by the crankshaft outlet of the diesel engine . Hence, the intensifiercapacity is a function of engine speed . The intensifier capacity is also a function of the currenttank pressureas can be seen from Eqn 1.4 .4GAS TEMPERATURE DUE TO COMPRESSIONFOR .VARIOUS INLET' T1:1v1PERATURES300INLET TEMPERATURE (deg C)ISENTROPIC COMPRESSION :	40n = 1 .3Z=1150loo5o0-5o5	6	7PRESSURE RATIO3 420 8 10 119250200P1 VD n engine 1 vm=R-T,(1 .4 .4)16where p1 is the current tank pressure, V D the intensifier displacement per engine revolution,nengine the engine speed in rpm and a y the volumetric efficiency . The effect of volumetricefficiency is neglected at this point . The intensifier capacity must be designed to satisfy the fuelrequirement of the engine at the worst case which is minimum tank pressure and full load overthe entire speed range6. However, the engine fuel consumption is a function of engine speed andload. Therefore, the intensifier capacity can exceed the engine fuel requirement, particularly, ifthe tank pressure is higher than minimum tank pressure and if the engine load is lower than fullload. This calls for an effective capacity control method to match engine fuel requirement andintensifier capacity.The engine fuel consumption can vary rapidly during the bus operation (e .g. departure from busstation, acceleration, etc .) . An accumulator is located between intensifier and injectors (ref . Fig.1.3.1) which stores high pressure CNG to ensure sufficient fuel for instantaneous high fueldemands . When the engine is started the accumulator provides fuel for the first few strokeswhile the intensifier builds up the gas pressure.1.3.6 General RequirementsThe fact that the compressor is located on board the bus imposes various requirements to thedesign which differ considerably from other compressor designs . Besides the technical aspectsdiscussed previously, a number of qualitative aspects must be taken into account.The engine room of an urban bus does not offer much space for engine add-ons. Therefore,every effort must be undertaken to keep the design as small as possible . Any space requirementfor intensifier driver configurations and transmissions must be kept in mind as well . Size andweight of the intensifier system are important design points.6ref. Fig . 1 .4 .217All moving engine parts of the diesel engine are required to last for at least 300,000 miles.Engine manufacturers want to stretch this limit to 500,000 miles before doing a complete engineoverhaul . Since the intensifier will be part of this engine, it is a goal to emphasize durability inthe design.Because of wear during regular operation, compressor seals and valves are not expected to lastfor 300,000 miles ; they must be changed as part of routine maintenance . Therefore thecompressor must be easily accessible in order to facilitate maintenance work.Cost is one of the most important factors since a CNG conversion is only attractive if it is asefficient as the diesel operation and if it is available at a reasonable price . Cost means both costfor the conversion kit and cost of operation and maintenance . The total cost can be considerablyreduced if it is possible to take advantage of existing hardware, such as transmissions,crankcases, etc ..1.4.7 SummaryAll technical and general requirements are summarized here for evaluation of alternative designconcepts, discussed in the next part.TECHNICAL REQUIREMENTS:- operation on board of an urban bus, model Detroit Diesel 6V-92 TA- maximum compressor capacity : m = 43 kg /hrm = 460 mg/engine revolution- inlet pressure : variable from 20 - 200 bar (300 - 3000 psig)- discharge pressure : fixed at 200 bar (3000 psig)- flow control unit to adjust compressor capacity to fuel requirement- air cooling18GENERAL DESIGN PARAMETERS (in keywords):- durability - sealability - ease of maintenance- size - weight - availability of existing hardware- cost	- complexity of the system1 .5 ObjectivesThe objective of the work described in this thesis was:(i) to discuss and evaluate alternative intensifier design concepts based on the requirementsstated in this chapter and to select the most promising concept(ii) to design and build a single-stage intensifier prototype(iii) to develop adequate testing facilities including a test stand, the arrangement ofmeasurement devices and a data acquisition program for steady-state and high-speed data(iv) to measure the performance of the single stage intensifier and assess design limitations(v) to design and build a two-stage intensifier prototype(vi) to measure the performance of the two-stage intensifier and compare the results withthat of the single-stage intensifier19Chapter 22. Evaluation of Alternative Design Concepts2.1 IntroductionThe purpose of this chapter is to discuss and list advantages and disadvantages of eachalternative design concept based on the requirements stated in the previous chapter. First, allcompressor types are considered . Then the most feasable compressor types are evaluated in apreliminary analysis. Based on this evaluation, the most promising design concept has beenselected.Compressors differ in a variety of characteristics, such as capacity, principle of design, pressureratio, end pressure, type of gas, working conditions, et cetera . In general, there are two majorcompressor groups : continuous flow compressors (i .e . turbomachines) and intermittent flowcompressors (e .g. reciprocating compressors) . Various members of these two groups arepresented in Fig. 2.1.1.From the group of turbomachines the centrifugal compressor is the only type that can operate atdesign pressure ratio of 10:1 under extreme conditions . However, a preliminary designcalculation showed that this compressor type is not suitable as on-board application. Assumingan isentropic efficiency of -85%, the rotational speed would exceed 100,000 rad/sec . Therequired rotor tip diameter is less than 2 cm which leads to a rotor tip speed of more than 500m/s1 . Most important is the fact that the rotor dimensions would be too small to machinesufficiently small clearance volumes.I A sample calculation is attached in APPENDIX A320Figure. 2 .1 .1 : Classification of compressor typesHence, intermittent flow compressors are of primary interest which can be subdivided in twogroups, namely rotary and reciprocating compressors . Rotary compressors have been excludedfrom the list of alternatives because their discharge pressure ratings are considerably lower than200 bar and the maximum stage pressure ratio is 4 .5 : 1 . In addition, rotary compressor typestypically consume up to 25% more power per unit flow rate than reciprocators [3].In general, the problem with rotary compressors is poor sealing of the sliding surfaces whichbecomes increasingly difficult with higher pressures . Vane type compressors are limited by themaximum allowable bending stress in the sliding vanes which is due to the pressure difference inthe neighbouring compression chambers . To reduce the stress, more vanes are required whichrenders the design more complicated but does not solve the sealing problem.COMPRESSORSl	~INTERMITTENT FLOW	CONTINUOUS FLOWIROTARY	RECIPROCATING	DYNAMIC	EJECTORSLIDING VANELIQUID PISTONSTRAIGHT LOBEHELICAL LOBECOMPRESSORACTUATIONHYDRAULICPNEUMATICMECHANICALCENTRIFUGALAXIAL FLOWMIXED FLOW21Rotary compressors are usually designed for a specific operation condition . Capacity controland off-design performance is not common and difficult to implement . Therefore, the focus ison reciprocating piston compressors.2.2 Reciprocating CompressorsIn this study the reciprocating compressors are grouped according to their actuation type . Threemain types of actuation have been investigated : hydraulic, pneumatic and mechanical actuation.The names of the groups refer to the medium that is used to generate the reciprocating motion ofthe piston.2.2.1 Characteristics of the Compression Cvcle,The operation of a reciprocating compressor is based on the principle of volume reduction via amoving piston in a cylinder which starts at bottom dead center (BDC) (Fig . 2 .2.1 A).This volume reduction causes a pressure rise of the gas enclosed in the cylinder (Fig . 2 .2.1 B),where :	(P2	yl	pl	v 2(2.2.1)kSubscripts 1 and 2 denote inlet and cylinder conditions, respectively and k is the isentropiccoefficient of the compressed gas.When the cylinder pressure exceeds the discharge (or receiver) pressure, the discharge valveopens and gas begins to flow out of the cylinder at discharge pressure until the piston is at topdead center (TDC) (Fig 2.2.1 C). The discharge valve closes at this point . Residual compressedgas is remains in the clearance spaces and expands as the piston begins to move in the oppositedirection .22A	IRQM. .RlwnTypical nonwater-cooled cast iron cylinder with short liner.DBCOMPRESSIONEXPANSIONEINTAKECDISCHARGEFigure 2 .2.1 : Typical reciprocating compressor cycle (Source : [3])23The clearance spaces are mainly due to the valves and due to a clearance distance between thepiston at TDC and the end wall of the cylinder which can be considered a safety margin to allowfor heat expansion of the rod.During the first part of the expansion stroke (Fig 2.2.1 D) both inlet and discharge valve remainclosed and the residual gas expands until the cylinder pressure drops below inlet pressure . Theexpansion is described asPz	—V	(2.2.2)P1 V2where subscripts 1 and 2 denote inlet and cylinder conditions, respectively and kex is thecoefficient of expansion. When the cylinder pressure is equal to the inlet pressure, the inletvalve opens and fresh gas enters the cylinder at inlet pressure until the piston reaches BDCposition (Fig.2.2.1 E). At BDC, a full compression cycle is completed.Assuming isentropic conditions, the compression coefficient (k) is equal to the expansioncoefficient . However, if heat transfer is considered, both the compression and the expansioncoefficient (ncomp and neXp ) vary during the stroke . The compression and the expansion arenow polytropic . The relationship between isentropic and polytropic coefficient is dependent onthe heat transfer and can be derived from the Second Law of ThermodynamicsT•ds=dh-v•dp,and the Ideal Gas Law for polytropic compression and expansion(2.23)dp+n.dv =0.P	VEqn. 2.23 can be modified and rewritten as follows :(2 .2 .4)dsce dT dpR RT	p(2.2.5)ds _ k	dT _ dp (2.2 .6)R k-1	T	p 'P24ds=k dp + dv _ dpR k-i p v	p 'ds-=k dv+1 dpR k-1 v k-1 pModifying Eqn. 2.2.8 gives(k -1) • ds	dv dpk•+—.R	v pHence, the Second Law for isentropic compression and expansion isds-= k dv + dpcv	v pand with Eqn. 2.2.40=n.dv + dpv pthe relationship between n and k can be stated as follows:ds-= (k - n) dv .	(2.2.12)cv	vEqn. 2.2.12 can be applied to compression and expansion of natural gas for both heating andcooling of the gas . The result is summarized in Table 2.1.dv/v ds/c, (k-n)Compression (HEATING) - + - n > kCompression (COOLING) - - + n < kExpansion (COOLING) + - - n > kExpansion (HEATING) + + + n < kTable 2 .1: Relation between k and n .(2.2 .7)(2.2.8)(2.2.9)(2.2.10)(2.2 .1 1)25INLET	DISCHARGE	INLET	DISCHARGE11R~R11SINGLE ACTING DOUBLE ACTINGDISCHARGE11Rf•R11Figure 2.2.2: Reciprocating compressor modesThe compression cycle has been described for a single acting compressor type, where the gas iscompressed only during the forward stroke.In a double acting compressor type, gas is both discharged and drawn into the compressionchamber simultaneously during forward and return stroke (Fig . 2.2.2). It has almost twice thecapacity of a single acting compressor, given the same dimensions . However, an effective rodseal is required to minimize gas leakage toDISCHARGE PRESSUREp2	the atmosphere . One of the majoradvantages of the double acting mode isthe fact that the compression stroke issupported by the incoming gas of theTMOSPHERICPRESSURE	INLETPRESSURE	opposite compression chamber . Note fromSHAFT	p iFORCE	Fig 2.2.3 that the shaft force isconsiderably reduced as it is a function ofthe pressure difference (Op= p2 - pi) ofFigure 2 .2.3 : Single acting versus double actingcompressor mode	the two compression chambersSINGLEACTINGDOUBLEACTING262.2.2 Volumetric EfficiencyThe volumetric efficiency (rl v)is defined as the ratio of measured mass flow to the theoreticalmass flow given by the compressor displacement and inlet condition of the gas.qv	 measured	(2.2.13)pbm •n•DD is the compressor displacement per stroke, n the compressor speed in and p the gas densityat inlet condition.The actual mass flow rate is lower than the theoretical mass flow because of residual gas beingtrapped in the clearance spaces . The clearance ratio so is defined as the clearance vooume of thecylinder with the piston at TDC devided by the piston displacement per stroke. The clearancespaces are due to following factors:• clearance pockets due to inlet and discharge valves•	stroke clearance between the piston at TDC (or BDC for double acting pistons) andthe top end (or bottom end) of the cylinder• crevice volumes between the piston and and the cylinder boreClearance volumes from 4% to 20% are common for reciprocating piston compressors . Thevolumetric efficiency can be calculated as a function of the pressure ratio if so is known.fly =1—soP-1	(2.2.14)The volumetric efficiency is very sensitive to changes in clearance ratio s o and pressure ratio asis shown in Fig. 2.2.1, where iv is plotted for three different eo as a function of the pressureratio.Besides the clearance volume and pressure ratio, the following factors also effect the volumetricefficiency:27• poor filling of the cylinder:The gas pressure drops acrross the inlet valve and the compression starts at alower pressure level . Thus, the pressure ratio is higher.• preheating of incoming gas:The fresh, incoming gas is heated and expands in the compression chamber, thus,reducing the amount of fresh, incoming gas . The temperature factor can reducethe volumetric efficiency by as much as 10% for a 5 :1 pressure ratio [2].• gas leakage through discharge valve:Insufficient sealing of the discharge valve allows compressed gas to flow backinto the cylinder at the beginning of the expansion stroke.• leakage across the piston seals• leakage across the rod seal• thermal losses and cooling:During compression and expansion of the gas heat is transferred both to and fromthe gas . As a result, the specific heat ratio changes constantly . The mean valuemust be determined experimentally . It is a compressor specific parameter.The effect of these factors cannot be determined accurately but must be represented in thevolumetric efficiency as a correction factor . This correction factor appears from early testexperience to be about :=Ps.0.01P 1(2.2.15)eon.This changes Eqn 2 .2.14 to1	ri g =1—eo• P2	-1	0.01	P1	P 1(2.2.16)28A correction factor can also be applied to account for changes in the compressibility . Since theintensifier operates over a large range of pressure ratios and temperatures, it is difficult to obtaina clear picture of the effect of the compressibility factor . Hence, the compressibility is neglectedin the calculation of the volumetric efficiency.Figure 2 .2.4 shows that the volumetric efficiency greatly influences the capacity of thecompressor and that its impact becomes more significant at high pressure ratios.PZCLEARANCE60¢ 40DU)utQ. 20PlU2QWJUACTUAL CAPACITY-.PISTON DISPLACEMENT--+7%14%,21%~KVE4% 6 EV-iw76 .4;B VEB .t%0U4 100121 107PERCENT PISTON DISPLACEMENT80 60 40 20 OVOLUMETRIC EFFICIENCYDISCHARGE PRESSURE 206 bar, n= 1 .3[bar]25020000 102	3	4	5	6	7	8	9PRESSURE RATIOTANK PRESSURE [bar]	5% CLEARANCE VOLUME10% CLEARANCE VOLUME	15% CLEARANCE VOLUMEFigure 2.2.4: Impact of Clearance Volume on Volumetric Efficiency.29The effects of volumetric efficiency on the compressor capacity render multistaging moreattractive, since the stage pressure ratio is (ideally) reduced to the root of the total pressure ratiostage — n 7r-C---total (2.2.17)where n denotes the number of stages . A 10:1 pressure ratio would therefore be reduced to astage pressure ratio of -3 .2 :1 . The volumetric efficiency of a multistage compressor is equal tothe volumetric efficiency of each of its stages . Thus, it is higher than in a single-stagecompressor.2.2.3Methods of Cauacity Control,The intensifier capacity is a function of three variables, namely, the engine speed, the load andthe current tank pressure . While engine speed and load vary constantly, the tank pressure (and,thus, the pressure ratio) changes slowly over time (ref. Fig . 1.4.1). Ideally, the capacity shouldbe varied in infinitely small steps (i .e . variable displacement) to match the intensifier capacitywith the current fuel requirement . However, most compressors are not variable displacementunits and since an on-board intensifier is driven at engine speed, it must be thought of alternativeways to control the intensifier capacity.Four methods of controlling the capacity of a reciprocating piston compressor have beenconsidered : speed control of the compressor driver, control of clearance volume, control of gasinlet and control of gas discharge . The choice depends strongly on the driver configuration . Itis also a matter of complexity, cost and efficiency. Speed Control of the Compressor DriverD ON/OFF-mode:The compressor is turned on or off depending on the current fuel requirement . An upper and alower pressure limit p, . and Amin can be preset to initiate the switching process . The30maximum allowable pressure difference (0 p = p,-pmin) depends on the accuracy of thecontrol mechanism and on how sensitive the injection and the combustion of natural gas are tochanges in the gas pressure . The system requires an electronic control device and a clutch toseparate the compressor from its driver . From an energy point of view the ON/OFF mode is thebest control, since it consumes no energy when it is not in operation.ii) Variable Compressor Speed :,The capacity of a positive displacement compressor is directly proportional to its speed.Therefore, the capacity can be controlled by controlling the speed . The compressor must beequipped with a control unit to change the compressor speed either linearly or stepwise (e .g.clutch or variable speed transmission), according to the current fuel requirement.iii) Variable Displacement Drive:The reciprocating motion for a positive displacement compressor can be generated by a fluid-activated cylinder (i .e. air- or hydraulic cylinder) . The compressor capacity can be controlled bycontrolling the amount of actuating fluid (i .e . air or hydraulic fluid) which determines thefrequency of the actuating piston and, therefore, the intensifier frequency . Variabledisplacement pumps can adjust the fluid flow regardless of its drive speed . Of major advantageis the fact that the fluid capacity can be varied regardless of changes in engine speed . Thissystem requires a load sensing unit which determines the hydraulic or air line pressure andchanges the fluid capacity accordingly . Variable displacement systems are ideal for anintensifier application ; however, they can only be employed in connection with a fluid driver.312.2.3.2 Control of Clearance VolumeFrom Eqn. 2.2.16 it follows that the capacity of a reciprocating piston compressor is primarilydependent on clearance volume and pressure ratio . The clearance volume can be increasedlinearly or stepwise using add-on clearance volume, which results in lower volumetric efficiencyand, therefore, lower compressor capacity (Fig 2 .2.5).Zero capacity is obtained when the cylinder pressure after the gas expansion at BDC is equal tothe inlet pressure. Additional cooling is required as increased heat generation is expected . Anelectronic control system regulates the timing and the amount of clearance volume added on.FULL LOAD	3/4 LOAD	Y2 LOAD% LOAD	NO LOADFigure 2 .2 .5 : The effect of add-on clearance spaces2.2.3.3 Control of Gas InletThe gas flow is regulated by delaying the closing of the inlet valve . During the compressionstroke a portion of the fresh gas in the cylinder is discharged through the inlet valve, thusreducing the amount of gas in the cylinder. After a controllable time interval the inlet valvecloses and the compression begins . This method is commonly known as "valve unloading" . Theeffect of "unloading" for infinite-step operation can be seen in Fig . 2 .2.6 .DISCHARGEWWS.FULL LOAD -► / / /DASHEDLINESD/	/ / /A,/ NO LOADINTAKE32-4	VOLUME ----!Figure 2.2.6 : Progressive indicator card illustrating the operation of infintite-step controlThis system can be used in combination with an overflow line from the upper part to the lowerpart of a double acting cylinder . The gas would then be moved between the two cylinder parts ifzero capacity is required.ARE PART2.2.3.4 Control of Gas DischargeThis method is based on a return (or "bypass") line which leads surplus compressed gas back tothe inlet port of the compressor . If the discharge pressure exceeds the required injection pressureby a preset margin, a valve opens the bypass line and the compressed gas reenters the intensifier.With the bypass valve open, the gas flows both through the intensifier and to the injector untilthe pressure decreases below a preset minimum pressure (the gas inlet line must be secured witha check valve to avoid that the compressed gas expands into the tanks) . Then the bypass valve(electronically or pilot pressure operated) closes and gas coming from the tank is compressedagain . This method is later referred to as "bypassing" .332.3 Hydraulic Actuation2.3.1 General,Hydraulic actuation means that hydraulic fluid is used to generate a reciprocating drive for theintensifier. A hydraulic drive circuit consists of four main parts:1.) an element to pressurize the fluid (e .g. hydraulic pump)2.) a double acting hydraulic cylinder3.) a switch valve to control the double acting cylinder4.) a flow control unit to adjust the driver frequency.In the early design phase, it has been considered to convert the existing parts of an engine brakesystem to a hydraulic driver' . The modification of the engine brake system has been turneddown because the dimensions of the hydraulic parts are too small to generate the required fluidcapacity and because of a lack of fluid capacity control.2,3,2 Circuit ArrangementFigure 2 .3 .1 shows a hydraulic driver configuration which consists of a hydraulic and a naturalgas circuit. The natural gas circuit contains a double-acting cylinder where the gas iscompressed (intensified) and a high pressure gas accumulator.In the hydraulic circuit fluid is drawn from a reservoir and pressurized by a variabledisplacement pump which is connected to the crankshaft of the diesel engine and operates atengine speed . It is then forwarded to a double-acting hydraulic cylinder where a pilot-pressureoperated switch valve controls the forward and return motion of the piston.'Engine brakes use hydraulic fluid to seal the exhaust valves of the diesel engine to compress the combustion air.The fuel injection is skipped and the compressed air is expanded to atmosphere. Hence, the engine consumsenergy rather than producing it . The engine brakes are not used in urban bus engines because they generate toomuch noise .34DOUBLE-ACTINGHYDRAULIC CYLINDERCNG-TANKCNG INTENSIFIERIVARIABLEDISPLACEMENTHYDRAULIC PUMP6V-92 TAHYDRALUICRESERVOIR L —MICROPROCESSOR1CNG TOINJECTORVFigure 2.3.1 : Hydraulic driver configuration.The piston frequency can be adjusted in infinitesimal steps according to the CNG fuelrequirement of the diesel engine by altering the swashplate angle of the variable displacementhydraulic pump . Hence, the pump maintains the required capacity regardless of its drivingspeed. An electronic control system senses the CNG discharge pressure signal and initiates astepper motor which readjusts the swashplate angle . If the measured discharge pressure exceedsor decreases below a preset pressure range (e .g. ±5%), the stepper motor responds by readjustingthe swashplate angle to reduce or increase the piston frequency, accordingly .352.3.3 Design SpecificationsTo design a hydraulic intensifier driver the dimensions of the CNG cylinder must be calculatedfirst, according to the maximum CNG fuel requirement of the diesel engine . The selection ofcylinder bore and stroke length is a compromise between large thrust forces and high pistonfrequencies . From previous experience with hydraulic equipment the maximum pistonfrequency should be limited to 30 cycles per minute (1 cycle = forward + return stroke) . Delaytime due to switching and acceleration and decceleration of the piston must be taken intoaccount. The maximum allowable piston speed for hydraulic cylinders is 1 .524 m/sec (300ft/min).Hydraulic systems are available at operating pressures between 100 and 400 bar (1500 and 6000psi) . The bore of the hydraulic cylinder can be calculated when the system pressure and thedimensions of the CNG cylinder are known.ROD DIAMETERFigure 2 .3.2: Piston rod - stroke selection chart2 [4]2 Here, the basic displacement is defined as the distance between hydraulic piston and intensifier piston . It isapproximately three times the stroke length.loo	1	s	s	s 1000	s	s s 7 s s lo,000TNp11ST—POUNDS36The selection of the rod diameter depends on the cylinder bore and on the maximum thrustforce. Figure 2.3.2.has been taken from a hydraulic cylinder manual . It shows the rod diameteras a function of thrust force and rod length. The chart has been used to roughly estimate the rodsize for an iterative design procedure . Buckling forces and stresses in the rod must be calculatedafter an adequate rod diameter has been selected.A preliminary design calculation has been carried out 3 , based on the maximum CNG fuelrequirement and on the design guidelines stated above . According to this calculation thedimensions of the hydraulic cylinder are as follows:Bore :	7 .62 cm	(3")Stroke:	30.48 cm	(12")Rod:	5.08 cm	(2")Displacement :	V	= 2164.2 cm3/cycleHYDRAULIC PUMP DIMENSIONSSTROKE : 30 .48 cm (12" ), ROD : 5 .08 cm (2")80 500704006050 300402003020100100 010.16 8 .89	8 .255	7.62HYDRAULIC CYLINDER BORE (cm)-PUMP SrLE	.0 OPERATING PRESSURE9 .525 6 .985 6.35Figure 23 .3 : Hydraulic cylinder bores versus operating pressure and correspondingpump size.3see Appendix A637The maximum piston frequency is 35 cycles per minute and the maximum thrust force is 86 .14kN.The selection of the hydraulic pump displacement depends on the hydraulic cylinder dimensions,on the maximum piston frequency and on the rotational pump speed . Given the displacement ofthe cylinder and the maximum frequency, the pump displacement is a function of operatingpressure and cylinder bore. The results are summarized in Fig. 2.3.3.A possible hydraulic circuit arrangement is presented in Fig . 2 .3.4. Note that the hydraulic fluidflow is controlled with a load sensing unit which maintains a preset pump capacity regardless ofchanges in drive speed or operating pressure . The load sensing unit consists of a proportionalvalve and a turbine flow meter.The following components are required:•	a pressure-compensated, load-sensing, variable displacement pump with remotepressure compensator control• a closed-loop electronic proportional valve to control the pump capacity:• a proportional-drive card with gain, deadband compensationand ramping controls• a turbine flow meter with signal conditioner• a pilot-pressure operated directional control valve with operatorto cycle hydraulic cylinder continuously• a hydraulic reservoir and an in-line return filterThe cost for the hydraulic driver system as shown in Fig . 2.3.4 is estimated at $ 88764 .4Price by Janox Fluid Power as of March 1991 . Note that the CNG cylinder is not included .382.3.4SummaryHydraulic actuation has been considered because hydraulic technology is readily available and awide variety of features and components can be selected .	CHYDRAULIC CYLINDERACCUMULATOR(double acting)IFIERINTEN!xSWITCH VALVE CIRCUITPROPORTIONALVALVETURBINEFLOW METERCRANKSHAFT6V 92 TA	(VARIABLE DISPLACEMENT)Figure 2.3.4: Hydraulic circuit arrangement .39The following are the advantages of hydraulic actuation:• The variable displacement pump provides a number of control features to regulate theintensifier capacity.• Hydraulic parts are interchangeable• Hydraulic systems operate efficiently• The hydraulic pump can be mounted directly to the crankshaft outlet of the dieselengine.The disadvantages are:• The system is heavy• A hydraulic reservoir is required which consumes space• The system is expensive2.4 Pneumatic Actuation2.4.1 GeneralThe concept of pneumatic actuation is similar to the hydraulic driver, except it uses compressedair as working fluid . Pneumatic actuation has been considered because an air system is on boardthe bus to operate brakes and pneumatic switches.The purpose of this section is to investigate the existing air system and to consider alternativeways of upgrading the system to power an intensifier driver . Furthermore, alternative air-drivenintensifier concepts are evaluated.2.4.2 Circuit ArrangementThe reciprocating motion of the intensifier is generated by a double acting air cylinder.Compressed air at 10 bar is supplied by an on board air source .40After the working stroke the air in the cylinder is expanded to the atmosphere.DOUBLE-ACTINGAIR CYLINDERCNG INTENSIFIERCNG-TANKSWITCHVALVECOMPRESSEDAIR SOURCECNG TOINJECTORFigure 2.4.1: Air-driven intensifier arrangement.The piston frequency is controlled according to the CNG fuel requirement by adjusting the airflow. Note from Fig . 2.4.1 that the bore of the air cylinder is considerably larger than the boreof the CNG cylinder which is due to the pressure difference of the two gases . The CNG circuitis similar as discribed in the previous section (see Fig. 2.3.1).2.4.3 Alternative Air SourcesAll transit and intercity busses that are powered by a Detroit Diesel engine are equipped with a700 Series Bendix air compressor s. This compressor runs at engine speed and delivers air at amaximum pressure of 9 bar (125 psi) . Its capacity is rated 0.43896 Nm3/hr (i .e . 15 .5 SCFM) at1250 engine rpm which turns out to be a maximum capacity of 0 .73632 Nm3/hr (i .e . 26 SCFM)at full speed . A pressure regulator controls the compressor capacity. For safety reasons, the5Compressor specifications are attached in Appendix A741compressed air is stored in a reservoir . Comparing the compressor capacity of the existing airsystem (26 SCFM @ 2100 rpm) with the required CNG capacity (35 SCFM @ 2100 rpm) itmust be concluded that the air system needs to be upgraded to supply an adequate amount ofcompressed air to the air driver . The following approaches have been considered to increase theair capacity.• Upgrading of the Existing System:The existing air compressor can be replaced by a larger compressor model of the same brand.The compressor capacity can be further increased by choosing a gear ratio such that thecompressor operates at maximum speed (which is in many cases higher than 2100 rpm). Asecond air compressor can be mounted to the available power take-off of the diesel engine, if thecapacity is not sufficient . With the above mentioned options the air capacity can be increased by300%.• Storage of Compressed Air:Compressed air can be stored in in tanks at 400 bar on board the bus . The tanks supplycompressed air to a double acting air cylinder which powers the intensifier. The air tanks can befilled in refueling stations or on key points of the bus route depending on the capacity of thestorage tanks . Multifuelling air compressors are capable of filling the air tanks in a few minutes.However, a modification of the infrastructure is required which is not aimed for in this study.Another disadvantage is the space requirement of the storage tanks which contributes to theexisting space problem on board the bus.• The "Skipfire" Method:The "Skipfire" method is based on the compression of air in one of the combustion cylinders ofthe diesel engine while skipping the fuel injection and the combustion . The process can be42electronically controlled using a modified engine brake system (ref. Chapter 2 .3) . The operationof the specified cylinder alternates between air compressor and combustion engine.This method is associated with considerable power pensities due to skipping the combustion . Asample calculation (attached in Appendix A8) shows that the power output of the diesel enginecan be reduced to 83% under certain conditions 6. Other disadvantages include the hightemperature of the compressed air due to the single stage compression and the heat transfer fromthe hot cylinder walls and the fact that a modification of the combustion cylinder is required.2.4.3 HASKEL Gas Booster ,A Haskel gas booster is a commercially available high pressure gas intensifier which is designedfor industrial gases that are required at high pressures . This plunger-type reciprocatingcompressor is powered by compressed air (8 - 13 bar) and covers a wide range of pressure ratiosand discharge pressures.Haskel gas boosters consist of a large reciprocating air-drive piston which is directly coupled toa small bore gas plunger as shown in Fig . 2 .4.2. The gas piston intensifies gas in a high-pressurebarrel section which is equipped with high pressure inlet and discharge check valves. The airdrive section includes a cycling spool and pilot valves that provide continuous reciprocatingaction when air is supplied to the air drive inlet . Dynamic seals isolate the gas compressionchamber from the air drive section . Cooling is provided by routing the cold exhaust air throughan individual jacket surrounding the gas barrel . Two-stage models are equipped with anintercooler which also uses cold exhaust air as cooling medium.6The calculation is based on 50 SCFM air flow at 1250 engine rpm and 125 psi air pressure . Thus, 98% of thecombustion cycles of one cylinder must be skipped to achieve the required air flow .43PILOT VALVEPoppet Design, All Stain-less Steel with MoldedBuna-N Seel.UPPER & LOWER CAPSHigh Strength AnodizedAluminum.AIR INLET PORTAIR CYCLING VALVELow inertia Alumi-num spool withinStainless Steel Sleeve.Air Pilot shifted —no wrings. Minimumfriction, yet no air leakage.AIR EXHAUST TUBE~~AIR PISTONHigh Strength Anod izedAluminum . Unique mini-mum-friction DynamicSeal.AIR BARRELFilament Wound Fiber-glass Reinforced PlasticFILTERED BREATHER TOCHAMBER BEHIND GASPISTON.NOTE : May be piped toremote area ifpumping flamableor toxic gases.CHECK VALVESSpring Loaded Poppets.Seats optically lapped.GAS END CAP300 Series Stainless Steel(Except Lower PressureModels — refer to pages 11 & 121.Figure 2 .4.2 : Haskel gas booster (Source: Haskel brochure M-26D) .44The frequency of the reciprocating actuation is dependent on the gas load (pressure ratio andrequired end pressure of the compressed gas) but it is primarily determined by the pressure andcapacity of the available air source . The more air available, and the higher the pressure, thehigher the operating frequency of the booster.A list of all commercially available Haskel gas boosters is attached in Appendix A9 as well as aset of performance calculations . Three groups of compressor types are available:• single-acting, single-stage (AG)• double-acting, single-stage (AGD)• two-stage (AGT)The performance calculations (attached in Appendix A9) have been applied to various AGD andAGT models (i .e . AGD-15, AGD-30, AGD-62, AGT-15/30) based on the required CNGdischarge pressure . An air supply of 75 SCFM at 9 bar was assumed . The capacity of the singlestage models AGD-30 and AGD-62 is 3 .77 SCFM and 3 .33 SCFM, respectively . The capacityof the two-stage model AGT-15/30 is 5.3 SCFM out of a required 35 SCFM.Upon a request to Haskel Inc ., a compressor package was proposed which consists of twodifferent models : a double acting, single stage booster (AGD-5) and a two stage booster (AGT-14/30). At tank pressures higher than 48 bar only the two stage model operates. If the tankpressuer drops below 48 bar, the single stage booster is automatically switched in series with thetwo stage booster as an additional compression stage' . The total air requirement for thisconfiguration is estimated at 356 SCFM . The total cost of both units is $18,953 .00 (not includedare a flow control system and an air supply)7 Specifications can be found in Appendix A9452.4.5 Alternative Air Driven Intensifier Desi gnsThe sizing of the air cylinder depends on the required thrust force for the CNG cylinder . Adesign procedure similar as presented in Appendix A6 can be applied to the pneumatic driver.Note, that the piston frequency and, thus, the intensifier capacity is a function of the availableamount of compressed air . When the dimensions of the pneumatic cylinder are fixed, the airsystem can be upgraded accordingly.2.4,6 SummaryPneumatic actuation was considered because of the availability of compressed air on board thebus. The existing air system is inadequate for an air driver and needs upgrading . The followingare the advantages of pneumatic actuation:• Pneumatic equipment is commercially available• A pneumatic driver is a variable displacement unit which can be controlled with the air• An air-driven intensifier system is commercially availableThe disadvantages are:• The compression of air is less efficient than the compression of hydraulic fluid• Compressed air expanded to the atmosphere can cause excessive noise• The existing air system needs to be upgraded• Commercially available gas boosters are extremely expensive46.5 Mechanical Actuation2.5.1 GeneralMechanical actuation is the conversion of a rotating motion into a reciprocating motion usingsolid material . Three types of mechanical actuation are investigated:• camshaft actuation• crankshaft actuation• rotating reciprocating pistons2.5.2Camshaft ActuationAll diesel injectors of the Detroit Diesel 6V-92 engine as well as the exhaust valves are actuatedby an overhead camshaft . At the beginning of the research on the conversion of the engine toCNG direct-injection, it has been considered to use the lift of the camshaft lobes to intensify thenatural gas . The idea was to modify the injector such that a small cylindrical chamberaccomodates natural gas at current tank pressure and a plunger (activated by the camshaft) bothcompresses the gas and forwards it directly into the combustion chamber . As the plunger movesback to its original position, fresh gas enters the injector chambercompleting the compressorcycle.The main advantage of this system is the fact that no compressor is needed to intensify the CNG.The forces on the plunger due to the gas compression are lower than for the diesel injection and,thus, do not create a problem.However, the major problem associated with this system is the lack of an effective flow controlmethod. The displacement of a camshaft actuated intensifier is a function of engine speed only,since the cam lift is constant .47Figure 2 .5.1 : Camshaft actuated plungerIn addition, the diameter of the original injector body is too small to accomodate a sufficientamount of gas for the combustion . A sample calculation (attached in Appendix A 10) shows thatfor a cam lift of 0.8255 cm (0 .325") a chamber diameter of at least 2 .8 cm is required to supplyenough gas for the combustion . The outside diameter of the injector is 2 .5 cm. The calculationis based on 100% volumetric efficiency for the gas compression . Hence, it is shown that theavailable space in the injector is not sufficient.In general, the main problem with camshaft-actuated piston compressors is to keep the clearancevolume small which is difficult to achieve because of the valve clearance and the typically smallcylinder dimensions . Small cylinder dimensions limit the capacity of the compressor.Therefore, a number of parallel operating plungers are necessary to achieve the required mass48flow rate which renders the system more complex . Therefore, the idea of camshaft actuationdoes not appear to be feasable.2.5.3 Crankshaft ActuationCrankshaft-actuated compressors are the most common among reciprocating compressors . Twogroups can be distinguished in terms of transmission of the rotating into the reciprocatingmotion :1) Direct transmission:The compressor piston is directly coupled with the crankshaft via connecting rod similarto a combustion engine . The piston both generates the reciprocating motion andcompresses the gas . This configuration allows single acting compression only.2) Crosshead transmission:A piston-like crosshead is guided in a cylinder to generate the reciprocating motion . Thecompressor is mounted on top of the crank case and the compressor piston is attached tothe crosshead with a rod . This configuration allows both single- and double-actingcompressor mode.It has been investigated whether a small combustion engine (representing the group of directtransmission) can be modified to become a CNG intensifier . The modification would includereplacement of the cylinder head by a compressor flange (which incorporates compressor valves)and replacement of the engine piston by a compressor piston with a lower radial clearance andhigh pressure seals . Ideally, the compressor would run at engine speed.Based on a 10:1 pressure ratio and 1 :1 speed ratio, a design calculation has been carried out (ref.Appendix A 11) to determine the required displacement per revolution . Thus, a displacement of50 cm3 per stroke is sufficient to meet the maximum CNG mass requirement assuming 10 %clearance volume.49Figure 2.5.2 : direct transmission (single acting) versus crosshead transmission (doubleacting)However, the sample calculation in Appendix A 11 shows that thrust forces of at least 5300pounds are expected due to compression which is clearly in excess of the maximum design forcefor a 50 ccm combustion engine . Therefore, the existing parts such as bearings, crank shaft andconnecting rod must be replaced . In addition, a flow control system is required.The main difference between a crosshead transmission and a direct transmission is theconversion of the crank case piston into a crossshead . Compared to a regular piston, thecrosshead has a significantly smaller radial clearance and oil grooves instead of piston rings.The main purpose of a crosshead is to carry the side loads and to eliminate radial misalignment.Crossheads are usually made of alumminum alloy or white metal which minimizes friction, yet,50is sufficiently strong to carry the side loads . A double-acting, crosshead-driven compressorrequires rod seals.The following are the advantages of using crankshaft actuation:• simple• consists of only a few components• inexpensive• crosshead transmission allows a double acting mode and a wide range of intensifierdimensions• small and lightThe following are the disadvantages:• operates at high speed which causes heat and excessive wear• needs a capacity control mechanism2,5,4 Rotating PistonsInspired by the function of a variable displacement hydraulic pump, it has been considered tomodify such a pump to a CNG intensifier or to adapt this system for a newly designedintensifier. All components of a hydraulic pump are designed for pressures up to 400 bar (5800psi) and the flow can be controlled in infinitely small steps . The pump capacity is a function ofthe angle of the swash plate which can be adjusted between an upper and a lower limit (i .e.between maximum and zero capacity) regardless of its rotational speed.A variable displacement pump is shown in Fig. 2.5.3. It consists of a swashblock (6) whichincorporates a rotating disk, called a swash plate . The swashplate can be inclined with respect tothe axis of the rotating shaft which changes the stroke of the axially arranged plungers . As theshaft rotates, the plungers move linearly in the cylinders, causing fluid to be drawn into the51cylinder and forwarded to the pump outlet (i .e . single acting compressor mode) . The plungerscan be equipped with gas seals, thus, allowing gas compression in the described mode.Figure 2.5.3: HYDURA PVWH open loop pump s. 1) Control system, 2) Cylinder mountedjournal bearings, 3) Swashblock lubrication, 4) Shaft, 5) Shaft bearing,6) Swashblock with bearing, 7) Plunger bearings, 8) Valve plate, 9) Valve plate port,10) Thru-shaft, 11) Quiet valve plate design, 12) FrameA sample calculation (attached in Appendix A 12) shows that the required pump displacement is8.036 ml/rev . The calculation is based on a 1 :1 speed ratio and a relative clearance volume of15% in each cylinder . The smallest pump size of this type is 10 nWrev which is suitable for aconversion.Unlike previously discussed piston compressors, the valves of this compressor type are radialslots, similar to those of a vane-type compressor . The cylinders move to and from the valves . Itis questionable whether efficient sealing can be achieved at the transition point from inlet tooutlet section, particularly when operating at high pressures and high pressure ratios . Gas, beingcompressible and less viscous than hydraulic fluid is expected to leak back into the inlet section1 " OILGEAR Controlled Power" Bulletin 47015B52of the compressor . In addition, efficient sealing of the sliding surfaces (i .e . valve plate (11) andshaft (10)) may cause a problem.The system looks very attractive because it provides an efficient capacity control mechanism, itis small in size, and it consists of only one element . The pump can be directly mounted to thecrank shaft of the engine and run at engine speed.However, the modifications include a number of unknowns, particularly in terms of sealing . Anewly designed valve plate must be considered as well as a lubrication device, since the pumpoperates at high speeds.2.5.5 SummaryThe following are the advantages of a mechanical driver:• simple design• small in size• light• low cost• more efficient than hydraulic and pneumatic driverThe disadvantages and expected problems are:• a flow control system is required in most cases• high piston speeds can cause excessive heat generation and wear2.6 Summary and ConclusionThe function of the alternative design concepts has been presented as well as the advantages anddisadvantages of their application as on-board intensifier . The purpose of this part is to evaluate53all models based on the criteria stated in Chapter 1 .4. In particular, the following points arediscussed :• Complexity of the system• Cost• Capacity control• Weight and space requirement• Power consumption and efficiency• Durability• Maintenance• NoiseThe sequence of the points listed above represents the priority of the design aspects.Both hydraulic and pneumatic driver systems are more complex than the mechanical driver.They consist of an intensifier cylinder and a separate circuit arrangement . Besides theintensifier, three basic parts are required, namely a device to compress the working medium (i .e.hydraulic pump or air compressor), a double-acting cylinder and a switch valve to control thereciprocating motion of the piston . Further, a control system is required. The mechanicalsystem consists of only two parts : a mechanical driver device (i .e. crank case) and a capacitycontrol system. The capacity control system must be chosen with respect to the intensifierdesign.The cost of an intensifier system is directly proportional to its complexity . It is difficult to giveexact prices on the different configurations because each design specification can vary in termsof size or quality reqiurement of equipment . A price breakdown has been attempted for thehydraulic configuration as well as the Haskel gas booster which allows to range the systemsamong each other. Thus, the pneumatic driver is the most expensive system (particularly withthe Haskel booster as intensifier) followed by the hydraulic system . The mechanical system isthe most inexpensive solution.54The ideal way of controlling the CNG flow is to use a variable displacement type compressorwhere the compressor capacity can be adjusted regardless of engine speed and gas inletcondition. Both hydraulic and pneumatic driver are variable displacement units as they vary theflow volume of the transmitting medium (i .e . hydraulic fluid or air) to adjust the intensifierfrequency. The hydraulic driver provides the best flow control device with the pump capacitybeing variably adjustable by infinitly altering the swashplate angle . The capacity of the pump isindependent of changes in driving speed.Both air and CNG flow must be regulated using the one of the flow control methods that havebeen discussed in Chapter 2 .2 .3. Most of these methods are associated with power losses andrequire a more complex system. The flow control of the pneumatic drive is easier to handlebecause of the lower pressure level of the compressed air.The hydraulic system is the heaviest among the alternatives . Besides the circuit components, thehydraulic system requires a large hydraulic reservoir (20 - 40 liters) . It is also the largest systemfollowed by the pneumatic actuator and the mechanical actuator.The discussion of the efficiency of an intensifier driver does not include the power losses of theintensifier, itself, but is only concerned with the power consumption of the driver based on themaximum power consumption due to gas compression.A typical performance curve of a hydraulic pump is shown in Fig. 2.6.1 . It is the curve of themodel which was used for the hydraulic driver in the sample calculation in Appendix A6 (i .e.Hydura PVWH-11, capacity : 41.31/min @1800). The overall efficiency of the pump dependson the operating pressure. The operating pressure of the hydraulic pump is a function of the tankpressure, since the differential gas pressure in the double acting cylinder determines the total55thrust force . It can be concluded that the hydraulicpump operates at overall efficiencies of 80% andhigher, depending on the current CNG inlet pressure.The efficiency of an air driver remains constant,since it operates at a constant pressure. Changes indrive speed may cause lower volumetric efficienciesbut do not significantly influence the overallefficiency. Air compressors (given atmospheric inletpressure and 8 bar discharge pressure) usuallyoperate at efficiencies between 80 and 90%.Compared to the hydraulic driver, its volume flow is considerably higher in order to generate anadequate piston frequency . In summary it can be said that the efficiency of an air driver (nottaking into account losses due to friction in the pipes and valves) is constant between 80 and90%, regardless of changes in pressure ratio and flow volume.The efficiency of mechanical systems (such as a crank case) is primarily influenced by friction,particularly, by friction due to high piston speeds and/or misalignment . The higher the pressureload, the more significant are the friction losses . The minimum efficiency limit is 90%.Depending on the quality of the parts the efficiency may be as high as 97% . The directtransmission compressor (i .e . conversion of a 50 cm3 motorcycle engine) is a special casebecause compressor driver and compressor are identical . Therefore, this driver can be rated at100% efficiency.In general, the mechanical driver provides the best efficiency for all load cases . Its behaviourchanges insignificantly for different loads (i .e. pressure ratios) . Every mechanical systemconsists of less parts than hydraulic and pneumatic driver systems .Figure 2.6.1: Hydraulic PumpEfficiency Chart56Power penalties due to losses in transmissions are not taken into account, because eachconfiguration can be equipped with the same transmission . Therefore, the losses apply to everysystem at same extent.The durability of the intensifier system is a function of the operating frequency . Intensifierswhich operate at high piston speeds are expected to deteriorate faster . Thus, hydraulic andpneumatic driver are considered more durable than the mechanical driver.The mechanical drive configuration, being less durable than the hydraulic and pneumaticsystems, requires the most maintenance work . Maintenance work means in particular valveservice and change of seals . Leakage and failure of components may cause increased service ofthe hydraulic system; however, the parts are interchangeable. The pneumatic system causes theleast problems in terms of maintenance, since an air system is part of the bus equipment andpiston speeds are very low.Noise, generated by the intensifier, is not a problem as long as it is less than the engine noise.The pneumatic actuation raises concerns about noise emissions, since its function is based on theexpansion of air to atmospheric pressure . Noise emissions of hydraulic and mechanicalactuation are subject to experimental evaluation but can be estimated lower than the pneumaticactuation.The results of the evaluation is summarized in Table 2.2 . Thus, the mechanical drive is superiorin most aspects . Particularly attractive is its simple design and the low cost involved . Thesefacts and the relatively high efficiency render it the most promising solution . The required flowcontrol system represents a major challenge for the mechanical configuration . However, anumber of control methods are available of which the most suitable can be selected . Poordurability might be expected due to high piston speed but is subject to experimental evaluation .57Power/Efficiency 80-90% 80-90% 90 - 97 %Flow Control variabledisplacementvariabledisplacementcontrol systemrequiredSystem Complexity complex complex simpleCost high high lowDurability high high lowWeight and Space large medium smallMaintenance Work low low highNoise low high lowTable 2.2 : Comparison of alternative drive systemsThe hydraulic system is very attractive because it is a variable displacement unit and, thus, morereliable than a mechanical actuation. The parts are durable and can be exposed to highpressures . Hydraulic technology as well as a broad selection of hydraulic components isavailable . Power penalties at low operating pressures reduce the efficiency of the system . Inaddition, the number of components increase both complexity and cost . The hydraulic systemranks very closely behind the mechanical system.The pneumatic actuation seems to be an excellent compromise between mechanical andhydraulic drives. It is a variable displacement unit and its function is similar to the hydraulicsystem with the exception that it uses compressed air . However, the capacity of the on-board airsystem is not sufficient to support an air driver for an intensifier . Considerable modifications arenecessary to increase the volume of compressed air, which means that the system becomes morecomplex and, as a result, more expensive (particularly the Haskel gas booster).58With regard to the evaluated design aspects discussed, it must be concluded that the mechanicaldrive configuraion seems to be the most promising solution . In particular, the double actingmode is favourable since it offers flexibility in selecting bore and stroke dimensions . Theconversion of a hydraulic pump implies too many questionmarks to present itself as attractivealternative . Hence, the double-acting, single-stage compressor has been chosen as intensifierprototype for an on-board bus application .59Chapter 3The Mechanical Drive Configuration3.1 Design DetailsThe graphs presented in this section illustrate the compression process, ideally observed undercertain conditions and should be regarded as design aid rather than strict design constraints.The primary purpose of the experimental tests is to investigate and understand the behaviour ofthe compression of CNG to high pressures and to solve the problems associated with it . Theacquired knowledge serves as a tool for the design of a second generation intensifier.3.1.1 Intensifier DimensionsThe sizing of the single-stage intensifier was carried out in 6 steps:Step 1: Calculation of the theoretical intensifier displacement Vth according tothe fuel requirement of the diesel engine.Step 2: Calculation of the basic intensifier displacement VDB, taking theestimated volumetric efficiency into account.Step 3: Calculation of the actual intensifier displacement VDA, after havingselected an appropriate speed reduction.Step 4: Selection of the stroke lengthStep 5 : Selection of the cylinder boreStep 6 : Selection of the rod diameter60While the calculations in Step 1 - 3 involve assumptions and decisions, Step 4-6 are carried outiteratively, which requires the change of previously selected parameters.Step 1 : The capacity requirement is based on the maximum CNG fuel consumption of thediesel engine at full load over the entire range of engine speeds (ref. Chapter 1 .3 .3) . It wasconcluded that the maximum CNG mass flow for the DDC 6V-92 TA bus engine is 460 mg perengine revolution. The equivalent displacement at 20 bar gas pressure (i .e . minimum density)and 288 K inlet temperature can be calculated fromVth_ m R•T (3.1 .1)and is460 .10-6 kg • 500.9	 J	 • 288K_	rev	kg K	.106 = 33.2 cm3/rev2 . 10 6 PaStep 2 : To obtain the required intensifier displacement (i .e . basic displacement VDB), thevolumetric efficiency must be taken into account . This was done by dividing the theoreticaldisplacement Vth by the volumetric efficiency at design pressure ratio.VDB = V (3.1 .2)rl,A displacement factor is defined as the inverse of the volumetric efficiency to illustrate theimpact of clearance volume on the intensifier dimensions.Figure 3.1.1 shows that the displacement factor is extremely sensitive to changes in clearancevolume. It is a design goal to minimize the intensifier dimensions, thus, it is important tominimize the clearance volume . The actual clearance volume due to valves and stroke clearanceis estimated between 3% and 10% .61EFFECT OF CLEARANCE VOLUMEPRESSURE RATIO 10:1, n = 1 .255	100/0 RELATIVE CLEARANCE VOLUME0 15•	1205040302010Volumetric Efficiency DisplacementFactor100908070>-0876Figure 3 .1.1 : Volumetric efficiency and displacement factor as a function of relativeclearance volumeAssuming 5% clearance volume and an isentropic coefficient of 1 .3 [6], a displacement factorFD of 1 .6 is calculated using Eqn. 3.13.VDB = 1.6 . 33.2 = 53.1 cm3/revStep 3: Knowing the basic displacement, an appropriate speed ratio must be selected . Giventhat the lifetime of both valves and seals suffer from high piston speeds, a speed reduction mustbe considered . Lower intensifier speeds likely result in higher efficiencies as has been discussedpreviously.Considering a speed reduction, the new displacement (VDA) is defined asVDA = r VDB	(3.1 .4)where r is the speed ratio (r = 3 for a speed reduction of 3 :1). Note that VDA' VDB.62Two means of speed reduction have been investigated : gear belt drive and gear box . The gearbelt drive is less expensive than a gear box and its horsepower rating exceeds 20 HP . Thetoothed gear belt prevents slip . The maximum speed ratio is limited by the maximum allowablecenter distance of the pulley wheels and by the maximum size of the larger pulley due to thelimited space available in the engine compartment.Figure 3.1.2 illustrates the relation between the intensifier displacement and its correspondingspeed reduction . The calculation is based on a 10:1 pressure ratio, 50 cm 3 piston displacementand 5% clearance volume . The choice of the intensifier speed is a compromise between highpiston speed and high speed reduction.Figure 3.1.2: Basic intensifier displacement versus intensifier speedThe intensifier speed should be kept as low as possible ; on the other hand, the speed ratio with agear belt drive should not exceed 4 :1 . This allows a selection of displacement between 100 and200 cm3 per intensifier revolution. The best compromise was considered to be the intercept ofEFFECT OF BASIC DISPLACEMENTON INTENSIFIER SPFH )650	50	1oo	150	200	250	300BASIC INTENSIFIER DISPLACEMENT (cm3/rev)2500crl50001063the "speed ratio"- and "intensifier speed"-lines (Fig . 3 .1 .2) . Hence, an intensifier displacementof 120 cm3 is used for further calculations.Step 4 - 6: The choice of bore, stroke and rod dimensions is an iterative process which isgoverned by the availability of piston seal and rod seal sizes as well as the stroke length of acrank case . During this iteration process, changes in intensifier displacement must beconsidered. Therefore, the initially selected basic displacement of 120 cm 3 can be considered aguideline rather than a fixed parameter.Based on this number, the relation between bore and stroke size is shown in Fig . 3.1.3.EFFECT OF BORE SIZEINTENSIFIER DISPLACEMENT: V =120 cm3/rev5	6BORE (cm)1 42 3 87 9 10 11R/B = 0.750.500.25THRUST FORCE (kN)CLEARANCE (%)*10020015010050302550Figure 3 .13 : The effect of the bore size on stroke length and thrust force.The stroke has been calculated as a function of bore and rod diameter for a double actingcylinder. To illustrate the effect of the rod dimension, three non-dimensionalized rod sizes havebeen investigated . They are given as ratio to the bore diameter (R/B = 0 .25, 0.50, 0.75) .64S =	 4V a 	 	(3 .1.5)[2- (/)2]B2 I TIt can be seen from Eqn. 3.1 .5 that the stroke length is only sensitive to the rod size for smallbore dimensions . The choice of bore and stroke dimensions is a compromise between maximumstroke length and maximum thrust load . Large stroke lengths are usually produced by heavy,large crankcases, whereas short stroke lengths need a large bore and, therefore, generate highthrust forces . The limiting factor for the bore size is the load capability of the crank case . Givena fixed stroke clearance (in Fig 3 .1.3 1 mm was chosen), the relative clearance volume increaseswith increasing bore size.The stroke size was determined by an 18 hp Kohler 4-stroke engine which became available tothe research group and was used as crank case . Its stroke is 8 .255 cm (3 .25") . Bore and roddimensions were obtained iteratively according to the equations used to generate Fig . 3 .1.3.SINGLE STAGE INTENSIFIER SPECIFICATIONSDIMENSION UNITSMETRIC IMPERIALSTROKE 8.255 cm 3.25 inBORE 3.175 cm 1 .25 inROD 1 .270 cm 0.50 inBASIC DISPLACEMENT 120.26 cm3 7.34 in3FORWARD STROKE 65 .36 cm3 3 .99 in3RETURN STROKE 54 .90 cm3 3 .35 in3Table 3 .1: Single stage intensifier dimensions65A design calculation of the rod is attached in Appendix A13 . It is based on the properties ofType 316 Stainless Steel and a limiting yield stress of 30,000 psi . The intensifier dimensions aresummarized in Table 3 .1.Based on the numbers presented in Table 3.1, the expected intensifier capacity can becalculated . At this design stage, the relative clearance volume of the lower compressionchamber was estimated approximately twice as large as the volume of the upper chamber . Thisestimate arises from the fact that the physical arrangement of the valves involves largerclearance spaces in the lower chamber and that the rod V-packing generates additional clearancespaceswhich is not present in the upper chamber . Therefore, a relative clearance volume of 4%and 8% was assumed for the upper and the lower compression chamber, respectively.Figure 3 .1 .4 : Expected capacity of single stage intensifier . Maximum intensifier speed : 700 rpm.3.1.2 ValvesThe compressor valves are actuated by the difference in pressure across the valve . A springassists the closing of the valve.INTENSIFIER CAPACITYAT VARIOUS INTENSIFIER SPEEDS AND PRESSURE RADIOSmaximum capacity : 43 kg/hrIntensifier Speed:700 rpm500 rpm300 rpm100 rpm0	2	4	6	8	10	12	14PRESSURE RATIO (P2/P l )50403020to066The following are desirable characteristics of a compressor valve:• low cracking pressure• low pressure drop across the valve• low clearance volume• low noise• no leakage• reliability• durabilityThe design goal is to find valves that have the least possible clearance space to achievemaximum volumetric efficiency. A set of compressor valves proposed by HoerbigerInternational was rejected because the clearance space of 6% and 16% for forward and returnstroke, respectively, is too large.Figure 3 .1.5 : NUPRO check valve ("CH" series) . 1 inlet body, 2 indicator ring,3 bonded poppet, 4 poppet stop, 5 spring, 6 0-ring body seal, 7 back-up ring,8 outlet bodyInstead, regular check valves have been chosen and modified to use as compressor valves.Check valves are designed for pressures up to 200 bar . The valve poppet incorporates a bondedViton 0-ring (Fig. 3 .1 .5-3) to provide better sealing, to eliminate noise and to cushion the valve67seat. Viton can withstand temperatures up to 200° C and is therefore suitable for the expectedcompression temperatures . The poppet is supported by a poppet stop (Fig . 3 .1.5-4) whichprovides a flow area for the gas and serves as seat for the valve spring (Fig . 3 .1.5-5).Figure 3 .1.6 : Axial cut through check valve ("CH" Series)The flow calculation (attached in Appendix A13) is based on a maximum gas flow per forwardstroke (since it is the largest chamber) at a 10 :1 pressure ratio . Under these conditions, the valveopens during the last 10% of the stroke . Hence, the instantanious mass flow is ten times thedesign massflow . It was concluded that the NUPRO'"' CH4 series valve was suitable . It is ratedat 400 bar (6000 psig) and its flow coefficient (c v) is 0 .67.To minimize clearance volume in the intensifier the cylinder caps were machined to house thevalve components . Each cylinder cap contains one inlet valve and one discharge valve . Theclearance space due to the valves is 0 .900 cm3 for the inlet valve and 0.0695 cm3 for thedischarge valve . This refers to a clearance space of 1 .48% for the forward stroke and 1 .77% forthe return stroke which is considerably lower than proposed by Hoerbiger . A 0.07 bar (1 psi)spring was chosen for the inlet valves and a 0 .7 bar (10 psi) spring for the discharge valves .683.1.3 SealingFrom a variety of sealing systems, "deep-type PARKER PolyPak" seals were chosen for bothrod and piston seals (cross section : 0.635 cm (0 .25 inch)).The seals consist of a V-cup which contains an 0-ring type "spring" which sits in the cup.Unlike regular lip-seals, the PolyPak system provides effective sealing for low and highpressures.Figure 3.1.7: The PolyPak sealing system [Parker Seals PolyPak Seal DesignHandbook PPD 3800 September 1989]At low pressures, the 0-ring generates a squeeze, thus, maintaining lip loading on both insideand outside diameter of the seal interface . As the system pressure increases, the pressure loadingof the lip takes effect and maintains the sealing . Under high pressures, the 0-spring stabilizesthe seal. These features are important for both the high system pressure (200 bar) and the widelychanging operating conditions.PTFE (i .e . Teflon) was chosen as sealing compound since it is capable of handling highpressures (>200 bar) and high temperatures (220°C) . The seals can be operated "dry" whichmeans that no lubrication is necessary .693.1.4 CoolingOne purpose of the experiments is to evaluate how sensitive the performance is to cooling.Three types of cooling are investigated:•	no cooling : the intensifier operates at ambient room temperature withoutadditional cooling• air cooling: cooling air at ambient room temperature is introduced with a fan•	water cooling: cold tap water is forwarded to a water jacket to cool the cylinder wall;(the valves are excluded from the water jacket)3.1.5 Capacity ControlThe flow control method chosen for the prototype intensifier is "control of gas discharge"(discussed in Chapter 2 .2 .3.4), since it requires the least modifications to intensifier and valves.This system is shown in Fig. 3 .1 .8 . It operates in two modes:1) Compression mode : This is the regular duty cycle where gas is drawn from the CNGtanks, compressed and delivered to an accumulator which stores compressed gas andforwards it to the injectors.2) Bypass mode : Since the intensifier runs at a fraction of engine speed, its capacitychanges according to the engine speed and gas can be produced in excess of the fuelrequirement (especially in off-design operation) . When the pressure in the accumulatorexceeds the injection pressure, a solenoid valve opens the bypass line which leadscompressed CNG back to the inlet port of the intensifier . During the bypass mode, gas iscycled through the intensifier and forwarded to the engine . In this mode, the gas pressurein the accumulator drops due to fuel consumption of the engine .70Figure 3.1.8 : Proposed flow control system (bypass system)71Eventually, the accumulator pressure decreases below a preset minimum pressure and thesolenoid valve closes the bypass line to reestablished the compression mode.The electronic signal for the solenoid valve comes from a pressure transducer measuring thepressure in the accumulator. An upper and a lower pressure limit can be adjusted which allowsminor pressure fluctuations of the discharge pressure.3.2 Single Stage Intensifier Components (Version 1 .1)The intensifier prototype as presented in Fig . 3 .2.1 consists of 7 main parts:1. Top Cap2. Cylinder3. Bottom Cap4. Tie Rods5. Cylinder Flange6. Piston7. Piston RodAll parts are made out of stainless steel type 316 since both natural gas and the environment ofthe bus operation contain corrosive materials . All adjacent parts that are exposed to highpressure CNG have been sealed with 0-rings.All parts illustrated in Fig 3 .2.1 have been machined in the workshop of the Department ofMechanical Engineering, UBC.The design goal was to minimize the size and the weight of the prototype as well as theclearance spaces . The components have been arranged to provide easy access for maintenancework. In particular, the exchange of rod and piston seals can be done without removing the72intensifier from the bus . Special tools have been made to facilitate removal and assembly ofseals.The existing piston of the crank case (Kohler engine) has been modified to hold the intensifierrod on a flange. The intensifier rod is screwed into the center of the flange and tightened with alock nut . This configuration allows a coarse adjustment of the stroke clearance.The top cap incorporates an inlet and a discharge valve. To minimize the clearance volume thevalves were arranged in axial direction . The top cap also includes a mounting hole for apiezoelectric pressure transducer which is used for a high-speed data acquisition of the cylinderpressure.The valves in the bottom cap required a different arrangement because of insufficient spacebetween bore and rod to accomodate the valve orifice . The problem was solved by machining aradial slot into the bottom cap providing an access to the horizontally arranged valves . The flowarea between slot and rod is twice as large as the area of the valve orifice . All edges have beenrounded to enhance aerodynamics . The bottom cap also incorporates the rod seal.The cylinder is axially centered between top and bottom cap and fastened with tie rods . Aclearance between the piston surface at TDC and BDC to the top and bottom cap, respectively(i .e . stroke clearance) allows an axial expansion of the rod.The clearance must be balanced against losses in volumetric efficiency due to increasedclearance volume. In this case, the total stroke clearance is 0 .89 mm (0 .035") of which 0 .635mm (0.025") is for the forward stroke and 0 .255 mm (0 .010") for the return stroke .73Figure 3 .2.1 : Intensifier Prototype Version 1 .1P I ST ON LIP-MEALSBOTTOM CAPDISCHARGE VALVEROD LIP-MEALCYLINDER FLANO E74The cylinder surface is polished and honed . After honing, the piston seals were worn in atmoderate intensifier speeds . The initial rubbing causes wear on the crevices and the cylinderwall is coated with the sealing compound (i .e . PTFE). Thus, the seals rub against a coating ofthe same material which increases the efficiency of the seals.The cylinder flange connects the cylinder configuration (i .e. top cap, cylinder and bottom cap)with the crank case. It also serves to cover the rod seals . The top cap is centered on a shoulderof the flange and sealed with an 0-ring. Venting holes have been machined in the side walls ofthe flange to prevent air compression by the crank case piston.A design criterion for the piston was to make it adjustable in axial direction in order to allowfine adjustment of the stroke clearance and to minimize clearance spaces due to fasteningdevices. The piston consists of three parts which are centered on radial shoulders (Fig. 3.2.2)and locked on the rod.To adjust the stroke clearance, the lower piston part is screwed onto the rod with the crank casepiston at BDC . The desired stroke clearance for the return stroke can be adjusted accurately,then the sealing rings and the other piston parts are assembled . The piston parts are sealed with0-rings to prevent gas leakage through the interior of the piston.The piston parts are fastened to the rod by applying torque to the upper piston part with thelower part held in the same position . The torque can be applied through two axial threaded holes(0 1/8") on the top surface of the upper and the lower piston part.Special tools have been made for this purpose . After the piston is assembled and locked, thethreaded holes are plugged with machine screws to reduce the clearance space .75Figure 3.2.2: Intensifier Piston Configuration763.3 Improvements (Version 1 .2) ,3.3.1 AlignmentMisalignment of moving parts such as the piston, the rod and the piston of the crank case causedlarge side forces which resulted in excessive wear of seals and low efficiency . The side forceson the crank case piston were in excess of its load carrying capability, thus, causing wear andmisalignment of the intensifier rod.Figure 3.3.1 : Crank case piston versus crosshead (designed and machined byTony Besic)The side loads were partly carried by the rod seals causing excessive wear and premature failure.The problem has been solved in three steps:i) The crank case piston was replaced by a crosshead (Fig. 3 .3.1) . Unlike a regular piston,the crosshead has a larger surface area, a very tight radial clearance and oilgroovesinstead of piston rings . It has a recess to center the rod attachement . The crosshead is77made out of aluminum which reduces weight . It is splash lubricated by the connectingrod which keeps friction coefficient between aluminum and steel sufficiently low.ii) To protect the rod seal from side forces, a plain linear bearing (Permaglide ® bushing')was installed on the cylinder flange (Fig. 3 .3.2) to guide the rod within a low radialclearance. The inside surface of the bearing is coated with bronze reinforced PTFE forreduced friction while allowing operating temperatures to 280°C and sliding velocities to2 mis.Figure 3 .3.2 : INA Permaglide® bushingUnlike regular oil impregnated bronze, the relatively soft Teflon surface prevents scoringof the rod, thus, improving the performance of the rod seals . The length of the bushingused is 2 .22 cm (0 .875").iii) With the crosshead and the Permaglide bushing, the reciprocating motion is rigidlyguided. However, due to machining tolerances and wear, there is still a chance of axialmisalignment between crosshead and intensifier piston . Therefore, a joint between'INA Permaglide ® Plain Bearings Catalogue 70378crosshead and intensifier rod was introduced (Fig . 3.3.3) which allows a radialmovement of the rod in the order of ±0 .25 mm.Figure 3.3 .3 : Floating rod attachement3.3.2SealingPreliminary tests proved PTFE to be the best sealing compound tested for this application.Compared to molythene and Viton, PTFE withstands higher operating temperatures withoutsignificant wear . Viton, which can also operate at high temperatures, generated considerablyhigher temperatures due to friction on the cylinder wall . The dynamic contact of Viton and steelcaused deposits on the cylinder wall which increased the friction even more . This chain reactionrendered the seal ineffective in a matter of minutes.While Viton requires lubrication to achieve satisfactory sealing, PTFE can be operated "dry" (i .e.without lubrication) . The deposit of Teflon on the cylinder wall has a positive effect on thesealing performance, as has been discussed in the previous section . High operating temperatures79seem to enhance the sealing performance of the lip seals as Teflon becomes softer and thepressure loading is more effective.While the piston seals were operating satisfactory, the rod seal needed improvement because ofgas leakage . The rod seal operates constantly at a maximum differential pressure of 200 bar(OP = Pdischarge -Patmospheric), whereas the piston seals are exposed to a maximum differentialpressure of 180 bar (depending on the current tank pressure Ap = Pupper chamber -Plower chamber) .Leakage across the rod seals would cause a fuel loss and a safety hazard.Tests showed that the single sealing lip of the rod lip seal was not sufficient to provide reliableand durable sealing under the dynamic conditions and radial misalignment . The single-lipPolyPak rod seal was replaced by a PTFE V-packing (Fig 3 .3.4) . The V-packing consists of anumber of V-shaped sealing rings (usually between 3 and 6) and a male and a female adapter.Axial pressure on the V-packing generates a static preload against the sealing surface . V-packing seals are designed for pressures up to 680 bar (10,000 psi).Following are the advantages of the V-packing compared to the lip seal:• variably adjustable preload• more sealing edges provide more effective and reliable sealing• flexibility to adapt to minor radial misalignment• failure occurs gradually as each individual ring wears out which makes maintenancework more predictable• the lifetime of ineffective V-packings can be prolongued by readjusting the axialpreload (e .g. adding a spacer or tightening the flange).• same compounds available and same positive operating behaviour as lip seals• no special tools are required for seal assembly80Figure 3.3.4 : Typical V-packing arrangement2 . The stack height for 5 rings (1/4" crosssection) is 2 .54 cm (1").The tapered design allows both sides of the ring to seal simultaneously as pressure is applied . Aminor disadvantage is the fact that the V-packing requires more space in axial direction, thus,requiring modifications to the bottom cap.3.3,3 LubricationThe performance of the V-packing and the Permaglide ® bushing can be improved byintroducing lubricant oil to the rod . However, a lubrication system is not aimed for, since itincreases the complexity of the intensifier.As compromise, very simple rod lubrication device was incorporated into the cylinder flange . Itconsists of a small oil reservoir and a pipe which forwards the lubricant to a circular groove,surrounding the rod (Fig . 3 .3 .6) . The groove is located between rod seal and linear bearing and2 "V-Packing" brochure, Power-Seal Corporation, Rochester Hills, Michigan81contains a felt ring which absorbs the oil and wets the rod as it reciprocates . No oil pump isrequired.Manual lubrication in a similar matter has been proven effective on Version 1 .1 . Tests showedthat oil travels along the rod into the compression chamber . In general, this is an undesirablefact because if small quantities of lubrication oil reach the diesel engine and burn, the engineemissions may be negatively affected . However, if diesel can be used as lubricant, the effect isinsignificant . The investigation of the behaviour of diesel oil and other lubrication oils is part ofthe performance tests and will be evaluated later.3.3.4 ValvesThe operation of the valves was found to be satisfactory for pressure ratios up to 5 :1 . Higherpressure ratios and, thus, higher compression temperatures caused a problem for the bondedpoppet of the discharge valve . When the pressurized, hot gas (200 bar, 200° C) pushes thepoppet open, the initial jet causes parts of the bonded 0-ring to tear off . Once the poppet seat isdestroyed the valve leaks. This happens within minutes while running at a 10 :1 pressure ratio.The problem has been solved by replacing both poppet and poppet-stop with a newly designedbrass poppet (Fig. 3 .3.5).The lower poppet area was reduced to increase the gas flow area and the surfaces lapped toimprove sealing . Gas exits to the valve outlet through circumferential slots in the tapered sidewalls . The width of the slots was chosen according to the dimensions of the original poppetstop. A centered recess on the upper surface provides a seat for the valve spring . The maximumvalve lift was set to 2.2 mm (0 .085") .82Figure 3.3.5 : Valve Poppet (designed and machined by Len Drakes)3.3.5 Assembly DrawingThe goal of the modifications was to minimize physical changes to the prototype Version 1 .1,yet, apply all improvements stated above . The cylinder configuration, the valves and the pistonhave not been changed except for minor modifications to the bottom cap . Figure 3.3.6illustrates the intensifier prototype Version 1 .2.A new seal flange was required to accomodate the V-packings . It is sandwiched between bottomcap and cylinder flange. A centered recess ensures alignment of the aluminum seal flange withthe bottom cap and the cylinder flange.The cylinder flange was modified to accomodate the lubrication device and the Permaglide®bushing. A centered recess on the bottom ensures alignment with the center of the crank case.The largest diameter of the intensifier (without crank case) is 15 .24 cm (6") . Its overall axialheight is 33 cm (13")83Figure 3 .3.6 : Intensifier Prototype Version 1 .284All workshop drawings and a parts list are attached in Appendix B1.A list of clearance spaces and their location is attached in Appendix B3 . The absolute clearancevolume is 1 .5205 cm3 and 2 .1755 cm3 for forward and return stroke, respectively. Thecorresponding relative clearance volume is 2 .33 % and 3.96 % for forward and return stroke,respectively. A plot of the theoretical volumetric efficiency is attached in Appendix B3.3.4 Two-Stage Intensifier Specifications (Version 2 .1)3.4.1 GeneralExtremely high operating temperatures and low volumetric efficiencies of the single stageintensifier at design pressure ratio led to the decision to modify the existing prototype to a two-stage intensifier . The modifications were based on the experience with the single stageprototype and can be considered an improved version rather than a completely new design . Thelower chamber of the double acting cylinder has been converted to operate as the second stageand the same bore and stroke sizes were used.The single stage intensifier components which have proven to operate satisfactory have beenadapted to the two-stage design, including:• the piston configuration and piston size• the piston seals• the rod V-packing (larger size)• the rod lubrication device• the Permaglide® bushing as linear guide-bearing• the floating rod mount• the crank case• the crosshead85BOTTOM CAFigure 3.4 .1 : Two-stage intensifier concept3.4.2 Sizing of StagesA stage volume ratio of v = 2.78 has been achieved by increasing the cylinder rod from 1 .27V2cm to 2.54 cm . Assuming 4% and 8% relative clearance volume for the first and the secondstage, respectively, the stage pressure ratios are plotted versus the overall pressure ratio in Fig3.4.2 . Note that the pressure ratio in the second stage remains 1 :1 at low overall pressure ratios.When the inlet pressure is high, the discharge pressure (200 bar) can be achieved with the firststage, only . Hence, no compression is necessary in the second stage and the pressure ratio is 1 :1.A bypass line was installed between interstage and receiver line which allows gas to flow fromthe interstage directly into the receiver line if the interstage pressure is greater or equal thedischarge pressure .86STAGE PRESSURE RATIOS2-STACIE INTENSIFIER, INTERC OGLING4	5	6	7OVERALL PRESSURE RATIO)0.53 .532 .521 .530 2 8 109 11Figure 3.4.2: Stage pressure ratios of two-stage intensifierThe mass-continuity equation is nowmSTAGEI - mSTAGE2 + mBYPASS	(3.4.1)where rhBYPASS is the amount of gas being forwarded directly into the receiver line and msTAGE1and tSTAGE2 are the massflows through the first and the second stage, respectively. Consideringthe gas conditions at each stage, Eqn. 3.4.1 can be written as:Pi Vi	- P2 .V2 . nV2 +mBYPASS	(3.4.2)with n 1, 1 and fl v2 being the volumetric efficency of the corresponding stage . Gas will flowthrough the bypass line as long as Eqn 3.43 is satisfied.mBYPASS - PI V • rj Y1 - P 2 V2 • ri v2  0	(3.43)Considering the stage volume ratio, this criterion can be simplified to:P2	.Pi . 2.78?Tin(3.4.4)At design pressure ratio, the maximum stage pressure ratio is 3 .01 :1 in the first stage and 3 .321 :1in the second stage .87The expected volumetric efficiency of the two-stage intensifier according to the assumedclearance volumes is compared with the single stage intensifier in Fig . 3.4.3.VOLUMETRIC EFFICIENCY2-STAGE INTENSIFIER, INTRRCOOLING4	5	6	7OVERALL PRESSURE RATIO3 8 119 toFigure 3.4.3: Expected volumetric efficiency of two-stage intensifierThe two-stage intensifier dimensions are summarized in Table 3.4.1.TWO-STAGE INTENSIFIER SPECIFICATIONSDIMENSION UNITSMETRIC IMPERIALStroke 8 .255 cm 3 .25 inBore 3 .175 cm 1 .25 inRod 2.54 cm 1 .00 inDisplacement Stage 1 65 .36 cm 3 3.99 in 3Displacement Stage 2 23.51 cm 3 1 .436 in3Table 3 .2 : Two-stage intensifier dimensions88Knowing the volumetric efficiency and the displacement of each stage, the expected intensifiercapacity can be calculated fromm = P 1 Vimn 'nmcRT,where VDI is the displacement of the first stage, and Hint the intensifier speed in rpm.Note from Fig. 3.4.4 that the maximum intensifier speed is 850 rpm to achieve the requiredcapacity at design pressure ratio (10:1) .(3.4 .5)TWO-STAGE INTENSIFIER CAPACITYBASED ON FIRST STAGE4	5	6	7OVERALL PRESSURE RATIO0504030201030 2 8 10 119Figure 3.4.4: Expected intensifier capacity at various intensifier speeds3.4.3 ValvesExperimental results indicated that the valves used in the single stage intensifier have excessivepressure drop at the required massflow . Therefore, the valve area has been enlarged by a factorof four. Utilizing existing parts of a Nupro CH8 Series check valve (ref . Fig. 3 .1.4 and Fig3.1 .5) a new inlet valve has been designed for the first stage with the goal of minimizing theclearance volume due to the valves .89ADJUSTABLE VALVE DISK	VALVE COVERVALVE SPRINGCNG INLET PORTBORE	VALVE POPPETFigure 3.4.5: Valve configuration of the first stageThe inlet valve consists of a valve poppet, a rod and an adjustable disk (Fig 3 .4.5). The Viton-bonded CH8 Series poppet is equipped with a cone to enhance the aerodynamics of the incomingflow. The adjustable valve disk serves three purposes:• guide the linear motion of the valve• hold the valve spring• adjust the valve liftThe disk is secured with a lock nut . A 0.07 bar (1 psi) spring was used.The discharge valve (CH8 Series) of the first stage is arranged similar to the single stagedischarge valve, except, it is rotated to the side due to lack of space on the top .90Lack of space was particularly a problem on the second stage where larger valves could not beused. Instead, the number of valves (CH4 Series) was doubled . The valves are located in thecylinder wall rather than leading into a radial slot as with the single stage model . The brasspoppets (see Fig . 3.3.5) are used in the discharge valves of the second stage.3.4.4 Assembly DrawingThe assembly drawing of the two-stage intensifier prototype is presented in Fig . 3 .4.6. Note thatthe cylinder is equipped with cooling fins to enhance heat transfer from the cylinder wall.All workshop drawings as well as a parts list are attached in Appendix B2.A list of clearance spaces and their location is attached in Appendix B4.The overall axial length of the intensifer (without crank case) is 30 .00 cm (13") and itsmaximum diameter is 15 .24 cm (6") .91Figure 3 .4.6: Two-stage intensifier prototype (Version 2 .1)92Chapter 4Experimental Arrangement4.1 Test Apparatus,A test apparatus) was designed to measure the intensifier performance . Figure 4 .1 .1 shows thatthe intensifier is driven by a variable speed electric motor, simulating the speed of the dieselengine. The motor is trunnion mounted, which means that the motor casing is mounted onbearings to allow frictionless rotational motion of the motor casing . With this arrangement thetorque can be measured directly using a strain gage type load cell which is mounted to a torquearm. Knowing the distance from the center of the motor shaft to the load cell, the measuredforce can be translated into torque.The electric motor has been upgraded to 25 HP and equipped with a DC motor controller . Themotor and the intensifier are mounted on a steel frame which is cushioned and isolated withrubber pads . A tachometer is mounted to the motor shaft . Its signal is sent to the motorcontroller as feedback and to the data acquisition system.A gear belt was chosen as means of speed transmission . The gear belt is simmilar to a tooth belt,except, it is designed for higher loads . It is important that there is no slip in the belt since theintensifier speed is calculated from the measured motor speed and the speed ratio of the gear belttransmission . The intensifier pulley is supported by a bearing block to avoid excessive bendingmoments in the crank case shaft.1 The specification of the equipment is attached in Appendix Cl .93Figure 4.1.1:Testrig94The intensifier pulley is supported by a bearing block to avoid excessive bending moments in thecrank case shaft . A flywheel is mounted to the shaft of the electric motor . Its purpose is togenerate additional inertia to overcome the typical cyclic peak torques due to the gascompression.4.2 Piping Plan of Single Stage Intensifier (Version 1 .2)Following data are taken to calculate the performance of the single stage intensifier:• CNG Inlet Pressure (bar)• CNG Inlet Temperature (°C)• CNG Massflow (kg/hr)• Cylinder Wall Temperature (°C)• CNG Discharge Pressure (bar)• CNG Discharge Temperature (°C)• Torque (N m)• Motor Speed (rpm)The specification of the measurement devices and the calibrations are listed in Appendix C2.Figure 4.2.1 shows that gas is drawn from a CNG storage tank and filtered (F) before it isforwarded to the main gas line . A switch board allows the selection of a specific tank andregulates the filling of the tanks with an external natural gas compressor.The inlet pressure of the gas is adjusted with a metering valve (MV) . Then the gas is forwardedto a Coriolis massflow meter2 (FM) which measures the amount of gas directly (rather thancalculating the massflow from the gas condition) . It uses an obstructionless U-shaped tubewhich vibrates at its natural frequency as the fluid passes through it . The tube twists as aresultof the flowing fluid and the angular movement of the tube created by the vibration.2Rosemount Technical Data Sheet 303195VVVI/IPIPELINEINSULATED PIPELINE® CHECK VALVEGAS FLOWBYPASSFigure 4.2.1 Piping Plan for Single Stage Intensifier.1CNG TANKS96Table 4 .1 : Legend for Figure 4 .2.1.Sensors measure the amount of twist which is proportional to the massflow rate through thepipe. The Coriolis meter operates independent of density, pressure and viscosity.The CNG massflow is measured prior to gas compression because the massflow meter is sharedbetween the intensifier and the diesel engine . The limitation of this setup is the fact that gasleakage is not taken into account in the massflow measurement.Then the temperature (Tl) and the pressure (Pl) of the incoming gas are measured . Tests haveshown that the instantaneous flow in the pipes due to filling of the compression chamber causesa pressure drop which reduces the amount of gas reaching the compression chamber and, hence,reduces the volumetric efficiency. Therefore, an accumulator (ACC) has been installed close tothe intensifier inlet port to compensate for frictional losses due to the flow through a 20 m pipe(O.D.: 6.35 mm) . The capacity of the accumulator is about 10 times the amount of gas drawnper stroke . This way, the oscillating flow in the pipe is damped out and the compressionchamber is charged without significant pressure drop.The gas is forwarded in both upper and lower chamber of the double acting cylinder,compressed to discharge pressure and forwarded into the receiver line . The top cap of theintensifier is equipped with a piezo electric pressure transducer which measures the relativepressure change in the cylinder (Pc).P1. . .CNG Inlet PressureT1 . . .CNG Inlet TemperatureP2. ..CNG Discharge PressureT2. . .CNG Discharge TemperatureTw. . .Cylinder Wall TemperaturePc. . .Cylinder PressureF. . .FilterMV. . .Metering ValveFM. . .Flow MeterACC. . .AccumulatorBV . . .Bleeding ValveRV. . .Pressure RegulatorTV. . .Throttling Valve97A thermocouple is mounted to the cylinder to measure the intensifier wall temperaure (Tw).The discharge temperature (T2) of the compressed gas is measured where the pipes of the upperand lower compression chamber join. The pipes are insulated from the discharge valves to thethermocouple to reduce heat transfer from the pipes.Both inlet and discharge pressure are measured with strain gage type pressure transducers . Thedistance from the main line to the pressure transducer has been extended to avoid malfunctionand damage due to the heat of the gas.The discharge pressure is adjusted with a pressure relief valve (RV)by varying the preload of aspring. The gas is then returned to the switch board and forwarded into a receiver tank.A bypass line has been installed to simulate the flow control system and to facilitate the startingprocedure. A metering valve (MV) opens and closes the bypass line . A manual relief valve(BV) allows venting the gas to the atmosphere . When the bypass line is open, compressed gasflows back to the accumulator in the inlet line . As a result, inlet and discharge pressure adjust tothe same level and no compression takes place . The inlet line upstreams of the point wherebypass and main line join is secured with a check valve to avoid gas expansion to the tanks.All CNG pipes and fittings used are made of stainless steel . The outside diameter is 6 .35 mm(1/4") except for the inlet line which is 9 .525 mm (3/8") to reduce the L/D ratio.4.3 Piging Plan of Two Stagg Intensifier (Version 2 .1)The two-stage intensifier uses the upper compression chamber as first stage and the lowerchamber as second stage. Note from Fig. 4.3.1 that the piping configuration between the CNG98tanks and inlet port of the first chamber is simmilar to the single stage configuration, except for alarger inlet pipe (1/2") between accumulator and inlet valve (because a larger valve is used).The interstage pressure (P2) and the interstage temperature (T2) before intercooling aremeasured. If the interstage pressure is higher or equal the discharge pressure (P3) the gas isdirectly forwarded into the receiver line . A special bypass line has been installed for thispurpose. The flow direction is controlled with a check valve.If P2 is lower than P3 the gas flows through an intercooler (IC) into an interstage accumulator(INT). The use of the intercooler is optional . Before the gas enters the second compressionstage the temperature of the intercooled gas is measured (T2').After the second compression stage both discharge temperature (T3) and discharge pressure (P3)are measured. The discharge pressure is controlled with the same pressure relief valve (RV)used for the single stage configuration.The intensifier is equipped with two piezo electric pressure transducers to monitor the cylinderpressure in both compression chambers (Pc1 and Pc2) . The cylinder wall temperature ismeasured as described above.No changes have been introduced to the bypass line.A high pressure accumulator is installed upstreams of the pressure regulator RV to damp outpressure fluctuations generated by the discharge of the compressed gas into the receiver line3 .3Due to the pulsing flow of the gas, the spring loaded pressrue regulator (i .e. a relieve valve) opened and closedwith each stroke causing an unsteady discharge pressure . With the accumulator installed, the pressure fluctuationsare damped out and a steady gas flow is maintained across the pressure regulator99PIPELINE O.D. : 6.35 mmPIPELINE O.D.: 12.7 mmCHECK VALVEREGULAR GAS FLOWBYPASSED GAS FLOW	CNG TANKSFigure 4.3.1 Piping Plan for Two-Stage Intensifier100Table 4.2 : Legend for Figure Data Acquisition SystemThe intensifier data acquistion system is equipped for both steady state and high speed dataacquisition. It is structured to operate both systems independently and to control and monitorthe process of acquiring data on-line via PC . The system is partly shared with the 6V-92 enginedata acquisition system.All steady state signals are sent to signal conditioning devices4, except for the interstage pressuresignal which is forwarded directly to an interface board . The signal conditoning providescomplete signal conditioning function including filtering, amplification, high noise rejection andwide range zero protection for all analog signals as well as the cold junction compensation forthe thermocouples . The following modules are used:• 5B31 : Isolated Voltage Input (Tachometer)• 5B32: Isolated Current Input (Massflow Meter)• 5B37-J-01 : Isolated Thermocouple Input - Type J (Thermocouples)• 5B38: Isolated Strain Gage Input (Pressure Transducers, Loadcell)These modules provide a high level voltage output of 0 - 5 V4Technical specifications of all data acquisition components are attached in appendix C3P1 . . .CNG Inlet PressureT1. . .CNG Inlet TemperatureP2 . . .CNG Interstage PressureT2. . .CNG Interstage Temperature before coolingT2' . . .CNG Interstage Temperature after coolingP3 . . .CNG Discharge PressureT3. . .CNG Discharge TemperatureTw. . .Cylinder Wall TemperaturePcl . . .Cylinder Pressure (first stage)Pc2. . .Cylinder Pressure (second stage)F. . .FilterMV. . .Metering ValveFM. . .Flow MeterACC . . .AccumulatorINT. . .InterstageTV. . .Throttle ValveBV. . .Bleeding ValveIC. . .IntercoolerRV. . .Pressure Regulator101ISAAC ISAAC 2000> DATAACQUISITIONq Crank Angle Position ( .Snftpot)q Cylinder Pressure (PCB) 	 CHARGEGy.Toque (Nm)Cylinder Wall Tempernh ire (°C) 	Motor Speed (rpm)	• Inlet Pressure (bar) 	• C)i diet Pressure (hot) 	• Inlet Temperate ire (°C:)	• Ch diet Temperate ire (°C) 	• Mncsflow (kg/hr)	• Interstage Pressure (bar) 	• Interstage Temp . RC (°C)	© Interstage Temp . AC (°C)6V-92	I	 1SensorsAMPLIFIERTRIGGERIDIGITAL DISPLAY / ALARMCrmat. 1 (outlet 1 C inlet 1 (outletcurrentJ temp. ) press.J press . 4-SIGNAL CONDITIONING MODULESmax .	max .	min .	max.alarm	alarm	alarm	alarm 1Stop5 5 5 5 5BStart3B3BMotor Speed rpm DisplayMOTOR CONTROL PANEL3B B35B5B5B5B53B6V-92SensorsV 1	rINTERFACE	C 	 Intensifier Sensors » SWITCHBOARD	C 	 6V-99 Sensors 	 >	BOX	 nIBM PC286 000A/D CARDPCL-818	Analog I/C)0	 IIIIIIIIIIIIIIIIUI	/j- 2,ft\Digital I/OPIFigure 4 .4.1 : Intensifier Data Acquisition System102To facilitate monitoring, a display panel has been installed featuring four digital displays . Inaddition, all displayed signals such as armature motor current, CNG end temperature and CNGinlet and end pressure are equipped with an alarm.The motor control panel is equipped with a motor speed regulator, a digital rpm readout andother motor control features . It interacts with the regenerative motor controller which controlsthe speed via tachometer feedback.After signal conditioning, all steady state signals are organized in an interface board andforwarded to an A/D board (PCL 818) which is part of a 286 IBM compatible PCThe procedure of steady state data acquisition is discussed in Appendix C4The high speed data acquisition system consists of a piezoelectric pressure transducer, an opticalshaft encoder, an external clock and trigger, an ISAAC computer and a general purpose interfacebus between the ISAAC and the host computer.The center piece of the high speed data acquisition system is the ISAAC5 computer . It is anintelligent data acquisition and control peripherical system that is connected to an ASCII-speaking host computer . It can acquire, store, mathematically manipulate and return to the hostdigitized values of signals generated by sensors, transducers and instruments . It can alsogenerate analog and digital signals over a wide range of speeds and voltages . Using the SSHmethod (simultaneously sample and hold), any phase shift between the (up to four) signals iseliminated.The system uses an external clock and trigger for the acquisition of data . The clock is the signalof the shaft encoder which provides one pulse per crank angle degree and 360 pulses total . In5 ISAAC 2000 Hardware Reference Manual103this setup, the amount of data as well as the clock speed is only dependent on the shaft speed anddoes not require internal clock speed adjustment for each test . The trigger is the index pulse,also provided by the shaft encoder but delivered on a seperate channel . It is sent when theintensifier piston is at BDC position. The high speed data acquisition begins at this position andtakes data over one entire (crank) shaft revolution . 360 cylinder pressure data points (in psi) areacquired in one test cycle.The required power is provided by a charge amplifier s . A scope has been installed to be able toconstantly monitor the cylinder pressure . The high speed data acquisition procedure is discussedin Appendix C4.4.5 Data Acquisition SoftwareThe data acquisition software is an in-house made PC-based program which directs the flow ofthe incoming voltage signals, converts them into engineering values, performs calculations,converts binary pressure data to ASCII and saves all data to disk.The system consists of three parts : the main program, a calibration file and a configuration file.The main program controls the acquisition of steady state and high speed data . All analogsignals are sampled 200 times and then converted in the PCL-818 to 12-bit digital numbers.Then each channel's average and standard deviation is calculated . The averaged data is used forintensifier performance calculations such as power consumption, isentropic efficiency andvolumetric efficiency . The calculated data, as well as the averaged data and the standarddeviation are stored on seperate files.6for charge amplifier specifications and settings see Appendix C3104The main program also directs the acquisition of high speed data which is saved in binary formunder a specified filename. Later, the binary file can be converted to ASCII and loaded in aspreadsheet.The calibration file contains information relating to the conversion of the voltage output signalsof each electronic sensor to engineering units . Thus, all calibration data is transferred to this filein the form of voltage-engineering unit pairs . If the calibration curve is linear, only two of thepairs are required, whereas if the curve is non-linear, it can be split into up to 10 pairs.Following is a sample part of the calibration file:CALIBRATION FILE FOR INTENSIFIER DATA ACQUISITIONLast modified 11-30-1992Intensifier Version 1 .02, Mechanical Drie, E-MotorTorque	(Measurement)0	(Voltage Range (# defined in software))N-m	(Units)-5,-.02271, 1 .369, 5, 0, 0, 0, 0, 0, 0	(Voltage Signal)-456.56, 0, 128.11, 456.56, 0, 0, 0, 0, 0,	(Corresponding Engineering Units)The configuration file contains fixed in : t nsifier parameters such as number of stages, intensifierdimensions and speed ratio. It allows the operator to set up the channels and configure thescreen.Calibration and configuration files are automatically accessed by the program at the beginningand at the end of a test session . The operator can select and modify the files before or during atest.The calibration and configuration files used for single stage and two stage intensifier tests areattached in Appendix C4 and C5, respectively . All calculations used for the intensifierpreformance tests are attaced in the same Appendices .105Chapter 5Results5.1 GeneralIn this chapter, the performance data of the intensifier is presented and the intensifer componentsare discussed. All data is plotted against the pressure ratio which is defined as : II= PendPinktThe discharge pressure was kept constant, while the inlet pressure was varied to achieve thedesired II-range, simulating the real application . Each performance test was done at constantintensifier speed . The structure of the data presentation has been chosen to facilitate thecomparison of single stage and two-stage intensifier performance.First, the typical performance data of both intensifier prototypes are presented, where only theinlet pressure was varied for a given intensifier speed . Then, the sensitivity of variousparameters is investigated, including:• the effect of intensifier speed• the effect of discharge pressure• the effect of cylinder-wall cooling (applies to single stage intensifier)• the effect of intercooling (applies to two-stage intensifier)• the effect of flow control methods (applies to single stage intensifier)In addition, the pressure-volume (p-V) diagram of the gas compression is analysed and thepolytropic coefficient of compression and expansion is determined .1065.2 Test LimitationsAs mentioned in Chapter 4, the flow meter is located upstream of the intensifier . Therefore, anyleakage through the rod seal is not accounted for because the measured flow rate can be higherthan the flow rate actually delivered by the intensifier.mmeasured = mdeliered + mleakageHowever, natural gas is scented with a strong odour for the purpose of detecting leaks inpipelines. As soon as the rod seals were worn sufficiently that natural gas could be smelled, thetest results were disregarded.Natural gas has been taken from a regular storage tank and, after compression, returned to astorage tank of the same size. Therefore, the testing time was limited by the pressure rise in thereceiver tank. For tests with higher intensifier speeds (and, hence, higher mass flow rate of gas),the number of data points is smaller as the pressure builds up faster . Usually, between 50 and150 data points were taken in each performance test . The intensifier prototypes were runningless than 100 hours.Typically, the time duration of the test was 10 minutes . During tests with the single-stageintensifier the discharge temperature continuously rose, even though the pressure ratio decreasedfrom about 12 :1 to 8 :1 . The operation was judged unsave when the discharge temperatureexceeded 220°C and the test was discontinued Thus, none of the data shown in Figures 5 .3 .1 to5.4.7 are steady-state experimental data . Since the high-speed cylinder pressure data weresampled in less than a second, these can be regarded as quasi-steady-state data, though subject toheat transfer.The typically high peak torque due to the compression in a piston compressor turned out to be aproblem for the electric motor. Even though it has been upgraded to 25 hp, it stopped due tooverheating when running at low speeds (i .e . less than 500 rpm) . As a result, the minimumintensifier speed tested was 100 rpm .1075.3 Accuracy and RepeatabilityThe process of redesigning and optimizing the single-stage prototype has taken more than oneyear. Tests have been conducted constantly during this time while improvements have beenintroduced . The data presented in this thesis are from the latest version . Tests were done on twodifferent days to verify the repeatability of the results which is ±5% . The data of the last day arepresented.All tests that have been performed to evaluate the effects of certain parameters have beenconducted on one day in the sequence presented.In a valid test the intensifier speed varies in less than ±1 % of its preset value.All steady-state data were taken continuously during a test . The results showed that the massflow rate readings fluctuated when adusting the inlet pressure . After about 30 seconds, steady-state conditions were re-established . Therefore, all readings that differ from the previous massflow rate reading in more than ±3 % have been excluded.The results are plotted against pressure ratio as mentioned above . However, pressurefluctuations of both inlet and discharge pressure were encountered which can not be detectedfrom looking at the pressure ratio . Thus, a clarifying statement is appropriate . A generalguideline has been established allowing a maximum fluctuation of the discharge pressure of ±10% (of full range) . This margin has been selected because it seems realistic for a real application.The inlet pressure is adjusted to the desired magnitude to establish steady state conditions . Itturned out, however, that the inlet pressure fluctuates, too l . The same pressure ratio might1 The expansion of the gas from tank pressure to adjusted inlet pressure cools the gas and water vapour cancondense and freeze at the orifice of the metering valve . If this occurs, the orifice area is reduced and the inletpressure decreases even more . Eventually, the gas flow blasts the ice from the orifice and the pressure risesinstantaneously .108therefore represent two different cases . For example, a pressure ratio of 10 :1 is achieved with210 bar discharge pressure and 21 bar inlet pressure, as well as 190 bar discharge pressure and19 bar inlet pressure . Both cases are considered valid . To give a complete picture as to how thepressure ratio was achieved, a plot of both inlet and discharge pressure is presented along withthe other data.After extracting the obsolete data, all data points were fit by a 3 rd order polynomial of the form:D = ao+a 1 x+a2 •x2 +a3 •x3where D is the regressed data, x the pressure ratio and ai the constants of the polynomial . Theregression has been performed on a Lotus spreadsheet.The obtained equation has been applied to the range of pressure ratios actually measured. Thisrange is between 7 :1 and 12:1 for the single stage intensifier and between 4 :1 and 13 :1 for thetwo-stage model.Figure 5.3.1: Sample data regression for volumetric efficiency . (100 rpm intensifierspeed and 200 bar discharge pressure).DATA REGRESSIONDATA OF THREE DIFFERENT TESTS, IOU RPM, 200 bar10006	7	8	9	10PRESSURE RATIO (P2/1'1)ACQUIRED EXPERIMENTAL DATA	FITTED CURVE5 131211109Figure 5.3.1 shows an example of a data regression performed for the volumetric efficiency ofthe single stage prototype . In this case, the data of three different tests was used to generate theregression . Usually, the data of only one test is used.After a flywheel and a support bearing had been added to the test rig (Fig . 4 .1.1), the torque dueto the friction of the bearings was measured . Following bearings and parts were considered:• bearings in the electric motor• trunnion bearings monting the electric motor• gear belt• flywheel and flywheel bearings• crankcase bearings• shaft support bearingFigure 5.3.2 : Frictional torque due to bearings and flywheelThe result of this measurement (Fig 5 .3.2) led to the conclusion to correct the torque . Theobtained data was fit by a linear equation which was used to correct the torque measurementaccording to the current motor speed as shown below.TORQUE DUE TO BEARINGS AND FLYWHEELCRANK CASE WITHOUT INTENSIFIER MOUNTED._' Torque (N-m).i Power (kW)S•	•	•.	.-500	0	500	1000	1500	2000	2500	3000MOTOR SPEED (rpm). Torque (N-m)	. Power(kW)	linear regression32.521 .50.50-0 .5110Tom,, = 7;._ – (1.371 + 3.94 . 10-4 • RPM)	(5.4.1)5.4 Sine1e Sta ge Intensifier PerformanceThe intensifier performance was measured at three intensifier speeds : 100, 150 and 200 rpm.Higher speeds have not been considered because of problems with the alignment of moving partsand valve problems . Both issues are discussed in detail later.5.4.1Intensifier CauacityThe intensifier capacity aimed for is 43 kg at design pressure ratio (10 :1) and 700 rpm intensifierspeed. Over the tested speed range, the intensifier capacity exceeds the expected capacity 2 .Figure 5.4.1: Single stage intensifier capacity (Version 1 .2).The reason for this is the clearance volume being smaller than originally assumed.2ref. Fig. 3.1 .4111Figure 5.4.1 shows the CNG massflow at different speeds over a range of pressure ratios . Themaximum massflow at design condition is 13 .2 kg/hr at 200 rpm intensifier speed . Higherspeeds were not attempted because larger massflows are not possible with the presentcompressor valves . In Table 5 .4 .1 the experimental results are compared with the expectedcapacity.The volume flow was calculated in atmospheric conditions (i .e . normal cubic meters per hour(Nm3/hr) and standard cubic feet per minute (SCFM)) from the measured massflow . Bothgraphs are attached in Appendix D2.200 12 .3 13.2 18.8 11 .1150 9.2 10.1 14.6 8.6100 6.1 6.9 10.0 5.9Table 5.1 : Comparison between expected and measured capacity of intensifier Version 1 . Volumetric EfficiencyThe theoretical behaviour of the volumetric efficiency has been discussed and defined earlier.The theoretical curves are presented together with the experimental data in Fig.5.4.2.Clearance volumes of 3% and 4% have been assumed for forward and return stroke,respectively3 and the isentropic coefficient of natural gas is 1 .3 [6] . The real-gas effects weretaken into account by measuring the CNG flow rate directly rather than calculating it fromsteady-state pressure and temperature.3 Compressor manufacturer specify only a minimum clearance volume . No data of volumetric efficiencies areavailable. The clearance volume of the single stage intensifier is defined in Appendix B3 . All calculationscarried out by the software are attached in Appendix C5.112Figure 5.4.2 : Volumetric efficiency of single stage intensifier for different speedsIt can be seen that the volumetric efficiency (VE) decreases with higher intensifier speeds . Themaximum VE at design pressure ratio is 67 .3% at 100 rpm intensifier speed and 200 bardischarge pressure.The results show a disagreement between the estimated averaged clearance volume of 3 .07 %4and the experimental data . However, some clearance spaces have been neglected which mayhave more impact on the results than originally assumed.5.4.3 Power ConsumptionThe intensifier power consumption has been calculated from the torque generated by electricmotor and the motor speed:=T,..•rpmMOWR•2 n•60 .1103(1')4ref. Appendix B3113The corrected torque has been used in the calculations as well as the directly measured CNGmassflow.Figure 5.4.3 : Intensifier power consumption (Version 1.2).POWER CONSUMPTION PER UNIT MASS FLOW RATEDISCHARUE PRESSURE : 200 bar0.190.180.17vu0 .130 .120 .110 .1876 1312 149	10	11PRESSURE RATIO (P2/P 1)Figure 5 .4.4: Intensifier power consumption per unit mass flow rate.114The result is illustrated in Fig 5.4.3 . It can be seen that the power consumption is more sensitiveto changes in intensifier speed than to changes in pressure ratio . This is evident from Eqn.1.3.3as both higher intensifier speed and higher inlet pressure increase the massflow and, thus, thepower consumption.5.4.4 Isentronic EfficiencyThe isentropic efficiency is here defined as the ratio of the theoretical (isentropic) powerconsumption for an ideal gas to the actual power consumption calculated from the torque and themotor speed . Corrected torque was used for the calculation.k-1k	pz k	1k -1 mcivc R Ti	p1	13.6406in which »t~,c is the measured gas flow rate.Figure 5.4.5 : Isentropic efficiency of the single stage intensifier (Version 1 .2)Tlbear .T,,.,,.- rpmmoeor '2 . 7r 6 . 101	•100 (in %)115Note from Fig. 5.4.5 that isentropic efficiency is more sensitive to changes in intensifier speedthan volumetric efficiency . Again, the highest isentropic efficiency is obtained with slowintensifier speeds . At 100 rpm and design condition, isentropic efficiency is 86.4% and changesvery little in the off-design region.The values for isentropic efficiency at design conditions and 150 and 200 rpm are 78.7% and72.5%, respectively . The performance seems to be increasingly efficient for very high pressureratios which was unexpected. This phenomenon is discussed in detail where the pressure-volume diagrams are discussed . It can be stated in advance, though, that the reason for this wasfound to be the valves which turned out to be too small to handle large mass flow rates . As aresult, the pressure build-up in the cylinder exceeded the discharge pressure, thus causing highertorques and excessive power consumption . Large mass flow rates occur when the inlet pressureis high or when the intensifier speed is high. This can be observed for the 200 rpm case in Fig5.4.5 ,where the efficiency is considerably lower compared to the lower speeds and isparticularly low for higher inlet pressures (i .e. lower pressure ratios).5.4.5 Operating PressuresThe operating pressures are presented in Fig .5 .4.6 for reference on how the pressure ratio wasderived. No data regression has been performed. Only 200 rpm data is presented, since theother cases behave in a similar way.It can be seen that the discharge pressure varies by +7% and -5% around the nominal value of200 bar .116Figure 5 .4.6 : Inlet and discharge pressures for the 200 rpm case5.4.6 Ouerating TemperaturesFigure 1 .4.5 shows estimated discharge temperatures exceeding 220°C following isentropicsingle-stage compression . Figure 5.4.7 shows experimental temperatures which confirms theprediction of discharge temperatures higher than 200 °C . It must be noted that the thermocouplewarms up slowly and since a performance test takes less than 10 minutes, a steady statetemperature is only reached at the end of the test.Furthermore, the discharge temperature measured may underestimate the actual dischargetemperature since the thermocouple is located where the discharge pipes of the upper and lowerchambers join (ref. Fig 4.2.1) . Although the pipes are insulated, some cooling of the gas likelyoccurs during the travel from discharge port to the location of the thermocouple.The inlet temperature is in many cases considerably lower than ambient temperature because thegas is cooled while being throttled from tank pressure to the adjusted inlet pressure . The actualOPERATING PRESSURESINTENSIFIER SPEED: 200 rpmMA. ..DISCHARGE PRESSUREINLET PRESSURE...r	MP	. .OMB	a. •7	8	9	10	11	12	13	14PRESSURE RATIO (P2/PI)250200Dozo500117inlet temperature depends on the throtteling and on how much the gas warms up as it travels tothe inlet port of the intensifier (distance : -20 m).Figure 5.4.7 : Discharge , inlet and cylinder wall temperature of single stage intensifierperformance test at 200 bar discharge pressure and 200 rpm intensifier speed.In general, higher pressure ratios generate higher discharge temperatures . Figure 5.4.7 showsthat the discharge temperature at higher pressure ratios is lower than at lower pressure ratios,which is due to the warm up time of the pipes and the thermocouple at the beginning of a test.Initially, the inlet pressure is low. It is slowly increased to achieve lower pressure ratios.Therefore, the discharge temperature is lower at high pressure ratios.The cylinder wall temperature reached a maximum of 115 °C . The high operating temperaturesled to the decision not to attempt an endurance test of the single stage intensifier . The impact ofthe high temperatures on the intensifier parts is discussed in Chapter 5 .4 .12.OPERATING TEMPERATURESINTENSIFIER SPEED: 200 rpm, DISCHARGE PRESSURE : 200 bar25020050010	11PRESSURE RATIO (P2/P 1)87 9 13 14121185.4.7 Cylinder Pressure DataThe main purpose of taking cylinder pressure data is to get an insight as to how the valvesoperate and how the gas behaves under compression and expansion . In addition, it is possible todraw the pressure-volume diagram and find the polytropic coefficient which can be used topredict the off-design behaviour of the intensifier.Figure 5.4.8 illustrates the cylinder pressure at 129 rpm intensifier speed . At design pressureratio the peak cylinder pressure is 220 bar which exceeds the design pressure by about 10%.This case can be considered satisfactory . At 8:1 pressure ratio the back pressure is alsoacceptable (12%).The cylinder pressure data are plotted against the angular position of the crank shaft . One fullcycle of the upper compression chamber (ref . Fig. 4.2.1) has been recorded for two intensifierspeeds and various inlet pressures . The piezo-electric pressure transducer accurately measuresthe change in pressure but does not measure absolute pressure . Hence, a reference pressure isrequired to scale the pressure data. The reference point chosen is that of the first crank angleposition (i .e. BDC). The pressure is adjusted to the steady-state inlet pressure which is recordedshortly before the high-speed data acquisition system is initiated . The discharge pressure wasrecorded as well to calculate the current pressure ratio . The time-constant setting of the chargeamplifier was set to "Medium" for all tests.At higher intensifier speeds, the back pressure increases rapidly which can be seen in Fig . 5.4.9.Higher intensifier speed means higher mass flow rate but the mass ,flow per stroke is the same,given the same inlet and discharge condition . The gas has to be forwarded through the valvesfaster at higher speeds. In other words, when the intensifier runs slower, the gas has more timeto flow through the valve . Therefore, only little pressure build-up occurs .119Figure 5 .4.8: Cylinder pressure data at 129 rpm intensifier speed and 200 bar dischargepressure.Figure 5.4.9: Cylinder pressure data at 193 rpm and 200 barAt design pressure ratio the back pressure is as high as 250 bar (i .e 25% back pressure) . Thisillustrates clearly that the massflow rate creates a serious problem for the valves . It can beCYLINDER PRESSURE vs CRANK ANGLEINTENSIFIER SPEW: 1291pm, DISCHARGE PRESSURE : 205 and 197 bar250Peak Pressure : 224 barDischarge Pressure	205 barPressure Ratio:197/20 .2 = 9 .75 : 1205/21 .2 = 9 .67 : 1205/25.5 = 8.04 : 12002,1501002Z) 5o— Inlet Pressure :197 bar0200CRANK ANGLE (°)0 100 300 4000	100 200CRANK ANGLE (°)IN . PRESS. 28 .9 bar	IN . PRESS. 23 .4 bar	IN . PRESS. 20 .7 bar300	400CYLINDER PRESSURE vs CRANK ANGLEINTENSIFIER SPEED : 193 rpm, DISCHARGE PRESSURE : 204 bar300Peak Pressure c©23.4 bar:0150 —2 10050Discharge Pressure : 204 barPressure Ratio:204/20 .7 = 9.86 : 1204/23 .4 = 8.72 : 1204/28 .9 = 7.06 : 1260 bar120predicted that the situation is more serious for off-design performance (i .e . lower pressure ratiosand/or higher speed) where the mass flow rate is higher and, therefore, the gas velocity is higher.This phenomenon explains the poor isentropic efficiency at 200 rpm, particularly for lowerpressure ratios.The volumetric efficiency is affected ,too, because the gas cannot be moved out of thecompression chamber sufficiently fast . Therefore, more residual gas remains in the clearancespaces and the pressure in the cylinder is higher than the discharge pressure at TDC position.Hence, the gas expands from a higher pressure and less fresh gas can enter the cylinder . Inaddition, the high compression generates higher temperatures which causes the fresh, incominggas to warm up and expand.5.4.8 The Pressure-Volume Dia gramAt design pressure ratio the pressure-volume diagram has been drawn for 129 and 193 rpmintensifier speed . The theoretical curve, based on 2.5% clearance volume, is plotted to comparethe experimental result with theory. The polytropic coefficient for compression and expansionhas been extracted from the experimental data by fitting an exponential curve of the form:Y=Ae X"where n is the polytropic coefficient and Ao a constant . Integrating this simple equation gives:hi(Y) = ln(Aa) + n • 1n(X)which can be transformed to:Y' =Ao +n . X•Data in this form can be fit by a linear equation using the data regression function of a Lotusspreadsheet .121Figure 5.4.10 : Cylinder pressure versus displaced volume at design pressure ratio.(Intensifier speed = 129 rpm)The excess pressure after the opening of the discharge valve can be seen very clearly in Fig.5.4.10 . This indicates that the discharge valve opens instantaneously (i .e . weak spring force) butthe small valve orifice causes a significant pressure drop. Comparing Fig . 5.4.10 and Fig.5.4.11, the excessive discharge pressure effect is stronger at higher intensifier speeds . Theperformance of the inlet valve is satisfactory since no pressure drop occurs during the filling ofthe cylinder. The disagreement of the expansion line with theory is due to excessive residual gasin the clearance spaces and due to the fact that the expansion coefficient changes over theexpansion stroke. During the first part of expansion the gas temperature is higher than thecylinder wall temperature. Hence, the gas is cooled . As the expansion continues, the gastemperature decreases below cylinder wall temperature and is heated.A constant polytropic coefficient is used for the computation of the theoretical curve over theentire expansion and compression which is obtained from the regression of the data.P-V DiagramIntensifier Speed : 129 rpm, Discharge Pressure: 205.8 bar25020050010	20_ Inlet Pressure 21 .2 bar30	40	50	60VOLUME (cmA3)THEORY 2 .5% Clearance Volume0 70 80122Figure 5.4.11: Cylinder pressure versus displaced volume at design pressure ratio.(Intensifier speed = 193 rpm)The volumetric efficiency can be determined from the p-V diagram by observing the ratio of thelength of the "filling" line to the total displacement . In Fig. 5.4.11 the length of the "filling" lineis 67 - 12.4 = 54.6 cm3 and the total displacement is 67 cm 3. Hence, the volumetric efficiencyis := 67 •100 = 81.7 %This number disagrees considerably with the data presented in Fig . 5.4.2 where the maximumvolumetric efficiency at design pressure ratio is about 67% . The reason for this is the fact thatthe relative clearance volume of the lower chamber is higher than in the upper chamber, thus, thevolumetric efficiency of the lower chamber is lower. Since the steady state data of volumetricefficiency is an average between upper and lower chamber, this average number will be lowerthan the number obtained from the p-V diagram (which is from the upper compressionP-V DiagramIntensifier Speed : 193 rpm, Discharge Pressure : 204.2 bar30025050010	20Inlet Pressure 20.7 bar30	40	50	60VOLUME (cm"3)THEORY 25/o Clearance Volume0 8070123chamber) . However, to confirm this theory, the volumetric efficiency of the lower chambermust be as low as 51% . Two possible reasons can be given to explain this phenomenon s:1) The lower compression chamber contains a number of clearance spaces whosemagnitude are not known. One of those clearance spaces is the volume of the pressurepockets of the V-packing that generate the preload to the rod (ref . Fig. 3 .3.4).2) The valve configuration of the lower chamber is unfavourable as the gas is forcedthrough a radial slot and has to turn 90° to exit through the discharge valve . This maycause excessive pressure build-up and, thus, low volumetric efficiencies, since theexpansion stroke starts at a higher pressure than discharge pressure.In the logarithmic p-V diagram (ln(p)-ln(V) diagram) the compression and expansion curvesappear as straight lines, assuming a constant polytropic coefficient . In Fig. 5.4.12 and Fig.5.4.13 the previously discussed cases illustrate good agreement for the compression stroke.The gap between theoretical and experimental expansion line is better visible in the ln(p)-ln(V)diagram. Initially, the slope of the expansion curve in Fig . 5 .4.12 is Hater than the theoreticalcurve with a constant expansion coefficient . At this point, the gas temperature is higher than thecylinder wall temperature and heat is transferred from hte gas to the cylinder wall . Hence, thepolytropic coefficient is smaller at the beginning of the expansion stroke and the slope is flatter.Note that the flat part of the expansion line may also be caused by gas leakage through thedischarge valve which can occur when the valve poppet travels from its maximum lift position toits seat6 .Only the data of last half of the expansion curve has been used for the data regression . Theexpansion coefficient of the lower speed case (i .e . Fig. 5.4.12) is slightly higher which isexpected since more time is avaliable for heat transfer.5Gas leakage must be excluded as possible explaination since the CNG massflow is measured upstreams of theintensifier. Thus, the volumetric efficiency of the lower chamber is not affected by leakage.6Both valve poppet and valve seat are metals which can cause the valve poppet to bounce of its seat or tremblewhen it first touches down . In the above case a 10 psi valve spring was used .124Figure 5 .4.12: ln(P)-ln(V) diagram at design pressure ratio . (Intensifier speed = 129 rpm)Figure 5 .4.13 : ln(P)-ln(V) diagram at design pressure ratio . (Intensifier speed = 193 rpm)The values of the expansion coefficients are 1 .213 and 1.202 for the 129 and the 193 rpm case,respectively. The compression coefficient is approximately equal for both cases (i .e. -1 .264).In(P) - In(V) DiagramIntensifier Speed: 129 ipm, Discharge Pressure 205.8 bar1Inlet Pressure : 21 .2 bar2	41n(V)_-._. THEORY 2.5% Clearance Volume5In(P) - In(V) DiagramIntensifier Speed 193 fpm, Discharge Pressure 204.2 bar2ln(V)THEORY 25'/o Clearance Volume5Inlet Pressure 20.7 bar125Finally, it must be remarked that no cooling was introduced to the intensifier during theperformance tests of the previously discussed cases.5.4.9 The Effect of Intensifier SneedAll data were plotted for various speed cases and the effect of intensifier speed has beenmentioned occasionally . The most important impacts of speed, based on test experiences and onthe figures presented above, are:• The intensifier performance is less efficient at higher speeds . Both isentropic andvolumetric efficiency are negatively affected by speed.• The p-V diagrams indicate a lower expansion coefficient for higher speeds wherecooling is not as effective, but the difference may not be significant.• Intensifier speeds higher than 200 rpm have been performed but knocking in the crankcase occured (possibly due to misalignment), causing excessive wear.• The main reason for the poor performance at high intensifier speeds is due toinadequate discharge valve size and unfavourable valve configuration in the bottomchamber.• Higher speeds generate higher operating temperatures . Even though the intensifierparts have not shown excessive wear due to heat, it can be assumed that higher operatingtemperatures increase the wear and reduce the lifetime of parts.5.4.10 The Effect of Changes in Discharge PressureSince it is not yet certain if the injection pressure will be 200 bar or lower, tests were undertakento measure the intensifier performance at lower discharge pressures . All tests were done on thesame day and under the same conditions . The intensifier speed was 150 rpm . All data wereregressed and fit by a 3rd order polynomial .126First the test at 200 bar was repeated . Then the discharge pressure was adjusted between 170and 180 bar and the inlet pressure was varied to achieve a variety of different pressure ratios (allwithin 8 :1 and 13 :1). The test was then repeated with a discharge pressure between 145 and 160bar.The results show that at the same pressure ratio the volumetric efficiency (rl v) is not verysensitive to changes in discharge pressure . At design pressure ratio, rk, varies between 71% and68% at different discharge pressures. The results are attached in Appendix D3.5.4.11 Capacity ControlThe flow control method originally considered was "bypassing", where the discharged gas iscircled through the bypass line and returned to the intake valve at discharge pressure.Figure 5.4.16: Back pressure due to bypassing.CYLINDER PRESSURE DUE TO BYPASSINGINI'E NSIFIER SF}±D : 150 rpm, VARIOUS PRESSURl S250Steady State Line Pressure :161 bar143 bar200CRANK ANGLE (°)300	4001007ref. Chapter 2 .2 .3 .4127Preliminary operation with bypassing indicated that the pressure drop across the discharge valvecauses excessive power consumption . To trace the origin of this phenomenon, cylinder pressuredata were taken at various line pressures (i .e . pinta = Pdischarge)• The gas was forwarded throughthe cylinder without compression. Theoretically, this operation should not cause any powerconsumption, except to overcome the valve preload due to the spring (i .e . 0 .7 bar)Figure 5 .4.16 shows that back pressure increases with increasing line pressure and that thepressure rise in the cylinder is considerably higher than 0.7 bar. The maximum line pressureavailable for this experiment was 161 bar but in a regular operation the gas is bypassed at at 200bar (i .e . discharge pressure) . The results indicate that the back pressure and , thus, the powerconsumption at 200 bar is even higher than illustrated in Fig . 5.4.16 . In addition, "bypassing"generates high temperatures . The pressure peaks are due to the dynamic response of the valvespring .CYLINDER PRESSURE DUE TO VALVE UNLOADINGINTENSIFIER SPFH1): 150 rpm, VARIOUS PRESSURES200150100250:JSteady State Line Pressure:165 bar133 bar101 bar77 bar50 bar100	200	300	400CRANK ANGLE (°)Figure 5 .4.17: Cylinder pressure due to valve unloading128Since "bypassing" seems to be an inefficient way of controlling the intensifier capacity, "valveunloading" has been considered as alternatives. This method is based on the control of the inletvalves . The gas compression is prevented by opening the valves during the compression stroke.Valve unloading was experimentally simulated by removing the valve poppet and running theintensifier at various inlet pressures.POWER CONSUMPTION DUE TO FLOW CONTROLINTENSIFIER SPEED : 150 rpmtoo	150	200LINE PRESSURE (bar)Figure 5 .4.18: Power consumption due to flow control methodsFigure 5 .4.16 illustrates that the pressure fluctuation due to "valve unloading" is considerablylower than due to "bypassing" . Unlike "bypassing" this method operates at the current CNGinlet pressure and, hence, becomes more efficient with lower inlet pressures.Both experiments have been done at 150 rpm intensifier speed . Figure 5 .4.18 illustrates that"valve unloading" is more efficient than "bypassing" . However, a drawback of "valveunloading" is the fact that it requires a more complex valve mechanism.8ref. Chapter 2 .2 .3 .31295.5 Two-Stage Intensifier PerformanceThe performance of the two-stage intensifier was measured at the same intensifier speeds as thesingle stage intensifier (100, 150 and 200 rpm) to facilitate comparing the data from the twoprototypes . The valve problems of the single stage intensifier have been solved and thealignment has been improved which resulted in less cyclic variation of the data . The data werefit by a linear regression as shown in Eqn. 5.3.1 . The range of pressure ratios has been extendedto 4:1 as a result of better valve performance. The pressure ratios were achieved by adjusting ahigh inlet pressure which was slowly decreased . Thus, the pressure ratio was slowly increased.Furthermore, data were taken after a 3 minute warm-up period . Therefore, all two-stageintensifier performance data can be considered steady-state.5.5.1 Intensifier CanacitvThe two-stage intensifier capacity of various intensifier speeds and pressure ratios is shown inFig. 5.5.1.Figure 5.5.1 : Two-Stage Intensifier Capacity .130Although the intensifier displacement has been reduced by 45% compared to the single stageintensifier, the capacity of the two-stage intensifier is only 25% less than that of the single stageintensifier (at 200 rpm) . The reason for this is higher volumetric efficiency due to larger valvesand lower stage pressure ratios.The maximum CNG mass flow rate at design pressure ratio is 9 .9 kg/hr at 200 rpm intensifierspeed. The results are summarized in Table 5.3.200 10.3 9.9 14.4 8.5150 7.7 7.3 10.6 6.3100 5.1 5.1 7.4 4.4Table 53 : Comparison between expected and measured two-stage intensifier capacity.The measured capacity is very close to that calculated in the design phase . The capacity curvesin Nm3/hr and SCFM are attached in Appendix El.5.5.2 Stage Pressure RatiosFigure 5.5.2 shows that the stage pressure ratios at design pressure ratio are 3 .8:1 and 2 .6:1 forthe first and second stage, respectively . The stage pressure ratios have been predicted as afunction of the stage volume ratio and the clearance volume of each stage and are shown in Fig.3.4 .2. At design pressure ratio the predicted pressure ratios were 3 .01:1 and 3 .321 :1 for first andsecond stage, respectively . The discrepancy between predicted and measured stage pressureratio is due to the clearance ratio being different than predicted and due to a heating effect of the131rod which greatly influences the volumetric efficiency of the second stage. The latterphenomenon is discussed in Chapter 5.5.8. The stage pressure ratios are not very sensititve tochanges in intensifier speed.It was found that the valve performance greatly influences the stage pressure ratios l . Theintensifier performance is very sensitive to small particles in the gas.STAGE PRESSURE RATIOS4.5	DISCHARGE PRESSURE : 200 bar2.5	M ~w AA .•.—JI..• M3 .534rM"Joft	."h 0. oft ;	iygt"M~!!4'rl'1•21.513	4	5	6	7	8	9	10	11	12OVERALL PRESSURE RATIO▪Stage 1, 100 rpm. Stage 2, 100 rpm. Stage 1, 150 rpm•Stage 2 150 rpm. Stage 1, 200 rpm Stage 2, 200 rpmFigure 5.5.2: Stage pressure ratios of the two-stage intensifier.5.5.3 Volumetric EfficiencyThe volumetric efficiency was measured for both stages. By definition, the volumetricefficiency of the intensifier is equal to the volumetric efficiency of the first stage. Figure 5.5.3shows the volumetric efficiency of the first stage. At design pressure ratio ri, ranges between90.5 % and 92.5 %. The increase in volumetric efficiency can be explained by lower pressureratio and low clearance volume of the first stage. In addition, the increased valve orifice'At one point, the stage pressure ratios changed rapidly and the intensifier had to be shut off. When inspecting thevalves, a small metal particle was found between the brass poppet of one of the discharge valves of the secondstage and its valve seat which caused leakage gas flowing back into the second stage. Thus, the volumetricefficiency of the second stage was drastically reduced which resulted in an increase in the interstage pressure.13 14132supports higher mass flow rates, thus, allowing measurements at higher inlet pressures . Thevolumetric efficiency is not very sensitive to changes in intensifier speed.Figure 5 .5.3: Volumetric efficiency of the first stage.VOLUMETRIC EFFICIENCY STAGE 1DISCHARGE PRESSURE : 200 BARlos806	8OVERALL PRESSURE RATIOFIITT1:1) CURVE2 4 10 1412VOLUMETRIC EFFICIENCY STAGE 2DISCHARGE PRESSURE : 200 BAR1101009060642 12 16148	10OVERALL PRESSURE RATIOFII1TU.) CURVEFigure 5.5.4 : Volumetric efficiency of the second stage .133The volumetric efficiency of the second stage is more sensitive to changes in pressure ratio thanthat of the first stage ; this is due to a larger clearance ratio . Figure 5.5.4 shows that rlv isbetween 75 % and 78 % at design pressure ratio . At 4:1 pressure ratio the volumetric efficiencybegins to exceed 100 % . This phenomenon can be explained by the fact that the gas flow ispartially bypassing the second stage if the inlet pressure is sufficiently high . In this case themeasured CNG massflow is higher than the CNG massflow forwarded through the second stage.Hence, the volumetric efficiency is higher than 100 %.5.5.4Power ConsumptionThe torque measurement shows cyclic variations which is due to sampling problems2 .Therefore, the scattering of the data points is more extensive . Furthermore, it was found that thepower consumption is influenced by the operating temperature . Two identical tests have beendone successively.THE EFFECT OF OPERATING TEMPERATUREIN'I ENSWIER SPN1-1): 100 RPM, DISCHARGE PRESSURE: 200 l3AR4	5	6	7	8	9	10	11	12PRESSURE RATIO	TORQUE TEST 1	. TORQUE TESL' 2—WALL TEMPERATURE TEST I	 .WALL TEMPERATURE TEST 2Figure 5.5.5: The effect of operating temperature on the measured torque.2AIl data are sampled 200 times and averaged, based on the acquired data . The sampling rate may not besufficient to reproduce the cyclic peak torques exactly .134The torque curve of both tests and the cylinder wall temperature are plotted in Fig. 5.5.5 . Thegraph shows that the torque, measured during the second test, is lower than that of the first test atlow pressure ratios and higher at high pressure ratios . Higher torque at high pressure ratios canbe explained by an increase in friction due to heat expansion of moving parts such as the rodand/or the crosshead . On the other hand, lower torque at low pressure ratios may be due tobetter gliding characteristics of the Permaglide® bushing. The change in torque is non linearand is not completely understood at this point . However, it was found that the volumetricefficiency and the stage pressure ratios remain unaffected . It must be concluded that the powerconsumption and, thus, the mechanical efficiency is extremely sensitive to friction caused byheat expansion . Therefore, only those data are presented that were taken when the cylinder wallwas at ambient room temperature . This allows fair comparison of tests at different intensifierspeeds and of tests with the single stage intensifier.POWER CONSUMPTIONDISCHARGE PRESSURE : 200 BAR8	10OVERALL PRESSURE RATIOFIIIT D CURVE42 6 1612 14200 RPM100 RPM150 RPM2.50Figure 5.5.6 : Power consumption of the two-stage intensifier.135At design pressure ratio the intensifier power consumption is 0 .75 kW and 2.1 kW for 100 and200 rpm, respectively . The corresponding mass flow for these two cases is 5.1 and 9.9 kg/hr.Figure 5 .5.6 shows that the power consumption is higher for higher speeds which is due toincreased mass flow and lower efficiencies.The power consumption per unit mass flow is plotted in Figure 5 .5.7 . The convenience of thisgraph is that the power consumption can be easily calculated for each CNG mass flow rate at agiven pressure ratio.Figure 5 .5.7 : Power consumption per mass flow rate of the two-stage intensifier.At design pressure ratio the minimum power consumption per unit mass flow rate is 0 .15kW/kg/hr. Note from Figure 5.5.7 that the power consumption is very sensitive to intensifierspeed. Heat expansion and increased friction at higher intensifier speeds are responsible forincreased power consumption per unit mass flow.POWER PER MASS FLOW RATEDISCHARGE PRESSURE : 200 BAR0 .30.058	10OVERALL PRESSURE RATIOFIIT l'H1) CURVE642 12 14 161365.5,5Isentrouic Efficiency,The results of the isentropic efficiency confirm good performance at low intensifier speeds anddecrease in efficiency at higher intensifier speeds . Figure 5.5.8 shows that the isentropicefficiency is sensitive to both pressure ratio and intensifier speed. The maximum isentropicefficiency at design pressure ratio is 77% at 100 rpm intensifier speed. The isentropic efficiencyis higher at lower pressure ratios which indicates good valve performance.ISENTROPIC EFFICIENCYDISCHARGE PRESSURE: 200 BAR6	8OVERALL PRESSURE RATIOFIITTW CURVE:)10 12 14Figure 5.5.8 : Isentropic efficieiency of the two-stage intensifier.Note from Figure 5.5.8 that the isentropic efficiency of the 200 rpm case can be higher at lowpressure ratios than that of lower speeds3 if the moving parts are well aligned . At higherpressure ratios the intensifier has warmed up sufficiently to show frictional losses.3 The test data for the 200 rpm case presented in Figure 5.5 .8 is different from that of the previous figures .1375.5.6 Operating TemperaturesThe performance tests with the two-stage intensifier covered a wider range of pressure ratios.Therefore, more data points were acquired and the curves are smoother . The operatingtemperatures increase slowly because the pressure ratio was slowly increased . Figure 5.5.9shows typical operating temperatures of the two-stage intensifier running at 200 rpm with 200bar discharge pressure.Note from Fig . 5.5.9 that both the CNG discharge temperature and the cylinder wall temperatureare considerably lower than those of the single stage intensifier due to lower stage pressure ratioand intercooling . At design pressure ratio the CNG discharge temperature is 100°C and the walltemperature is approximately 65 - 70°C . The temperature is not very sensitive to changes inintensifier speed.120110100. 90b 80OPERATING TEMPERATURESDISCHARGE PRESSURE: 200 BAR, INI'ENSIFIL( SP	 H:~ ): 100 RPM4	5	6	7	8	9	10	11	12	13OVERALL PRESSURE RATIOINTERSTAGE Temp.BEFORE COOLINGCYLINDER WALL Temp.INTERSTAGE Temp.AFTER COOLINGINLET Temp.Figure 5.5.9 : Operating temperatures of the two-stage intensifier .138As CNG flows through the interstage it is cooled due to heat exchange between the CNG pipesand the air at ambient room temperature. Intercooling can reduce the CNG interstagetemperature by 50°C depending on pressure ratio and intensifier speed.5.5.7 The Pressure-Volume DiagramCylinder pressures were measured simultaneously in both stages during one entire intensifierrevolution. The reference point for the cylinder pressure of the first stage is the measuredsteady-state inlet pressure . The reference pressure is the discharge pressure of the first stage (i .e.interstage pressure) . Cylinder pressure data at 105:1 pressure ratio and 100 rpm are discussed indetail in this section. P-V diagrams of higher intensifier speeds and lower pressure ratios areattached in Appendix E2.Figure 5.5.10 shows the p-V diagram of the first stage at 4 .02 : 1 stage pressure ratio . To benoted is the fact that there is no pressure drop across the discharge valve of the first stage.CYLINDER PRESSURE vs VOLUMEStage 1, Intensifier Speed : 100 RPM30	40	50	60VOLUME (cm^3)Theory 25'/o Clearance Volume. Experimental Data70 80Figure 5.5.10: The p-V diagram of the first stage at 100 rpm and 4 .02 : 1 stage pressure ratio .139The pressure peaks in the discharge line are caused by the spring response . The steep slope ofthe expansion line indicates low clearance volume and confirms good sealing of the valvepoppet . A slight pressure drop can be detected from the filling line . The ln(p)-ln(V) diagram ofthe same data points is shown in Figure 5 .5.11 . The polytropic coefficients have beencalculated as described in Chapter 5 .4 .8. The calculated values are plotted with experimentaldata . The p-V diagram of the second stage at 2 .62 : 1 stage pressure ratio is shown in Figure5.5.12. Again very little pressure drop occurs for both filling and discharge . The slope of theexpansion line is not as steep as in the first stage . The reason for this is larger clearance ratio inthe second stage . Figure 5 .5.12 shows a bulge at the beginning of the discharge line.Figure 5 .5.11: The ln(p)-ln(V) diagram of the first stage at 100 rpm and 4 .02 : 1 stage pressureratioThis bulge is caused by the spring preload of the discharge valves . The cylinder pressureslightly exceeds the discharge pressure to open the discharge valve . As long as the pressuredrops back to discharge level, this phenomenon is not a concern.In (p) - In(V) DiagramStage 1, Intensifier Speed : 100 rpmn = 1 .51	1 .20IIIIII11IITh2	3	4	5In (V) Theory 25''/o Clearance Volume2.5054.531. Experimental Data140Figure 5.5.12 : The p-V diagram of the second stage at 100 rpm and 2 .62:1 stage pressure ratio.CYLINDER PRESSURE vs VOLUMEStage 2, Intensifier Speed : 100 rpm15	20	25VOLUME (cm^3)Theory 7 .27% Clearance Volume250100500DischargeStage Pressure Ratio : 2.62 : 1ExpansionCompressionFilling305. Experimental DatatoIn(p) - In(V) DiagramStage 2, Intensifier Speed : 150 rpm5.64.44.240 .5. Experimental Data2	2 .5	3hl(v) Theory 7.27% Clearance Volume1 .5 3 .5n = 0.67= 1 .53Figure 5.5.13 : The 1n(p)-ln(V) diagram of the second stage at 100 rpm and 2 .62:1 stagepressure ratio141The logarithmic curve of the second stage is shown in Fig. 5.5.13. Note that polytropiccoefficient of the expansion is considerably smaller than that of the first stage . The polytropiccoefficient of the compression is very similar to that calculated in the first stage.The p-V diagram for both stages is shown in Fig. 5.5.14 . All pressure-volume diagrams inAppendix E2 show both stages in on one graph . The logarithmic curves are not presented.Rather, the polytropic coefficient of compression and expansion for different intensifier speedsand inlet pressures are discussed in the next section.Figure 5 .5.14: The p-V diagram for both stages at 100 rpm and 10 .5: 1 overall pressure ratio.All data were taken with one inlet valve of the second stage plugged . At higher mass flow rates(i .e . m > 20kg/hr) the cylinder pressure data show a significant pressure drop across the inletvalve. This can be seen clearly in Fig . 5.5.15, where the inlet pressure is 75 .3 bar. The gascompression takes place in the first stage, only . In the second stage gas is only forwardedwithout compression . However, the high mass flow through the valves cause a pressure dropCYLINDER PRESSURE vs VOLUMEStage 1&2, Intensifier Speed : 100 rpmInlet Pressure: 18.9 barInterstage Pressure: 76 barDischarge Pressure : 199 barOverall Pressure Ratio 10 .5 : 10	10	20	30	40VOLUME (cm!`3). Experimental Data	Theory50 807060250200500142across the inlet valve . Hence, it must be concluded that two inlet valves are necessary for thesecond stage to support the maximum CNG mass flow.Figure 5.5.15: The p-V diagram of both stages at 100 rpm and 2 .9:1 overall pressure ratio.5.5.8 The PolvtroDic CoefficientThe polytropic coefficient was calculated from the p-V diagrams at various pressure ratios andintensifier speeds using the method described in Chapter 5 .4 .8.The results of Table 5.4 indicate that cooling is effective for the first stage . According to Table2.1, the gas is cooled during compression if n <k, which is the case for all speeds and pressureratios . The expansion coefficient for the first stage is slightly higher than k. Therefore, the gasis cooled during expansion . The polytropic coefficients of the second stage clearly indicate thatthe gas is heated during both compression and expansion . With respect to the isentropiccoefficient k, the compression coefficient is higher and the expansion coefficient is smaller . Theheating of the gas appears to be due to heat transfer from the piston rod . The temperature of theCYLINDER PRESSURE vs VOLUMEStage 1&2, Intensifier Speed : 100 rpmStage 2	▪._ •	Stage 1	%I. . .▪ ti .Inlet Pressure : 75.3 barInterstage Pressure : 220 barDischarge Pressure : 220 barOverall Pressure Ratio 2.9 : 1•...NM	N./. .-M.--,ps-0	10	20	30	40	50	60	70	80VOLUME (cm^3)Experimental Data25050143rod increases due to friction on the linear bearing and the rod seals . Since the rod surface islarge compared to the displacement volume, the heat transfer is very effective.Speed Pressure STAGE 1 STAGE 2(rpm) Ratio Compression Expansion Compression Expansion100 10.50 : 1 1 .19 1 .38 1 .55 0.6710 .15 :1 1 .19 1 .41 1 .58 0.676.77 : 1 1 .18 1 .37 1 .65 0.644.61 : 1 1 .17 1 .36 1 .83 .622.90 : 1 1 .24 1 .32 N/A N/A150 10 .40 : 1 1 .19 1 .31 1 .64 0.708.87 : 1 1 .18 1 .30 1 .69 0.696.40 : 1 1.17 1 .27 1.83 0.67200 12 .10 :1 1 .21 1 .41 1.63 0.72Table 5.4 : Polytropic coefficients of the two-stage intensifier.An empirical equation for the polytropic expansion coefficient is suggested [2] according to theinlet pressure of the gas and its isentropic coefficient in the form : n = 1 + C*(k - 1), where C =0.88 for pressures between 10 and 30 bar and k = 1 .3. For pressures higher than 30 bar, n = k.Therefore, the polytropic expansion coefficient should be 1 .264 at design pressure ratio . Thevalues presented in Table 5.4 show that the experimental expansion coefficients are somewhathigher than suggested by the empirical equation, indicating that the gas is cooled during theexpansion.The steady state polytropic coefficient of the first stage is plotted in Fig . 5.5.16 . against the stagepressure ratio of the first stage . The polytropic coefficient is calculated from Eqn 2 .2.1 and canbe written as:144(5.5.2)V1POLYTROPIC COEFFICIENT STAGE 1DISCHARGE PRESSURE : 200 BAR1 .291 .281 .27> .. 1 .26`~-~{ 1 .25J 1 .24Gi 1 .231 .22J1 .21x 1 .21.19I- 1 .181 .171 .162 .5 3 .5STAGE PRESSURE RATIO3 4 4.5FIITTED CURVEFigure 5.5.16 : The steady state polytropic coefficient of the first stage.Note from Figure 5 .5.6 that the data points seem to converge towards a linear drop withincreasing stage pressure ratio . Note that the data points at lower stage pressure ratios areconsiderably lower than that of higher stage pressure ratios which is due to a warm-up delay fothe thermocouples . Figure 5 .5.16 shows that higher intensifier speeds can generate higherpolytropic coefficients which indicates less heat transfer at higher speeds . The data points of the100 rpm case show a very linear behaviour because the test is longer and the number of datapoints is higher than that of the 150 and the 200 rpm case . The 100 rpm case was fit by astraight line :n = 1.2866 - 5.08*10-3*T1 1	(5 .53)where H1 is the pressure ratio of the first stage.1455.5.9 The Effect of Intensifier SneedThe effect of intensifier speed on the intensifier performance, based on the presented data andtest experience is summarized as follows:• The power consumption and, thus, the isentropic efficiency is sensitive to changes inintensifier speed . Frictional losses increase with increasing speed at higher pressureratios.• Intensifier capacity and volumetric efficiency are not affected significantly by theintensifier speed . The flow area of the valves is sufficiently large to support higher massflows at higher intensifier speeds.• The operating temperatures are not very sensitive to intensifier speed.• Misalignment and wear of moving parts are concerns for operation at higher intensifierspeeds5.5.9 The Effect of Changes in Discharge PressureSimulating the requirement of a lower injection pressure, a test was performed with 150 bardischarge pressure . While the volumetric efficiency is not affected by changes in dischargepressure, the isentropic efficiency is considerably lower at lower discharge pressures . Note fromFigure 5 .5.17 that an increase of 4% in isentropic efficiency can be observed at design pressureratio and 150 bar discharge pressure . The compression forces are smaller due to the lowerpressure . Therefore, lower side loads are generated which generate less friction .146THE EFFECT OF DISCHARGE PRESSUREINTENSIFIER SP1+3) 100 RPM0.20 .18000.160 .14 v~0.1220 .10.080 .06 00.048PRESSURE RATIO14Figure 5.5.17 : The Effect of Changes in Discharge Pressure.5.5.10 The Effect of IntercoolinEINTERCOOLING90	DISCHARGE PRESSURE : 200 bar, INTENSIFIER SPN7.7 ) : 100 rpm80NiV 70—~ 60g50FORCED CONVECTION	•,,,_• .•.A. .,• • ••~ • I M ••• •H• •TYPICAL INTERCOOLING4030r M•~ IM r •SF' rwM•rMM~• ~•y.ru . .mbe' .lam g •3	4 5	6 7	8	9PRESSURE RATIO10 11 12 13 14Figure 5 .5.18 : Intercooling Temperatures .147The CNG temperature after intercooling was held constant between 10 and 15°C . Note fromFigure 5 .5.18 that intercooling is more effective at higher pressure ratios because the stagepressure ratio of the first stage and, thus, the interstage temperature is increasing.Using intercooling, the inlet temperature of the second stage was reduced by 20% . Figure5.5.19 shows that the isentropic efficiency is sensitive to intercooling where an increase of 4% inisentropic efficiency can be observed at design pressure ratio . The volumetric efficiency is notaffected by intercooling . Lower cylinder wall temperatures and CNG discharge temperaturescan be observed when intercooling is used.THE EFFECT OF INTERCOOLINGDISCHARGE PRESSURE: 200 bar, INTENSIFIER SPEW : 100 rpm8PRESSURE RATIO0.20 .180 .160 .140 .120.10 .080 .064 6Figure 5.5.19 : The Effect of Intercooling.5.6 Comparison of Single Stage and Two-Stage Intensifier ,The intensifier performance has been improved in many respects with the two-stage intensifier.The results are summarized as follows :148• Volumetric Efficiency: Due to lower stage pressure ratio and better valve performance thevolumetric efficiency of the two-stage intensifier is 20% higher than that of the single stageintensifier. Furthermore, the volumetric efficiency of. the two-stage intensifier is not as sensitiveto changes in intensifer speed and massflow as that of the single stage intensifier.• Isentropic Efficiency : The isentropic efficiency of both intensifier prototypes is verysensitive to changes in intensifier speed . Friction forces are responsible for lower efficiencies athigher intensifier speeds and higher pressure ratios . The isentropic efficiency was not improvedby increasing the number of stages.Figure 5 .6.1 : Power Consumption per Unit Mass for Single Stage and Two-Stage Intensifier.• Power Consumption : The power consumption per unit mass flow (kW/(kg/hr)) is shown forboth intensifiers at three intensifier speeds in Figure 5 .6.1 . The graph shows that the powerconsumption is sensitive to changes in intensifier speed for both intensifiers . The increase inpower is related to friction. The power consumption can be lower with the two-stage intensifier.POWER CONSUMPTION PER UNIT MASSSINGLE STAGE AND T W O-STAGE INTENSIFIER8	10	12PRESSURE RATIOSingle Stage 100 rpm	Single Stage 150 rpmSingle Stage 200 rpm	Two-Stage Intensifier0 .350 .30 .250.20 .150.10 .0564 1614149However, the two-stage intensifier seems to be more sensitive to intensifier speed than the singlestage intensifier.• Valve Arrangement: Size and configuration of both inlet and discharge valves were found tobe important parameters for the intensifier performance . Note that the improvement of theintensifier performance is not only due to the increase in number stages but also due to valveimprovements. Increasing the valve orifice and the number of valves solved problems with backpressure in the cylinder as is shown in the p-V diagrams of single stage and two-stage intensifier.• Operating Temperatures : CNG discharge temperature and cylinder wall temperature havebeen drastically reduced with the two-stage intensifier. The maximum CNG dischargetemperature of the single stage intensifier can be reduced by 115°C using multistaging if noartificial cooling is applied . Furthermore, the operating temperatures of the two-stage intensifierdo not increase significantly with increasing intensifier speed . The lower operating temperaturesare due to multistaging and intercooling . The temperatures can be reduced even more ifeffective intercooling is applied. The cylinder wall temperature can be reduced by 50°C with thetwo-stage intensifier.• Capacity Control: The capacity control methods discussed in Chapter 2 .2.3 are moredifficult to implement in a two-stage intensifier . Physical arrangement of control devices andincreased clearance volume are a concern with valve unloading . Bypassing is not recommendsince the stage pressure ratio cannot be maintained .150Chapter 6Conclusions and Suggestions6.1 ConclusionsThe objective of this research was to design and build an intensifier prototype which compressesnatural gas from a variable, elevated-pressure storage source (20 - 200 bar) to a constant, highdischarge pressure (200 bar) . The intensifier is intended to deliver variable amounts of gaseousfuel (5 - 43 kg/hr) to a heavy duty bus diesel engine which has been converted to a gas-dieselengine. The intensifier performance was to be measured and the effects of operationalparameters were to be investigated. Furthermore, the intensifier design was to be evaluated.Based on the CNG pressure and mass flow requirements of the gas-diesel engine, a number ofalternative intensifier design concepts have been studied and evaluated . The crank-shaft drivenreciprocating intensifier design was selected because of lower cost, simplicity of design and thepotential to operate at higher efficiencies compared to hydraulic and pneumatic reciprocatingdrivers.A single-stage double-acting intensifier prototype was designed and built . Based on early testresults and experience with the intensifier operation, a new, improved intensifier design (Version1 .2) was proposed. Improvements were adressed to alignment-related problems of moving parts,rod seals, discharge valves and rod lubrication . Steady-state and high-speed cylinder-pressuredata were recorded at different intensifier speeds and pressure ratios .151Low volumetric efficiency, high operating temperatures and excessive valve flow velocities athigher mass now rates called for multistaging and further improvements. A double-acting, two-stage reciprocating intensifier (Version 2 .1) was designed and built which uses the displacementof the forward stroke as first stage and the displacement of the return stroke as second stage.Bore and stroke dimensions of the single-stage intensifier were adapted to the two-stage design.An adequate stage volume ratio has been achieved by increasing the rod diameter . The valveorifice area and the number of valves have been increased . The two-stage intensifierperformance was measured at the same intensifier speeds and similar pressure ratios.The following is a brief summary of the performance of the single-stage and two-stageintensifier prototypes Version 1 .2 and Version 2.1:• A constant CNG discharge pressure of 200 bar was achieved with both single-stage and two-stage intensifiers.• Pressure ratios between 6 :1 and 12:1 were performed with the single-stage intensifier . Thetwo-stage intensifier covered pressure ratios between 4 :1 and 14 :1 . Hence, the design pressureratio of 10 :1 was exceeded with both intensifier prototypes.• The minimum measured power consumptions per unit mass flow at design pressure ratio are0.143 kW/(kg/hr) and 0.137 kW/(kg/hr) for single-stage and two-stage intensifiers, respectively.• The maximum isentropic efficiencies at design pressure ratio are 86% and 77% for single-stage and two-stage intensifiers, respectively.• The maximum volumetric efficiencies at design pressure ratio are 67 .3% and 92 .5% forsingle-stage and two-stage intensifiers, respectively .152• At design pressure ratio the stage pressure ratios of the two-stage intensifier are 3 .8 :1 and2 .6:1 of first and second stages, respectively . The stage pressure ratios are not sensitive tochanges in intensifier speed.The following conclusions can be drawn from the intensifier performance:• The measured intensifier capacity of both single-stage and two-stages intensifier agrees verywell with the predicted intensifier capacity in the tested speed range . However, performancetests have been done in the lower speed range (100 - 200 rpm) only . Higher intensifier speedshave not been attempted due to friction- and alignment-related problems.• Lower CNG discharge pressures (150 -175 bar) do not significantly affect the single-stageintensifier performance . The performance of the two-stage intensifier was found to be moreefficient when operating at lower discharge pressure (150 bar).• Cooling reduces the CNG discharge temperature and the cylinder wall temperature of thesingle-stage intensifier . However, it does not significantly improve the single-stage intensifierperformance.• Intercooling can reduce both the CNG discharge temperature and the cylinder walltemperature . It has been shown that effective intercooling results in higher isentropic efficiency.• Intensifier power consumption and isentropic efficiency of both intensifiers were found to bevery sensitive to changes in intensifier speed . Friction due to side forces and excessive heatexpansion (as a result of the side loads) appears to increase with higher intensifier speed .153Therefore, the isentropic efficiency is lower at higher intensifier speeds . Perfect radialalignment of all moving parts was found to be vital to ensure an efficient and reliable operation.• The orifice size of the intensifier valves was found to be an important performance parameter.Cylinder pressure data shows that small valve orifices can cause excessive valve flow velocitieswhich result in cylinder back pressures and low volumetric and isentropic efficiency, particularlyat higher mas flow rates . A bore-to-valve area ratio of 16 :1 or less was found to be adequate forthe required mass flow rate.• Capacity control remains a major challenge for the crank-shaft driven intensifier prototype."Bypassing" and "Valve Unloading" have been simulated on the single-stage intensifier."Bypassing" was found not suitable because of excessive flow velocities through the valveswhich can cause back-pressure in the cylinder and, thus, creates a power loss . "ValveUnloading" causes only a modest pressure rise in the cylinder. However, the physicalarrangement of an unloading device and a potential increase in clearance volume are problemsnot yet solved at present . Capacity control becomes increasingly difficult for the two-stageintensifier with the above mentioned methods because the stage pressure ratio may not bemaintained . Variable displacement and frequency control appear to be the most promisingsolutions.• With a crank-shaft drive, high piston forces give rise to high normal and frictional forces inthe rod guide or crosshead which can generate excessive frictional heat . Because of theseproblems and difficulties with capacity control , the reciprocating crank-shaft driver can not berecommended as on-board intensifier actuator . However, some intensifier components haveproven to operate efficiently and reliably and can be recommended for further intensifierdesigns. These components include:* The Parker PolyPak piston seals (PTFE)154* The V-Packing rod seals' (PTFE)* The inlet valve configuration of the first stage (Version 2 .1)* The discharge valve of the first stage (Version 2 .1)* The valve poppets (brass) used in the discharge valves of single-stage and two-stageintensifier* The Permaglide® bushing as means of eliminating radial misalignment of the rod* The floating rod attachmentThe following can be concluded from multistage compression:• The volumetric efficiency of the two-stage intensifier is significantly higher than that of thesingle-stage intensifier at the same operating conditions.• The isentropic efficiency was not improved with multistaging.• The operating temperatures have been greatly reduced with multistaging.• The incoming gas of the second stage appears to be heated by the large, hot rod surface duringcompression and expansion . Cylinder pressure data show that the expansion coefficient of thesecond stage is significantly lower than k (nexp = 0.7) . The consequence of this was that themeasured interstage pressure differed considerably from the design value predicted with equalpolytropic coefficients for compression and expansion.6.2 SuggestionsBetter intensifier performance and higher efficiency at low intensifier speeds are evident fromthe test data . It was found that eliminating side loads and ensuring perfect radial alignment of'Minimum : 5 V-rings + male and female adapter155moving parts becomes increasingly difficult with a crank-shaft-driven intensifier at higherintensifier speeds . Excessive wear can be expected . Furthermore, the capacity control and off-design performance of a crank-shaft-driven intensifier are serious concerns . Due to high cyclicpeak torques, a very strong mounting and a flywheel may be necessary to operate a crank-shaft-driven intensifier on a bus.The above mentioned difficulties render the hydraulic driver configuration more promising thanthe crank-shaft driver . A double-acting hydraulically driven intensifier can be designed tooperate at a maximum piston frequency of 30 cycles per minute . The control of a variabledisplacement pump can adjust the piston frequency, and thus the intensifier capacity, accordingto the current tank pressure . Efficient off-design performance can be expected because of lowpiston speeds and no excessive mass flow through the valves . Although the hydraulic driverrequires more space, little additional space for the hydraulic pump is required at the engineoutlet. The location of the intensifier can be selected according to availability of space in theengine compartment of the bus . Initially, the cost of a hydraulic driver was a concern.However, older transit buses are upgraded with power steering and the hydraulic system can bepartly shared with the intensifier driver . Availability of a low cost variable displacementhydraulic pump greatly reduces the estimated cost (see Appendix A5) of a hydraulic driver.A hydraulically driven intensifier design is proposed, similar to what has been described inChapter 2 .3. All components used in the two-stage intensifier as described in this thesis can beadapted to the double-acting hydraulically driven CNG intensifier. An arrangement of twocheck valves (CV) and a solenoid valve (SV) shown in Fig . 6 .2.1 allows to operate theintensifier variably as single-stage or two-stage compressor . The operation is described asfollows :If the CNG tank pressure is higher than 60 bar, a single-stage operation appears feasablesince the pressure ratio does not exceed 3 .3 :1 . The solenoid valve is closed and gas iscompressed in both compression chambers and fed to the accumulator, seperately . If the156tank pressure (monitored by a pressure transducer) decreases below 60 bar, the solenoidvalve opens and gas is compressed in both compression chambers subsequently . Theintensifier now operates as a two-stage intensifier . An adequate stage volume ratio canbe achieved by sizing the rod diameter appropriately.DOUBLE-ACTINGHYDRAULIC CYLINDER DOUBLE-ACTINGCNG INTENSIFIERSWITCHVALVEVARIABLEDISPLACEMENTHYDRAULIC PUMPI1.SVCVMICROPROCESSORCNG TOINJECTORCNG-TANK6Figure 6.2.1 : Proposed variable displacement, hydraulically-driven reciprocating intensifierprototype, variably operated as single-stage and two-stage intensifier.The advantage of variable staging is the fact that the intensifier operation can be adapted to thecurrent tank pressure to achieve maximum efficiency . At low pressure ratios single-stagecompression is desirable because the displacement per cycle can be increased and the pistonfrequency can be reduced . At high pressure ratios a two-stage intensifier has the potential tooperate more efficiently than a single-stage intensifier and the operating temperatures can bereduced. Furthermore, advantage can be taken from intercooling .157Chapter 7References1	John B. Heywood, "Internal Combustion Engine Fundamentals", McGraw - Hill, 19882	Gunawan, H ., "Performance and Combustion Characteristics of a Diesel-Pilot GasInjection Engine", M .A.Sc . Thesis, Mech. Eng . Dept ., The University of BritishColumbia, 1992.3 Beck, N.G., Johnson, W.P., George, A .F., Petersen, P .W ., van der Lee, B . and Klopp, G .,"Electronic Fuel Injection for Dual Fuel Diesel Methane", SAE Technical Paper 891652,1989.4	Gordon J . Van Wylen, Richard E . Sonntag "Fundamentals of ClassicalThermodynamics", 3rd Edition, ISBN 0-471-80014-7, John Wyley & Sons, Inc ., 19855	Hill, P. G., Peterson, C. R ., "Mechanics and Thermodynamics of Propulsion", SecondEdition, Addison-Wesley Publishing Company, 19926	Technisches Handbuch, "Verdichter", 4 ., unveranderte Auflage, VEB Verlag Technik,Berlin, 19861587	A . W. Loomis "Compressed Air and Gas Data", Third Edition, Ingersoll-Rand Company,Washington, NJ ., 19828	Schrader Bellows "H-2 Series Heavy Duty Industrial Interchangeable HydraulicCylinders", Brochure CYL-PH2, Akron, Ohio, 19879	Donald R. Askeland "The Science and Engineering of Materials " , PWS-KENTPUBLISHING COMPANY, Boston, Massachusetts, 198410	"SWAGELOK" Catalogue11	Ouellette, P., "High Pressure Injection of Natural Gas for Diesel Engine Fueling",M .A.Sc. Thesis, Mech. Eng. Dept ., The University of British Columbia, 1992.12	Hill, P.G ., Pierik, R .J. and Hodgins, K. B ., "Intensifier-Injector Technology", US PatentNo. 5 067 467, 1991.13	Yuen, D., Hodgins, K .B., "Data Acquisition System for Alternate Fuels Engine Testing",Unpublished Report, Mech . Eng. Dept ., The University of British Columbia, 1991 .APPENDICES160APPENDIX AlTHERMODYNAMIC PROPERTIES OF CNG:The following paragraph shows the derivation of the thermodynamic properties ofnatural gas. Since the composition of natural gas varies throughout the continent(particularly in terms of CH4 content), a specific configuration had to be selected for thecalculations. Here, two natural gas compositions have been investigated and compared;a standard composition ("STANDARD") and the natural gas composition of BritishColumbia ("BC HYDRO"). For the calculations the values of BC HYDRO have beenutilized.~'y{	'	•.;~w:.v:: vv	: .~~t .f	~ ::fv..	.:..fv.	r`}F{~f••v/	4 r	f	~%.~r%:{1~••`S<<x{}~ f r	f .•,~.	f};%r•	f ff~`f~~ :i	p	{f }w~f f flffri	~p~{	/s;^y~~	f'1f.: f. f %	Ir / ~	•• "`^ }	f}J f	~'	:\:t•}:?ff..ff, f :: rf :	f ri i"+.'•:+fffr'~+•f{f'• r:f~:+• .{?~rf`	f.•:~' f	fif	frF:ri'~ f ~'::	r rf.7~. ,•.'?rf ,	f r/ /f'.,	r/te;i./}:	f/{?:. .fc~ffl3~'}	'':`r,,f.f	r. ., . ::.fl.•:{sr?•: .:.:: .:.r .•/,{/.f..:r: •./.r::•:::. :::•7.~Q::~	. . . .v	a	.xr. . . r	 :r•rr:METHANE	CH4.~	..y ./:'~:e~~'	: .{ irf{`	•f ::	f ~fF~f +~~r~} F	1+vn	r:	•r	vi?•'fr ff :?+f'::!•r:?•i	ri ;}f ffffffrrr. :f .	f. .8:{fy::.iw.: ::::: n?:.:::~. . ~/f frf.•. f.f ..	{	f	 ::f :f~	;.f/'p'{	r	Ir/	}• f ` :}}f'rf,.f••'r'r}.•.'••'	••..	iN!:If r	f:%: '	f .~f"r ,"•'.",..•fi••	 0%:•"::;••1.	%?rp}if%f.•: J:f v'~ri	}::•f''	#. ;	.•y,: f }?.: .;%J~fi/ .rs ?.: .~y/,	}}ff	 ;s	} :•iz~: :.:: :.t:::	•'/rF~•{•`•":.'+i/	~w:::: :ry::}::}:•}i::ri•:?rr/.•.}p :}'fffi•~:4.~.v; /.::: :: r::::::.	r nr ..93.8`r{{'~4f~F	:,1++':r:+f l•}'F	'F: }••	r	:r	f }r{{' ''	~•	firf	~f/fr	rF	~/	f%{~~	ff~f.	•' :ff f:	f~ :.:f .:'}~•' .~i:^rf	//Yf r+:	r{F.:f}	•f}ffi:',	frft: . j	i	%`f;~.i. ~f .f~	i '+ .f:::  }:ffv ••:	.G% F'{ :fir/ :/ f/fff• !f {fff/xf{:./rr`f{c:f	fr	.,/,	, f {/'/ff?`//	. .	f': , ff..ff	fr.'rl?:{•}fF+•.y~+f' :. . . .? ii	f/f . . is}r~:{• ./T•::::.:	~i. ~':••}.'Fi:•iii: iF:?ri.rr.96.4	+/- 0 .3ETHANE	C2H6 3.6 2.5	+/- 0 .3PROPANE	C3HR 1 .1 0.2	+/- 0.1BUTANE	C4H 10 0.3 0.07 +1- 0 .04CARBON DIOXIDE CO? 0.2 0.20 +/- 0.1NITROGEN	N7 0.9 0.53 +/- 0.04OTHER 0.1 0.1SUM 100.0 100.00Table A1.l : Composition of Natural Gasgi = %vol * (Mi/M),`: .: .}rr.}:.:.;::.v::?y	~yj}% i	~ ~	T6	. F	~. ..~•• 'll		••:	isJ. . :.:.6. :}:}?•:,	..:•?:r-: f?r~[::}:~~'~`''+1i	~ ~{ f:•.	~; .•.	•r.,l „	%ff:!f fff.:..	%: i f	~.rv' fff.	fi{f+"}'f'/{::%''•. },.&I}r  r ir4 .rf	i .~?.f:}~:.}}}{?}}::is}:•}"•}:iiiii}i:? :Y' :•}:"f.•Fri ::v: jFn . .	:iiv •~~f?x~fr?l r.4~~•~€~•~:F•.}r•}r	:4'~:f•'•:>i::}r•.%.`x;:x••'/}/n//.•:. ,:. . .f::'.?•?..;?. .	•.}	F .~:•}'+ff::::: •vv.•r:f:f :v:f/. i	r?rf ::,~S:rid::•:}".:'{.iT. .S'}'}}Y::::.F'f	. : fff:•:r. `•f.[. . . A•I~`~f~.~/;`'.?r:fff;f"	;:rli'~}f.. .l}~i~ : "A' .;nF . . }Y .::	 %ff :.Ffr :.. :: ,f.~sfar~• .	~f•. rf	~/?.ff%r ~{:://..: ;'•,~ :+fi'f:5ff>~}.+.} lrf~:? f	:j	QQf.~i ff : •: }:. :!	f.F"''	6 i	~'f •+F`•, }	{f yt}r:;:'rfrr "f.; •~f..v:.; :: r>:<•f.}	.if~~ %	:f. /f•~f ffi' fffi?fF~f~L~	fi •	rf,`.:fr~`~::	:,~f ::'Irfrf`/.Fi . . . }:5£f ,l,.?;.;i+ . f::•:'Fr.:	5 ~iy.CH4 0.8784 0.9316 0.938 0.964C7H6 0.0632 0.0453 0.036 0.025C3HR 0.0283 0.00531 0.011 0.002C4H1(l 0.0102 0.00245 0.003 0.0007CO? 0.0051 0.0053 0.002 0.002N7 0.0147 0.00895 0.009 0.0053H7O 0.0001 0.00108 0.001 0.001T_L1_ • • w 1 _ _ 1 1 1coomposiitIon of Natural as161':f.I.':	'• MA:: . . ''•}}fF,.:•i}}'•}i:fi?•:fi f'::i:i}ff'':f'~ :	/+S.~fr:r:8f/F:.; r:.r. f :. ::. . . . ;~ru F/:: }yF• , ..	,/	Ff:ff	;%	ff ~'••}	/f,'%	i `:'~JT F~,f /` IF/.~J:: ii:.i>	•r 'fM .'•:' :J .:riJ[rfiF. ;:n••f'%$	ry'?'if Ff}{. ::l:ix•. :::::./-:::::-y/%::::::':,•ff	k : .	f ..fif	: : +fi?s'. ::}.f i``g••'' ''••:'#	 '. .~i''.''.'.'•`::L. :	:.. .Ff IF	{+:.	ff. . n ..i.}ffF: :i?' ~i•}iii~.C::::F"''"'; :;'i'	f o. .. •} ",¢.f..	a	.•.` :T•"yr'%lF:	//'•	.<;fi?f	F //~Y :iI F ~	Ff/F k y	1~	`F{N~~~g}omI,•,,,,~~F••~+~F/.•:`~bFr`iS<i:v;{	:f<•.	•f{f f.. rv; .; F. 1r,76,. .	~~qf• ; .f ~~•,.•;.:~; .•vfr : ri:. n F :rh.0 r/{Y.•. :: ~ . . ; r f.}ff	.y. ..r	/	r	.:	Fa	i	;	+ lr •' _,	}/,,	v .,t,;~'f I t ;	t..	. R •	~:•' L~,	„.•:, ,	.V ,- f{t+l?~r ~$:~ .	i5:1Yf,:•r~%i%'F	~J'	. F	/	+r	v/F.	. f~.. . :~f. . :~f~r«4}rl,.:C A 16.04 15 .04552 15.46256C7116 30.07 1 .08252 0.75175ClHR 44.097 0.485067 0.088194C4H n 58.1243 0.174372 0.040687CO?, 44 .01 0.08802 0.0880228.02 0.25218 0.148506HBO 18.02 0.01802 0 .01802SUM 17.14569 16.59773Table A1.3: Mass of Natural GasMOLECULAR WEIGHT OF CNG:STANDARD :	M= 17.14569 kg/kmolBC HYDRO :	M= 1659773 kg/kmolINDIVIDUAL GASCONSTANT FOR CNG:	R= R*/MSTANDARD :R*= 8.31434 kJ/(kmol *K)R = 0.484922 kJ/(kg*K)BC HYDRO : R = 0.500932 kJ/(kg*K)DENSITY: p = p/(R*T)ATMOSPHERIC PRESSURE :	p = 101.325 kPaTEMPERATURE:	T = 293 KSTANDARD:BC HYDRO :p = 0.713142 kg/m 3p = 0.690351 kg/m3162CNG CONVERSION FACTOR:n%F. r,~,', .,'r','F:i::i'~•f	~dii+;:{%	:~~:,~~vv::i\}:' • { +	•':'Fit:r:L:1~	(	 r}rx :fF.	0.. :.::.:	+ri :y{::~:? }r~.t~'{~~ :.}.rf.. :J'{:{•r•.	~?41S<% fiii`: :Ff i:.~'~• : . .~~ . ..f+: .	. .~?{	. : ::.:•.•v.{r	}i}:`fi~:i'i.+'••i::f~$:	fii}?`{{ }:v:.•	`Y l :{C:~iF+ii r	/.'tr. ;'~$ .}{,.,.: v:~.Y///~•:~.•:f.~.~►~~~.i~{•	 /: :rfv::: f..:: ?i::::::::: :: : : : :::::::•::::::	!i	/i.'i:f.{•i?::~/. f{ffi:~:.::h}:::.~::..::{{:: :ri::{.~.::.v:•:]ir:.•%r'~':irr~:. .::. v{.:. .::i! f	: ` /fifv{nr.	' :.?•rf.~ ,. .?vv.,,.+rfi.	; n . r. x• . ; . r.. . . . . . . . . .$?M~ . . . . :. . .	. ..:.fi:::::?:.:w::::::::	:.: ff.::{:~:.. fii .~:::f • ::ri' ''f, ~•	~;.fJf:•.:fi	t '.:i//+{ii>:.~`~	'+•~ :`.~!'A . . . .	{`i':	 :i :+:rf r'i`v ::	 :. ..•f. .~:tip/v:}:::::::}::::::::::: •:n::n:r ffw::::::::::::::::.~:?v:.ff:.~',~.}}:Cn 'r'. .: +?~lC~~ .t~~7`?`Ai.nr?G	'+f:	f.. ~{fr`f:r~fr~£i~'iiri'v:::::f:'i{'i::'iif4::r: x: ~: ."~'~i?> "~} .:	 ~CT.:r,n . .. .	: 'Fri:F	~ 9	E	~ :::fi:$yiiii~:;::iF }:•ii}F}:•}}r•}}i. :• ~	+	~ 9CH4 0.8784 0.9316 0.72 0.632448 0.670752C2H6 0.0632 0.0453 0.5 0.0316 0.02265C3HR 0.0283 0.00531 0.36 0.010188 0.001911C41110 0.0102 0.00245 0.25 0.00255 0.000612CO2 0.0051 0.0053 0.74 0.003774 0.003922N2 0.0147 0.00895 1 0.0147 0.00895H2O 0.00108 0.00108 1 0.0001 0.00108SUM 1 .00 0.99999 0.69536 0.709878Table A1.4: Gravitation Factor of Natural GasCALCULATION OF LHV:Table A1.5: Lower Heating Value of Natural GasSTANDARD: LHV = 48707.84 kJ/kgLHV = 34735.64 kJ/m3BC HYDRO: LHV = 49098.47 kJ/kgLHV = 33895.19 kJ/m3CH4	0.8784	0 .9316	50010	43928.31	46689.31C2H6	0.0632ClHs	0.0283C4H ~ 0	0.0102SUM	0.0453	47484	3000.988	2151 .025	0.00531	46353	1311.789	246.1344	0.00245	45714	466.2828	111 .999348707.84	49098.47163APPENDIX A2ENGINE SPECIFICATION DATAThe engine specification data was provided by the Detroit Diesel Company . The DetroitDiesel 6V92-TA two stroke engine belongs to the group of heavy duty engines . Itconsists of 6 cylinders in V-configuration with a displacement of 92 cubic inches percylinder. TA stands for turbocharged and aftercooled . Out of a variety of models, thestrongest engine has been investigated (285 hp) to represent the case of maximum fuelconsumption and, hence, maximum compressor capacity.GENERAL DATA :6V-92TA COACH64,84 x 5,00 (123 x 127)552 (9 .05)17.0 : 11740 (533)not applicable4direct injection63 .5" VEE - 2 CycleturbochargedModelNumber of CylindersBars and Stroke - in (mm)Displacement - in3 (L)Compression RatioPiston Speed - ft/min (m/min)Valves per Cylinder:IntakeExhaustCombustion SystemEngine TypeAspirationPHYSICAL DATA:Size:Length - in (mm) 37 .6 (955)Width - in (mm) 37 .6 (955)Height - in (mm) 48 .0 (1220)Weight, dry - lb (kg) 2020 (920)FUEL SYSTEM:Fuel Injector Part No ./Timing 5234850/1.460Cart Code 0002Fuel Consumption - lb/hr (kg/hr) 103.5 (46.9)Fuel Consumption - gal/hr (L/hr) 155 (58 .7)Fuel Spill Rate - lb/hr (kg/hr) 489 (222)Fuel Spill Rate - gal/hr (L/hr) 73.2 (277)164Total Fuel Flow - lb/hr (kg/hr)	593 (269)Total Fuel Flow - gal/hr (L/hr)	88.7 (336)COOLING SYSTEM:Coolant Flow - gal/min (LJmin) 160 (606)Maximum Top Tank Temperature - °F (°C) 210 (99)Minimum Top Tank Temperature - °F (°C) 160 (71)Minimum Coolant Fill Rate - gal/min (LJmin) 3.0 (11 .4)PERFORMANCE DATA:Power Output - bhp (kW) 285 (213)Full Load Speed - r/min 2100Peak Torque - lb ft (N m) 870 (1180)Peak Torque Speed - r/min 1200BMEP - lbf/in2 (kPa) 92 (1200)r{ '::•:. .:: f%f!` . J}} }~ }/• .•• '.•,`~~,•'r.,' -.•~. .	}K 9 ;r!	/ . .r~~/~'~f ~	.4• •6: ..&	b	S~$ t :'!	,4	f ".. :~•'%f'•~':	~ ~	~ bT	$ ]fl > ` . :;KV~:~rml:,.ffiv i:' ~'	~f}}.:fN~ a! •ff:/~~ . :.	iPf ~3`~://fem. •.,/~/ ~!	f : ` ~::r;Sf{f+f"i}'l..9	0	9	!•. :'•::Fj:?	.::.• tir'f~l~ ~i i G :• r~u: ff~~f?if ::f~	r~r` ~ /fi3' f~ ~,~f~.f	r	! f. . . rff. .{}:	•••~~~{fS: ~'''~:.	•'•~flf	f~ff.~~{i:!/	rf'::{i' ..+% i/r:	!ieY {l~~r~.	:f • • •f•'	rf:~f~r.F/.	.:••~.;:• : :	Jx6~ rd.•.~1~ r.Yi:+fY.i'	r4	~	'f ,	.'f	.+Y'i	!~ Y. .. :.f .9,•`.+}r~ !!7 :	' E	S~i`	••~fr$"	:i J/	1rSk :: C	!>'.F ; 6 F~,!i	T f	n~ }~:::f • :ffffr:: :rr: i:f.	rf •'• .r	Fr.ff	.flF~• -f'f: hr . f : ~ .J•r~	off iYff. •.C: firf/~..•.~~~?	{ [..	.Cf.	+. . .	{ .~:,{r• Y :.2100 285 (213) 713 (967) 0.358 (218)1950 276 (206) 743 (1007) 0.355 (216)1800 265 (198) 773 (1048) 0.350 (213)1600 248 (185) 814 (1104) 0.341 (207)1400 228 (170) 855 (1159) 0.335 (204)1200 199 (148) 870 (1180) 0.334 (203)1000 161 (120) 846 (1147)	_0 .343 (209)Table A2.1 : Performance Data of Detroit Diesel 6V92-TA Engine.The information of power, torque and brake specific fuel consumption (BSFC) wasbacked up with a graph. The values for power, torque and BSFC have been deducedfrom this graph in 50 rpm intervals . Since data are only available down to 1000 rpm, thecurves had to be extrapolated to cover the lower speed range between 600 and 1000 rpm.This was done by fitting a polynomial to the existing data whereP = ao +a1 •rpm+a2 rpm2 +a3 •rpm3BSFC = ao + a1 rpm + a2 • rpm2 + a3 • rpm3 + a, • rpm'165Based on the lower heating values, the diesel fuel consumption (BSFC) has beenconverted to natural gas.Figure A2.1: Power output curvePOWER OUTPUTDETROIT DIESEL 6V-92 TA 285 HP250200X 150500	1000	1500ENGINE SPEED [RPM]	DETROIT DIESEL	FITTED CURVE0 500 2000 2500BRAKE SPECIFIC FUEL CONSUMPTIONDETROIT DIESEL 6V 92TA 285 HP0	500	1000	1500	2000	2500ENGINE SPF11) [RPM]. DIESEL-FUEL DETROIT	FITTED CURVE	CNU-FUEL300T 2503a150Figure A2.2: Brake Specific Fuel Consumption curve166The lower heating values for natural gas has been taken from Appendix Al (BC Hydro).LHVnatural gas = 49098 kJ/kgThe lower heating value for light diesel [7] is:LHVdiesel = 42500 kJ/kgTo get the maximum fuel consumption in kg/hr for the entire speed range, the BSFC(kg/kW hr) has been multiplied by the regression data of the power output (kW).CNG FUEL CONSUMPTION (MASS)DETROIT DIESEL 6V 92TA 285 HP1500	2000ENGINE SPEED [RPM].~ _ kg/hr — g/revFigure A2.3: CNG fuel consumptionThe fuel consumption has then been converted to volume, using the density value ofnatural gas calculated in appendix Al (BC Hydro)167CNG FUEL CONSUMPTION (VOLUME)DE;1ROIl' DIESJL 6V--92TA 285 HPwoo	1500ENGINE SPWI) [RPM]20000 500Figure A2.4: CNG fuel consumption in normal cubic meters per hour (Nm"3/hr) andstandard cubic feet per minute (SCFM)168APPENDIX A3SAMPLE CALCULATION: CENTRIFUGAL COMPRESSOR:The purpose of this calculation is to calculate the tip diameter as well as rotational and tip speedof a centrifugal compressor, assuming an efficiency of 85%.Figure A3.1 : Comparison of calculated efficiency contours with test data on centrifugal pumps(5] .400.5- Calculated--- Tested0.50 .60=0.6 0.7313,=o'15° ~0 .80 .850=0.850 .70 .08IRe * - 10 8f0 .9I°0 .9I20106420 .06 0 .08 0.1 0 .6	0 .8 1 .00 .40.2nsX and Y axis of Figure A3.1 are defined as follows:169and	ds=_ SZ• IXns	30(4) 4Pwhere:	S1	 rotational speedQ	 gas volume at inlet conditionA po	 pressure differencep	 gas densityD	 rotor tip diameteri) CNG Inlet Density:p _ 2 . 106	kg/P R T 500 .288 -13.889 m3ii) CNG Volume at Inlet Port:CNG mass flow :	Q = 60 m3/hrQ = 0.01667 m3/seciii) CNG Pressure Difference:APo= Pout- pin 200 - 20 =180 barApo= 18*106 Paiv) Rotational Speed:ns was chosen from Figure A3 .1 for - 85% efficiencyhence, ns = 0.5Y Z3n, - (40.5 ( 18-10 Tsand	1=	 13.9/0.01667= 148,660 rod/sec170v) Rotor Tip Diameter:ds was chosen from Figure A3 .1 for -85% efficiency,hence, ds= 5D = /0.01667= 0.0191 m18 . 10613.9The rotor tip diameter should be 1 .9 cm.vi) Tip Speed:Ut= .Q*D/2Ut= 148,660 * 0.0191/2Ut= 1422.5 m/svii) Shaft Speed:n	 -60U, - 60 -1422.5D•,r	0.0191 . 7rn = 4.468*10 6 rpm171APPENDIX A4PERFORMANCE CHARTS OF COMMERCIALLY AVAILABLECOMPRESSOR UNITSIn the following section performance charts of piston compressors are presented. BothNorth American and European companies have been contacted. In particular, thesecompanies are :	• Sulzer (Switzerland)• Neuman & Esser (Germany)• Atlas Copco (Germany)• Leobersdorfer Maschinenfabrik AG (Austria)• Ingersoll-Rand (Canada)Most of the compressors that deliver gases at 200 bar discharge pressure are aircompressors. Their capacities are usually much larger than required . Smaller air cooledcompressors which generate 200 bar consist of 4 or more stages since air is drawn fromthe atmosphere. The conversion of these compressors is questionable, since a great dealof modifications must be undertaken . Most compressors are designed for one or morespecific capacities at a fixed inlet and discharge pressure rather than a whole range ofpressures and capacities . They lack a suitable capacity control method . Themodifications add to the general high cost of a compressor.Thus, it can be assumed that a newly designed intensifier is less costly and more suitablethan the modification of a commercially available unit .172200100602010610	m 3/h	50	100	200	500	1000	2000	5000	10000111	III	 N2=::~~~s''2 911	 ~'	 nnn~1~~1•~1 i 1__I ilull"a oast aD2~i-2A $ 4D 28 viiir	nn~	n 1O .IIIIIIIumn1111111n1 1n 01 nIklh	=• I nII 1woman 1LIEN 1D 30-tA 1 2D200-18	,,,"') IVv 1 1 5 f '11	1 D2-i'Illill1D65-1AFigure A4 .1 : Selection chart of Sulzer "Labyrintpiston"-Compressors .1737030201042010865432I v	 -. ..JI%V- Bauart	Baugrol3elV 1V 2V 50V 80-)BauartenstehendliegendBoxerBaugroBe125200320305080Ii`i 1t BauartenI BVstehendliegend-L---~---~-...I~ II_I-1r-ter WI&.O•s1	i	I I500	1000	5000	10000 20000Figure A4.2 Capacity chart of Neuman & Esser piston compressors.m3/h 100174Gas-KolbenkompressorenSonderausfuhrungen bis 300 bar im Trockenlauf300 -200 -150-100-f 50-40 -6" 30 -cC 20-10-8-6-5-4-3-2-I	i40 50 60 80 100	200 300 500	1000	2000	5000	10000 20000	50000 100000Liefermenge bezogen auf 1 bar Saugdruck —)-SkaFigure A43: Capacity chart of Atlas Copco dry-running piston compressors .175Leistungsdiagramm der Iuftgekuhlten Hochdruckkompressoren450 (6525)400 (5800)350 (5075)300 (4350)250 (3625)200 12900)V119150 121751V23 V 26100 (1450)V 770 (1015)50 (725)30 (435)!V13V 15	V 16	II 8V 1420 (2901bar (PSIG)I /min (CFM) 100 I150	200	300	400 500	700	1000	15	2	3	4	5	7	10 00013 5)	15 .3)	17 1)	(10.61	(141) (177)	(247)	(353)	(53)	(70.6)	(105 .91 (1412X1766)	(2472)	(353 .2)Figure A4.4 : LMF air-cooled high pressure compressors .176Engineering DataMotor Dimensions WeightsMax. Prase. P. D. Rating Length WWth Haight Ban	Bass Mtd.PSIG Bar CFM Mirth HP KW Model RPM IN MM IN. MM IN MM Lba. Kg. Lba. Kg.500 34.5 7 .4 0 .21 2 1 .5 231 660 36.0 914 20.0 508 19 .0 483 150 68 290 132500 34.5 36.0 1 .01 10 7.5 772 800 52 .0 1321 24.0 610 31 .0 787 425 193 890 404500 34.5 49.5 1 .40 15 11 .2 1572 900 59 .0 1499 28 .0 711 34.0 864 725 329 1180 535800 55.2 30.3 0.86 10 7 .5 1572 555 59.0 1499 28 .0 711 34.0 864 725 329 1150 5221000 68 .9 41 .2 1 .17 15 11.2 1572 750 59.0 1499 28 .0 711 34.0 864 725 329 1180 5351000 68 .9 13 .1 0.37 5 3.7 774 800 52.0 1321 24.0 610 31 .0 787 425 193 890 4043000 207 11 .0 0 .31 5 3.7 223 985 42.0 1067 21 .5 546 28 .0 711 310 141 500 2273500 241 36 .8 1 .04 20 14.9 1574 1000 58.4 1483 27.0 686 45 .5 1156 780 354 1145 5195000 345 23.2 0.66 10 7 .5 H15T4 631 58 .4 1483 27.0 686 45 .5 1156 780 354 1265 5745000 345 32.7 0.93 15 11 .2 1-11574 890 58 .4 1483 27 .0 686 45.5 1156 780 354 1295 5875000 345 36.8 1 .04 20 14 .9 H15T4 1000 58.4 1483 27 .0 686 45.5 1156 780 354 1320 599Figure A4.5: Ingersoll-Rand high pressure, air-cooled piston compressors177APPENDIX A5HYDRAULIC DRIVE PRICE BREAKDOWN'All parts as shown in Figure 2 .3 .1 . This configuration represents hydraulic parts onlyand does not include the intensifier parts.Hydraulic Pump (variable displacement)1 .55cu.in./rev max. displacement	$ 1,431 .00Remote Pressure Controller	$ 251 .00Proportional Valve	$ 680.00-Subplate $ 340.00- Control Card	$ 814 .00Directional Control Valve	$ 300.00-Subplate $ 230.00- Auto Cycle Valve	$ 671.00Hydraulic Cylinder (with cushions)	$ 724.00Turbine Flow Meter	$ 645 .00- Signal Conditioner $ 850.00- Digital Readout	$ 738.00-Cables $ 100.00Return Line Filter	$ 60.00Sundries	$	 250.00SUBTOTAL :	$ 8083.00 $ 8083 .00EXTRAS:24 VDC Power Supply (2 .4 V (3 24 V)	$ 113.95Potentiometer to Control Proportional Valve	$ 55.0010 Gallon Reservoir (complete with suctionand return line)	$ 575.00Hydraulic Oil (per litre)	$	 2 .30SUBTOTAL :	$ 793.00 $	 793.00TOTAL :	$ 8876.001 Source: Janox Fluid Power, Richmond, B .C., Canada, March 1991178APPENDIX A6SAMPLE CALCULATION: HYDRAULIC DRIVERThe calculation shows the dimensions of one hydraulic driver configuration and howthey have been derived . This is only one of several solutions.Conditions:Tank Pressure : P1= 20 bar (300 psi)Injection Pressure : P2= 200 bar (3000 psi)Fuel Consumption : m= 45 kg/hr @ 2100 rpmGas Cylinder Dimensions:Bore :m= 15 kg/hr @ 600 rpmBg= 7 .62 cm (3 .0")Stroke: Sg= 30.48 cm (12.0")Rod: Rg= 5 .08 cm (2 .0")Clearance Volume : co = 3% (for forward stroke)eo = 5% (for return stroke)Bh= 8.89 cm (3 .5")Hydraulic Cylinder:Bore:Stroke: Sh= 30.48 cm (12.0")Rod : Rh= 5.08 cm (2 .0")Gas Properties:Gasconstant R= 500 J/kg KPolytropic Coefficient n= 1 .3Inlet Temperature T1= 288 K11 Volumetric EfficiencyA, =1-so Pz -1 - Pz - 0.011	Pl1=1-0.03 (200) 11 .3-1 -10 . 0.01=0.7536the volumetric efficiency for the forward stroke is 75 .36%.1791200 1 .3R=1-0.05 (—20	-1 -10 . 0.01=0.6561the volumetric efficiency for the return stroke is 65.61%.2) CNGMassflow perstroke(forwardstroke):volume:VF = Bc 1r ScVF =7.622 •7r 4 . 30.48	VF =1390cm3masssflow:m=P1VF-A vR•1m20 . 10 5 . 1390 . 10 0.7536500 . 2881nF =1.455 .10-2 kgstroke3) CNGMassflowper stroke (returnstroke):volume:VR=(BB-"' it	SGVR = (7.622 - 5.082 ) . 7r 4 . 30.48	VF=772 .2 cm3massflow:m=P1 • VR X vR . 7j20 . 105 . 722.2 . 10 -6 . 0.6561m =500 . 288180mR =0.66 . 102 kgstroke4) Total disulacement Der cycle :,D=Vf+Vr= 1390+722 .2=2112 .2cm35) Total CNG Massflow Der cycle (forward + return stroke):total mass per cycle:6)#of cycles Der minute re quired:i) mass requirement @ 2100 rpm : mreq= 45 kg/hrmreq= 45/60 kg/min = 0.75 kg/min# of cycles= m"Q=	 0.75 3 = 35.46 cyclesmwT 21.15 . 10 -	minii) mass requirement @ 600 rpm :	mreq= 15 kg/hrmreq= 15/60 kg/min = 0 .25 kg/min# of cycles= m"4 =	 0.25	 =11.8 cyclesmTOT 21.15 -	min7) Required numb capacity:assuming that the pump turns with engine speed:the hydraulic cylinder volume is:Vg = Bg•n•4•Sg+(Bq–RA)•n•!•SrrVg = 8.892 • n • 0.25 .30.48 + (8.89 2 – 5.082 ) • n . 0.25 . 30.48V,, = 1891.94cm 3 + 1274.16cm3 = 3166.1cm 3kgmTOT = mp+mR = (14.55+6.60)•10 -3= 21.15 . 10 -3cycle181i) Hump capacity for 2100 rpm:Vpump= VH* # of cycles per minuteVpump= 3166 .1 / 1000 * 35.46 = 112.27 Uminpump displacement per revolution:D = Vpump/2100 = 53 .6 cm3ii)	 pump capacity for 600 rpm:Vpump= VH* # of cycles per minuteVpump= 3166 .1 /1000 * 11 .8 = 37 .36 Uminpump displacement per revolution:D= Vpump/600 = 62 .27 cm3A pump displacement of 62 .27cm3 /rev (3 .8 ci/rev) is required tooperate the hydraulic pump at enginespeed. Eventual fliud shortages can becompensated with a hydraulicaccumulator.8) Maximum force on rod:A f = 45.60 cm2Ar = 25.33 cm2P2 = 200 barP1 = 20 barTHRUST:F=AF' P2 —AR -PIF = 86,134 NTENSILE:F,= —AF P,+AR •P2F,s = 41,540 NThe maximum thrust force is 86,134 N .182APPENDIX A7SPECIFICATION OF THE EXISTING AIR COMPRESSOR'Following are the technical data of the air compressor which is used on the transit busses.Its purpose is to power the air brakes .Tu-Flo 700 SpecificationsAverage weight	 46 lbs .'Number of cylinders 	 2Bore size	 2 .750"Stroke	 1 .81"Displacement at 1250 RPM	 15 .5 CFMMaximum recommended RPM	 3000 RPMMinimum coolant flow (water-cooled) atMaximum RPM	 2 .5 GPMMinimum RPM	 5 GPMMinimum coolant flow (air-cooled)	 N/AApproximate horsepower required at1250 RPM at 120 PSIG(naturally aspirated)	 3 .5Turbocharge limitsMaximum RPM	 2200 RPMMaximum pressure (gauge)	 15 PSIG"Maximum inlet air temperature 	 250 degrees FMaximum discharge air temperature	 400 degrees FMinimum pressure required to unload(naturally aspirated)	 60 PSIGMinimum oil pressurerequired at engine idling speed	 15 PSIGMinimum oil pressure required atmaximum governed engine speed	 15 PSIGOil capacity ofself-lubricated model	 N/AMinimum discharge-line size	 1/2" I .D.Minimum coolant-line size 	 3/8" I .D.Minimum oil-supply line size	 3/16" I .D.Minimum oil-return line size 	 1/2" I .D.Minimum air-inlet line size	 5/8" I .D.Minimum unloader-line size	 3/16" I.D.'Installed weight determined by final mounting configuration."Mmdmum pressure (gauge) at maximum 1900 RPM. 25 PSTG.Figure A7.1: Technical Data Sheet of Tu-Flo 700 Compressor'Source : Bendix Air Compressors (brochure)Tu-Flo 700CompressorMP *rm. igawadoubles	heavy tandem	transitconventionalCOMPRESSOR BUILD-UP TIME0-120 PSI6900 CUBIC INCH RESERVOIR400loo1500 1800 2100 2400 2700 3000COMPRESSOR SPEED (RPM)01200183APPENDIX A8SAMPLE CALCULATION : THE "SKIPFIRE" METHODThe purpose of this calculation is to evaluate the power loss due to compression of air fromatmospheric pressure to 8.5 bar (125 psi) in a combustion cylinder of a 6V-92 engine . Thecompression is achieved by skipping the fuel injection and, thus, the combustion and forwardingthe air into a receiver vessel . No turbocharging is assumed.GIVEN:	cylinder volume :	1 .5071(92 in 3 )compression ratio :	17:1engine speed:	1250 rpmpower output @ 1250 rpm:	150 kWREQUIRED: air pressure:	8 .5 bar (125 psi)flow capacity :	50 SCFMCALCULATE:1) clearance volume:	eo = 1/18 = 0 .0555 = 5.55%co = 0.0555* 1507 = 83 .7 ccm	2) volumetric efficiency :	i v = 1 - e0*[8.513 - 1] = 0.767r1 v =76.7%3) number of strokes n required to achieve required air flow:n QV rl qwhere:	Q	air flow (in SCFM)	V	piston displacement (in ft 3)184n	engine revolutions per minuten Q = 92 50	=1224 rpmV . T ly	0 .7671231224 strokes must be skipped to deliver 50 SCFM air @ 125 psi1 - 1250 -1224 = 1 - 0.0208 = 0 .9792 .-. 98%125098% of the strokes are not used for combustion, hence, 98% ofthe power of one cylinder is not available5) absolute power loss :	0 .98* 160 = 24.48kWthis is the power loss only due to skipping the combustion in one cylinder6) relative power loss :	1- 1501504.7 =16.3290the relative power loss is 16 .36% not including the power consumption for theair compression .185APPENDIX A9SAMPLE CALCULATION: HASKEL GAS BOOSTERSThe performance of various booster models is calculated as is suggested in the HASKELbrochure M-26D. Prior to the performance calculations a booster model selection chartis presented. Then the requirements are stated and the calculations for selected modelsare carried out. At the end, a booster combination is presented as suggested by theHASKEL Inc. head office.MAXIMUM OPERATING PISTON DISPLACEMENTTYPICAL GASACTUAL VOLUMEDELIVERYATPRESSURE WITH: MAX.-	PRESSUREMAX . MIN. "P*"AIR DRIVE PRESSUREGAS SUPPLY COM-BASIC DISPLACE- (Air Drive 35 SCFM at 85 PSI)•MODEL RATEDGASOUTLETNORMALGASSUPPLYPRESSURE "Ps" PRESSIONRATIO INNUMBERPERCYCLEIN s•LOADTZMENTERO100 PSI DRIVESUPPLY OUTLET=PI!. Q P319 ACT- @ PSI(SUPPLYPRESS.PSIGPRESS.PSIGPsPSIGOUTLET"Po" (REFERPAGE 81 LITRE CYCLESPER INs/MINBINACi LIT BARMINLITACTBar Bar Bar MIN. LIT/MIN 49MIN @BARMINAG-1 .5 20.7 ATM . 207 1 .5 PA+Ps 101 .31 .9522 215 66 60112 29g 0110 2510 @ 161 5AG-4 ' 400 ATM 600 4 PA 10:1 164 250 2500 914 ,X170 5.74 @ 193AG-7 710 254 1 .7 1050 7 PA 20 :1 216 124 02637 9 .67X170 2 21 @ 33 8= AG-15 ' 2250 50 2250 15 PA 20 :1 6'01 124 126 9 280	20 7 8 5 X1 72 4AG-30 003 1100 ao0 30 PA 25 :1 051 124 3846 29 2.33 @ 34 5 3565 @ 12150AG-62	_600 260 6 p0 60 PA 25:13 .1. 051 SO 4.06 910 Q 690 236	22907' 11,250 250 11,250 1 .2 149 55 .7 @ 1500 16.0 @ 5250AG-75 775 17 775 75 PA 25:1 .020 124 2 .44 .913 Q 103 .262 @ 362Sa, AG-152 2°380o° j5j 23 80 150 PA 25 :1 02 SO 1 . 7 35651038 ~108@~ 724AG-233 22,5001550 25017 22,5001550 225 PA 25:1 1 .2.020 69 841 .38 26 .8 @ 2000.439 @ 138 3 .4 Q 15750.056 @ 1084AG D-4 50 39 6 1240y, 9 ATM 868 l PA + Ps 10 :1 245 7735 4@Q 7~ 7 .10	2620AGO-7 2500 500 7 PA + Ps 20:1 26.4 100 04343 13 81668 @ 9003.70 @172 1 .7 172. AGD-15 : 5000 50 3505 15 PA + Ps 20:1 12.4 100 212043	7 .24 7@	20 102 @ 1350O 1 67AGD-30 6° o° 100 60200 30 PA + Ps 25:1 62 100 10.202 215 a3.52 @ 3545	462 .393 @ 12 700 9000 200 9000 62 428 160 @ 1000 31 @ 5200AGD-62 620 14 620 60 PA + Ps 25 :1 .102 69 7 .01 2 .62 Q 69 .0 .508 @ 359AGD•75 20,0001380 25017 0 .0001380 75 PA + Ps 25:1 2 .4.039 100 2403 .93 92 @ 15001 .51 @ 103 20 .6 @ 6750.338 @ 466AGD-152 20,000 250 20,000 150 PA + Ps 25:1 2.4 69 166 59 @ 000 9.5 @ 125001380 17 1380 .039 2.72 .967 Q 138 .156 @ 862AGD-152H 25,02040 217 0 25,02040 150 PA } Ps 25 :1 2 .4039. 69 2 72 1 08 @6414 393@1 1138AGT-4 1250.86 2 s/ ATM 125086 .2 4 PA +Ps 100;1 10.0.164 250 250041 .0 1035 @ 2517 .0 @ 1 .70 129 Q 3052 .11 @ 21 .0AGT-7/15 4000 25 6PA 15 P A	2 Ps 50:1 13 .2 100 113.6 586@70 49@86 .22 PA PA 100:1.13 .2 63	100AGT-7/30 3j j 30	+ 4 Ps 216 100 2131 .6 47 59@ 7 231.146 @ 172AGT-15/30 8500 3 .5 15 PA 30 PA {- 2 Ps 50:1 102 100 60.2 4 16 Q 207 2482 @ 12786v AGT•15/75 1.897 50 3.5 PA 75 PA+5P5 100:1 602 100 0620.2 293@103 400@ 60004v~i AGT•32/62 9062100 1007 30 PA 60 PA + 2 Ps 50:1 602 69 42087 . 12 .90 @ 690 2474 @ 46220016000 100 3 .1 310 121 @ 500 9 .59 @ 6500AGT . 30/75 1100 7 20 PA 75 PA + 2 .5 Ps 60:1 .051 100 5 .08 1 .98 @ 34 .5 .157 @ .448AGT-32 /152 200001380 100 7 PA 150 PA + 5Ps 100:1 6.2 69 428 127 @ 350 3.78 @ 122507 .102 7 .01 2.08 @ 24 .1 .062 Q 845AGT-32 /152H 25 000 100 7 PA 150 PA + 5Ps 100:1 6.2 69 428 154 Q 700 7.86 @ 140001720 7 .102 7 .01 2 .52 @ 48 .3 .129 @ 966AGT•62/152 1 38080 1007 40 PA 150 PA 12.5 Ps 60:1 3.151.0 69 214.513 8137 (a?1 609	6 .107Q@1 890025 000 100 3 .1 214 89.6 @ 2000 11 .6 @ 15500AGT•62 /152 H 1720 7 40 PA 150 PA + 2.5 Ps 60:1 .051 69 3 .51 1 .47 @ 13 .8 1 .47 @ 1070Figure A9.1 : Model Selection Chart. Maximum Air Drive Pressure : 10.3 bar (150 psi)ty)aOOm021 .00o.soNSICmECo0 .60GCCq 0 .40U-I-Zg 0 .20WL)a• 0II PSIG(BAR)186OUTPUT FLOW SCFM = Kd x Ve x Db x PS + 14 .714 .7~	AIR DRIVE FLOWI s~	(Std . FL3/Min .)P	'Mt'*1 .000020(1A)5010	15	20 25 30	0	20 40 60 80SINGLE STAGE	TWO STAGECr = COMPRESSION RATIO = (Po + 14 .7) (Ps + 14.7)40	60	80	100	120(2 .8)	(4 .1)	(5.5)	(6 .9)	(8 .3)Lp = PRESSURE LOADING ON AIR DRIVE1 .00l	1	lPs - Gas Supply, PSIGPo Gas Output, PSIGPa = Air Drive, PSIGalculate Lp from chart— bottom of page —For Two Stage Models:Note:1. Interstage Lp will bethe lower of the Ps vsPo quotients.2. Total Lp will be thehigher of output Lpvs interstage Lp.100 120140(9 .7).80Ve =	.60VOLUMETRICEFFICIENCY . ACTING SINGLE STAGE	DOUBLE ACTING SINGLE STAGE	TWO STAGEAG MODELS	AGD MODELS	ACT MODELS-11	-11 -M -17 -7s -112 -217 -4	-7 -Is -20 - 2 5 -112	-Ins -7/51 11/20 -15/71 1-32/52 20/71 3V112 127152.-7.Ir,r .-42rH 35r, r. P. ►.0 12.1r .-2.1r,7sso► .20►.-s► ,Iser.71ISO100NP.>,►.-lr,mr. P.11 219.	Am.AIM.InkZ . Lw1 IN PSI On., , LIt dly.Use Vo'on.bin Efficiency Cone2 .71 112 .72	21 24717711 14711 12111 11121 1/ N)A	0	1	C	C	1211 1 01211 ,141810	c	c.222 .144 11112 .511C.1431415C.05117 .121C1.4514115D.711DIMD	(11211141 .1)A1201)1 .46217 .141 .444 110Ass1111) 11 .42C C.247211.711 .2ss .17212 4111014124711 .0210.17111111E121 .51	110 .11	110 .1)D I D	DFigure A9.2 : Performance Data187Figure A9.3: Performance CalculationsGIVEN:REQUIRED OUTLET PRESSURE Po : 206 bar (3000 psig)AVAILABLE GAS SUPPLY PRESSURE Ps : 20 .6 bar (300 psig)REQUIRED GAS FLOW : 35 SCFMPRESSURE RATIO : 10:1AVAILABLE AIR DRIVE PRESSURE Pa : 130 psiAVAILABLE AIR DRIVE FLOW : 75 SCFMnote that both air pressure and air flow have been chosen relatively largeAir or gas flow under the pulsating conditions characteristicof an air driven gas booster will vary widely with relativelyminor changes in external plumbing, internal friction, man-ufacturing tolerances, etc. While it is not possible to makeprecise predictions, it is relatively easy to make reasonablyaccurate estimates by the following simplified method : —Gas flow in "actual" volume per minute (volume at actualpressure and temperature) is piston displacement per cycletimes cycling "speed " (cycles per minute) times volume-tric efficiency . This can be converted to "standard" vol-ume (volume at standard atmospheric temperature andpressure) by multiplying by the ratio of supply pressureto atmospheric pressure, and dividing by the ratio ofsupply temperature to atmospheric temperature . (Allpressures and temperatures must be absolute).PRESSURE LOAD — "Lp" is the air drive pressure required'to produce the outlet pressure at "stall" or zero speed . Thedifference between this "load" and the actual air drive pres-sure determines rate of flow through the air system andtherefore determines cycling speed.Note : This family of boosters includes units with a varietyof stroke lengths and with one, two or three drive cylin-ders . Air drive volume, and therefore cycling speed varieswidely from one model to another . However, ALL modelsuse the same air caps, valve and essentially identical airsystems . Each will use essentially the same quantity ofdriving air for any specific combination of "load" andair drive pressure . Cycling speed for each model will bein direct proportion to its speed at zero load with 100PSI (7 bar) air drive . It is threfore possible to show per-formance by a single graph showing relative speed versusload . For convenience the graph shows "displacement"(speed times volume per cycle).With single acting single stage units the pressure load isoutlet pressure divided by drive area ratio . With doubleacting single stage units pressure load is the "lift" ordifference between outlet and supply divided by the drivearea ratio .	-With two stage units the pressure load on each of the twostrokes must be determined and the higher value used indetermining speed or displacement. On the output strokethe load is outlet pressure minus assistance from supplydivided by second stage drive area ratio . On the interstagestroke the interstage pressure acts on the DIFFERENCE be-tween areas of the first and second stage pistons . The inter-stage pressure is a fixed factor times supply pressure butcannot exceed outlet pressure . Pressure load for the inter-stage stroke is determined by supply pressure, or by outletpressure whichever gives the LOWER load . (See equationsin chart below).VOLUMETRIC EFFICIENCY : — The actual intake vol-ume per stroke is less than "piston displacement" since partof the available space is filled by re-expansion of high pres-sure gas remaining in the clearance volume at the end of thecompression stroke. The volumetric efficiency, "Ve" , is theratio of actual intake volume to the displacement . It is 100%at a compression ratio of 1 :1 (outlet = supply) and decreaseslinearly to zero at a compression ratio such that the re-ex-panded gas completely fills the available space and no freshgas is taken in.The following shows procedure for performance calcu-lations with typical examples.1881) AGD-15 (double acting, single stage)maximum rated outlet pressure : 345 bar (5000 psig)maximum pressure ratio : 20:11) BASIC DISPLACEMENT Db : 20.4 liters/min (0 .72 actual CFM)(from chart)2) PRESSURE LOAD Lp : Lp = (Po - Ps)/15 = 2700/15 = 180 psifrom the chart it can be seen that 180 is too high2) AGD-30 (double acting, single stage)maximum rated outlet pressure : 620 bar (9000 psig)maximum pressure ratio : 25:11) BASIC DISPLACEMENT Db:(from chart)2) PRESSURE LOAD Lp :10.2 liters/min (0 .36 actual CFM)Lp = (Po - Ps)/30 = 2700/30 = 90 psig3) DISPLACEMENT FACTOR Kd : 0.7(from chart for Lp = 90 psig)4) VOLUMETRIC EFFICIENCY Ve :	0 .7(from chart for category C)5) ACTUAL FLOW VOLUME : Db*Kd*Ve = 0.36*0.7*0.7 = 0 .17646) STANDARD FLOW VOLUME : Db*Kd*Ve* (Ps+ 14.7)/14.7 == 0.1764*314.7/14.7 = 3.7764 SCFMrequired are 35 .3 SCFM which means that the capacity of model AGD-30 is about oneorder of magnitude too low.1893) AGD-62 (double acting, single stage)maximum rated outlet pressure : 620 bar (9000 psig)maximum pressure ratio : 25:11) BASIC DISPLACEMENT Db : 6.99liters/min (0 .247 actual CFM)(from chart)2) PRESSURE LOAD Lp : Lp = (Po - Ps)/60 = 2700/60 = 45 psig3) DISPLACEMENT FACTOR Kd : 0.9(from chart for Lp = 45 psig and -85 SCFM air flow)4) VOLUMETRIC EFFICIENCY Ve :	0 .7(from chart for category C)5) ACTUAL FLOW VOLUME : Db*Kd*Ve = 0.247*0 .9*0.7 = 0.155616) STANDARD FLOW VOLUME :	Db*Kd*Ve*(Ps+ 14 .7)/14.7 == 0.15561*314.7/14.7 = 3.3313 SCFMrequired are 35 .3 SCFM which means that the capacity of model AGD-62 is also aboutone order of magnitude too low.4) AGT-15/30 (two stage)maximum rated outlet pressure : 586 bar (8500 psig)maximum pressure ratio : 50:11) BASIC DISPLACEMENT Db : 10.1 liters/min (0 .355 actual CFM)(from chart)2.1) PRESSURE LOAD Lp during output stroke:Lp = (Po - 2*Ps)/30 = 2400/30 = 80 psig2.2) PRESSURE LOAD Lp during interstage stroke:a) Lp = Ps/15 = 300/15 = 20 psigb) Lp = Po/30 = 3000/30 = 100 psig1902.3) TOTAL PRESSURE LOAD Lp :	80 psig3) DISPLACEMENT FACTOR Kd :	0 .75(from chart for Lp = 80 psig)4) VOLUMETRIC EFFICIENCY Ve :	0.93(from chart for category D)5) ACTUAL FLOW VOLUME: Db*Kd*Ve = 0.355*0 .75*0.93 = 0.24766) STANDARD FLOW VOLUME :	Db*Kd*Ve*(Ps+14 .7)/14.7 == 0.2476*314.7/14.7 = 5.3 SCFMthe capacity of a two stage model of reasonable size is also too low.It has been shown, that all the selected models are not suitable . Various other modelshave been investigated but their performance is even less suitable .191SPECIFICATION OF BOOSTER PACKAGE PROPOSED BY HASKEL L INC.Haskel Inc.100 East Graham PlaceBURBANK, CA. 91502Two boosters in series:first booster model number :	8AGD-5air drive pressure :	110 psigrequired air drive quantity:	174 SCFMbooster cycles per minute :	42second booster model number :	8AGT-14/30air drive pressure :	110 psigrequired air drive quantity:	182 SCFMbooster cycles per minute :	87system total required drive air :	356 SCFMgas supply pressure :	300 psigcalculated gas interstage pressure :	805 psiggas outlet pressue:	3000 psiggas outlet flow :	36.04 SCFMNOTE: Gas temperature is not considered in the listed performance data . Flow rateswill be effected by supply or discharge gas temperatures.COST :	AGD-5:	$ 9,294.00AGT-14/30: $ 9,659.001 Source: Fax, May 29th, 1992192APPENDIX A10SAMPLE CALCULATION : JNJECTOR DIMENSIONSThe purpose of this calculation is to determine the dimensions of an injector chamberwhich holds CNG for the combustion . A plunger which is activated by the camshaftcompresses the gas and forwards it into the combustion chamber. In particular, thediameter of the cylindrical chamber is interesting since it is limitted by the outside shapeof the injector.GIVEN:	CNG inlet pressure:	20 bar (300 psi)CNG inlet temperature :	285 Kgas constant :	500 J/kg Kfuel requirement :	15 kg/hr @ 600 rpmvalve lift :	0 .8255 cm (0 .325")1) Mass Requirement per Injector:m = 15 kg/hr @ 600 rpm(15 kg/hr)/60 = 0.25 kg/min @600 rpm(0.25 kg/min)/600 = 4 .166* 10-4 kg/revolution4.166* 10-4 kg/revolution)/6 = 6 .944* 10-5 kg/revolution and cylinderm = 6.944*10-5 kg/revolution and cylinder2) Volume Requirement of the Injector Body:V= mxRxTPV = 6.944 . 10 -5 . 500 . 285 = 4.9476 . 10 -6m32 . 10 6V=4.95cm3this result is obtained assuming 100% volumetric efficiency3) Chamber Diameter for a Given Lift:sv = d' 4 • L where L is the lift (i .e . 0 .8255 cm)193d =J4V =2.77 cmL•7rThis dimension is too big since the outside diameter of the injector body is 2 .54 cm. Tobe noted is that the effects of volumetric efficiency are not taken into account and thatthe injector needs an appropriate wall thickness which limits the chamber dimensionseven more . Hence, this intensifier type can be considered for lower pressure ratios(higher inlet pressures), only.194APPENDIX Al 1SAMPLE CALCULATION: CONVERSION OF A 50 ccm COMBUSTIONENGINEPurpose of this sample calculation is to find the number of strokes that are required torun a 50 ccm displacement compressor at a 10 :1 pressure ratio and full mass flow . Theresult is satisfactory if the required number of strokes is lower or equal to the actualnumber of strokes.GIVEN:	inlet pressrue:outlet pressure:gasconstant R:inlet temperature:mass flow :20.6 bar (300 psi)206 bar (3000 psi)500 J/kg K288 K45 kg/hr @ 2100 rpm15 kg/hr @ 600 rpmCompressor Dimensions:bore : 3.81 cm (1 .5")stroke: 4.445 cm (1 .75")displacement : 50.677 cm3clearance volume : 10%1) Volumetric Efficiency:=1–cro• [1112	-1206 1—3A 0 = 1–0.1•	-1 =0.512=51.2%20.62) Massflow per Stroke:m= P, V A = 2.06 .106 .50.667 .10.0.512R • T °	500 . 288tit = 3.712 * 10-4 kg/strokeP11953) Number of strokes required to meet the massflow at 2100 rpm:required massflow : 45 kg/hr = 0.75 kg/minrequired massflow = displacement/stroke * required # of strokes/minrequired # of strokes =	 0.75 4 = 2020.5 strokes/min3.712 . 10-4) Number of strokes required to meet the massflow at 600 rpm:required massflow : 15 kg/hr = 0.25 kg/minrequired massflow = displacement/stroke * required # of strokes/minrequired # of strokes =	 0.25	 = 673.5 strokes/min3.712 . 10-a5) Thrust force:F=p*AF = 3000 psi * 1 .52 * n / 4 = 5301 .5 lbsIt can be seen that the mass low requirement is satisfied at 2100 rpm, however, at 600rpm there is a shortage of approximately 10% . Thrust forces of at least 5300 lbs areexpected .196APPENDIX Al2SAMPLE CALCULATION: CAPACITY OF A HYDRAULIC FLUID PUMPThe purpose of this calculation is to determine the capacity of a hydraulic pump which isrequired to meet the gas consumption of the bus engine at design pressure ratio as it isconverted into a gas compressor.GIVEN :	Maximum fuel consumption of Detroit Diesel engine, model 6V 92 TA:m = 45 kg/hr @2100 rpmrim = 15 kg/hr @ 600 rpmclearance volume :	c = 15%gas constant :	R = 500 J/kg Kpolytropic coefficient : n = 1 .3gas inlet temperature : T = 288 Kgas inlet pressure :	p = 20 barspeed ratio :	1 :1 (i .e . pump operates at engine speed)1) Fuel consumption per revolution in g/rev:m = 45 . 1000= 0.35 g/rev @ 2100 rpm60 . 2100m = 15 . 1000 = 0.416 g/rev @ 600 rpm60 . 6002) Volumetric efficiency at design pressure ratio:a 1-E P " -1Pl=1-0.15 (10)13-1]=26.83%1973) Required pump displacement (ml/rev):V = m R•T•XPv = 0.416 .10-3 .500 .288 .0.2683 = 8.036 10.6 m3/rev2 . 106V = 8.036 ml/revHence, a displacement of -8 ml/rev is required to provide sufficient gas for thediesel engine.198APPENDIX A13PISTON ROD CALCULATION;The following calculation has been carried out in imperial units . It is based on theproperties of Type 316 Stainless Steel [4]:Yield Strength	30,000 psiTensile Strength	75,000 psiBoth tensile and thrust forces have been obtained, assuming a double acting piston withmaximum load (i .e . where the pressure on one side is 3000 psi and on the other side 300psi).Figure A13 : Forces and stresses for various rod sizesThe inerta forces have been neglected at this point since the weight of the piston is notknown. Figure A13 illustrates the load and stress behaviour for various rod sizesbetween 1/4" and 1" . It must be reminded that large rodsizes require a correction of thebasic displacement . In order to meet the given yield stress a rod size greater or equalROD SELECTION CHARTBORE 1 .25", P1= 300 psi, P2 = 3000 psiTHRUST FORCETENSILE FORCE	 YIELD STRESS100	00 .25	0 .3	0.35	0.4	0 .45	0 .5	0.55	0 .6	0.65	0 .7 0.75	0.8	0.85	0.9	0 .95	1ROD DIAMETER. (inch)T THRUST FORCE — TINSILE FORCE	THRUST STRESS — TENSILE STRESSIi 2so70605040302041990.35" must be chosen . A fatigue calculation has not been carried out since the number ofcycles of the prototype will be too low to require a proof of strength under the givenconditions.For the prototype intensifier the rod size was chosen 0 .5 inch (1 .27 cm) . The safetyfactor is 1.92 .200APPENDIX A14FLOW CALCULATION OF CHECK VALVEThe information for the the following gas flow calculation has been taken from the datasheet of the "SWAGELOK" catalogue [10] (distributor of NUPRO check valves) . Thedimensions in the equations are given in imperial units.1) Given :• Discharge Pressure P1 = 3000 psig• Flow Coefficient (CH4 model) Cv = 0 .67• Specific Gravity of Methane SG = 0.554(referred to air @70° F)• Absolute Gas Temperature	T = 520° F(in °F + 460)The critical operation is at 10 :1 pressure ratio and full fuel flow . At this point theopening time of the valve is the shortest and the gas flow is a maximum . The gas flowthrough the inlet valve is not as critical since the opening time is longer due to immediategas expansion to inlet pressure.2) Required Gas Flow:VDB = 35 SCFMThe volume flow in the forward stroke is 54.35% of the total volume flow according tothe intensifier dimensions.VDBF = 0 .5435 * 35 = 19 SCFMThe opening time has been assumed (1/12) of the stroke to allow for pressure build up inthe cylinder and for the lift of the poppet. Therefore, the adjusted volume flow isVDBF' = Q = 19 * 12 = 228 SCFMz2013) Pressure Drop:The equation for the gas flow is given as(p 2 -p 2 )Q=16.05 .0	SGTThe purpose of the calculation is to evaluate the pressure drop across the valve at thegiven conditions . The downstreams pressure is2P2 = P12 — 	 Q	 • SG • T = 2978.3 psig16.05 • Cvwhich refers to a pressure drop ofAP = P1 — P2 = 3000 — 2978.3 = 21.7 psigin metric units the pressure drop isAP=21.7 =1.47 bar14.7which is less than 1% of the total pressure and can be considered minor.Hence, the CH4 series valves are used as compressor valves .202APPENDIX B1INTENSIFIER PARTSSINGLE STAGE INTENSIFIER (Version 1,2)Following is an assembly drawing of the single stage intensifier, a parts list and machine shopdrawings of the intensifier components . All dimensions are in imperical units because all shopequipment measures imperial units . The machining tolerances are±0.002" unless otherwise specified.The scale of the drawings does not represent the original dimensions as the drawings have beenreduced to fit in the format of the document .203Figure B1 .1: Assembly Drawing of Single Stage Intensifier (Version 1 .2)	ID I A W N S Y : CIUMOp11 A.CtHngerMATIIIAI : StOklba SteelASSEMBLY DRAWINGINTENSIFIER (Version 1 .2)DRAWING IYMIUAD-1 (1)ICU/ : APPIOVIDDAIS : NOVenlber 1997DEPARTMENT OF MECHANICAL ENGINEERINGTHE UNIVERSITY OF BRITISH COLUMSIA204	. ThosierVerso.MT1 Discharge Valve Body	CH4 Series 2 Stainless Steel NUPRO2 Hex Head Nut	3/8" UNF 4 Steel3 Valve Spring (10 psi)	CH4 Series 2 Steel NUPRO4 Lock Washer	3/8" 4 Steel5 Poppet Stop	CH4 Series 4 NUPRO6 Valve Poppet	CH4 Series 4 Viton 0-Ring NUPRO7 Top Cap	DWG # : TC-1 1 Stainless Steel Block 2 .5x2.5x1 .58 0-Ring	AS 568 - 131 2 Viton9 Cylinder	DWG # : IC-1 1 Stainless SteelHollow BarL = 5 .500"I.D .: 0.790"O.D . : 1 .790"10 Tie Rod	DWG # : TR-1 4 Stainless Steel Bar 01/2"x 8"11 Piston (upper part) DWG # : WP-1 (3) 1 Stainless Steel Bar 01.250"x 1"12 Piston (middle part) DWG # : IPP-2 (3) 1 Stainless Steel Bar 01.250"x 1"13 Piston (lower part) DWG # : IPP-3 (3) 1 Stainless Steel Bar 01.250"x 1"14 PolyPak Piston Seal 25000750-375 2 PTFE Parker Seals15 0-Ring	AS 568 - 018 2 Viton16 Piston Rod	DWG # : PR-1 1 Stainless Steel Bar 01/2"x 11"17 Inlet Valve Body DWG # : VB-1 1 Stainless Steel Bar 01 1 /4"x 2.5"18 0-Ring	V19-OR-0908 2 Viton Swagelok19 Female Connector SS-400-7-4 4 Stainless Steel Swagelok20 Valve Spring (1 psi)	CH4 Series 2 Steel NUPRO21 Bottom Cap	DWG # : BC-1 1 Stainless Steel Block 2 .5x2.5x2.522 Rod Seal V-Packing ID : 1/2", OD : 1" 1 PTFE Power-Seal Corp.23 0-Ring	AS 568 - 124 2 Viton24 Allen Screw 1/4" x 1" 4 Steel25 Rod Seal Flange DWG # : SF-1 1 Aluminum Bar 04" x 1 .5"26 Lubrication Device 1 Steel 1/8" Pipe, Ellbow27 Velt Ring ID : 1/2", OD : 1" 1 Velt Sheet28 Permaglide® Bushing PAPZ 0814P10 Teflon INA Bearing29 Cylinder Flange DWG # : CF-1 1 Stainless Steel Bar 06"x3.5"30 Allen Screw 3/8" x 2.5" 6 Steel31 Lock Washer 3/8" 6 Steel1 A11 dimensions in inches20532 Center Ring	DWG # : CR-1 1 Stainless Steel Bar 04.5" x 0.75"33 Lock Nut 7/16" UNF thread 1 Steel Bar 0 1" x 1"34 Hex Head Bolt 5/16" x 1 .5 UNF thread 4 Steel35 Lock Washer 5/16" 4 Steel36 Crosshead Flange DWG # : RM-1 1 Steel Bar 0 3.5" x 1"37 Rod Extension DWG # : RM-1 1 Steel Bar 0 2" x 1"38 Flat Head Screw #10 x 1/4" 8 Brass39 Crank Case KOHLER Model K361 1 Cast Iron KOHLER40 Cross Head 1 Aluminum Bar 0 4" x 6"Table Bi: Parts List for Single Sta ge Intensifier (Version 1 .2)206view A -A000i0 I . vs-,.o.00~thread 8-32.DEPARTMENT OF MECHANICAL ENGINEERINGTHE UNIVERSITY OF BRITISH COLUMBIASCALE:DATE : June-24-19922:1 APPROVED SY : DRAWN B Y:ChristophAichIngerMATERIAL : 316 Stainless SteelPISTON (UPPER PART)INTENSIFIER (Version 1 .2) DRAWING NUMBERIPP-1 (3)207view A - A~o .00z0 0.854 -o.00z.r	 .0D.}50'a I~~33/L4 ~,0a —0 0o	 ///	 Cy/}0It,.0000,~	 d0.50-0.0000.854!8	~ 1 .2't~t'o°~zA ADEPARTMENT. OF MECHANICAL ENGINEERINGTHE UNIVERSITY OF $itITISH COLUM IASCALE:DATE : June-24-1~2 :1 I APPROVED BY : DRAWN B Y:Christoph AlchIngerMATERIAL : 316 Stainless SteelPISTON (MIDDLE PART)INTENSIFIER (Version 12) DRAWING NUMBERIPP-2 (3)208DATE : June-24-1992SCALE :-DEPARTMENT OF MECHANICAL ENGINEERINGTHE UNIVERSITY OOF 1tRI~$SH COLUMBIA2 :1 APPROVED BY : DRAWN BY :Chrtstoph AlchlnperMATERIAL : 316 Stainless Steelview A - A	o I.2Jtlt r o .o01O. ?50oocb 0.500«0001	0.150-o .00lthreadILL -Z0PISTON (LOWER PART)INTENSIFIER (Version 1 .2)DRAWING NUMBERIPP-3 (3)209tbtn ,ad i/i UNF8t$ rei61 10 s/aDEPARTINIENT OF MECHANICAL ENGINEERINGTHE UNIVERSITY OF BRITISH COLUMBIASCALE:DATE : May-08-19921 :1 APPROVED SY :MATERIAL : Stainless Steel type 316	DRAWN BY :Chdstoph AlChlnperPISTON RODINTENSIFIER (Version 1 .2) DRAWING NUMBERPR-1 (1)210view A - A2451 .1-708LA . APART IENT OF MECHANICAL ENGINEERING[#1E UNIVERSITY OF BRITISH COLUMBIA,xSCALE :	2 :1DATE: June-24-1992APPROVED SY :	DRAWN BY : ChristophAlchingerMATERIAL: 316 StOlnleSS SteelCYLINDERINTENSIFIER (Version 1 .2) DRAWING NUMBERIC-1 (1)211thrtarl 3/e UNF	0thread'12- UNFA-ky4 pieces	y	 Qf O50 0DEPARTMENT tF MECHANICAL. ENGINEERINGTAE UN1vt:1tS ' I Y Of BRITISH COLUMBIASCALE :	 1:1	 I APPROVED BY :	 DRAWN BY :Chrlstoph AlchingerDATE : May-08-1992	MATERIAL : Stainless Steel type 316TIE RODINTENSIFIER (Version 1 .2) DRAWING NUMBERTR-1 (1)212os, ogrNaWdctCaoC00a0>0s02LLU'0}0mwJw-C0eccoo-4-,	ILn'o/'00	ji,/W<0Wr02130 I .5O$00 9hb	 09/160 3.SS0 4 e.2EPARTRIENT OF IiECHANICAL ENGINEER NGTHE UNIVERSIT•r' OF BRITISH COLUMBIASCALE :	2 :1 DRAWN BY :Chrhtoph AlchhgerAPPROVED BY .DATE : MOy-081992	MATERIAL : AluminumROD SEAL FLANGE214wwaoULU3. <o iozUz3wwgu..wzJRstt78/{9191h,Qee0Ao4iO8	r-/a°SS zaJ2150WZiwUvLuLt.NUZO0W1NP°"`dt1dn9456i-408s2165a0OOS'000FoQ3	rlXS-1217w= v1!=~Lt1D70Oa00b"0	00i;OToo o -ZOa'Ur'r7J000 I0+,9/c0.GS,,OS-00SZI 0c0	050	OSI ' I00 0-	Zi000-aLLLUz3Ocs{C0ONcNmWJ.S'Iv08000070IIEI\/-jljjl ,Frti	.a V.U-8	Ui2180donQc'oOOQO_ioOro•QQOToo o -StC O000roQ•e	\ \219APPENDIX B2INTENSIFIER PARTSTWO-STAGE INTENSIFIER (Version 2.1)Following is an assembly drawing of the two stage intensifier, a parts list and machine shopdrawings of the intensifier components . Dimensions and tolerances are as indicated in AppendixB1.Some intensifier parts have been adapted from the single stage intensifier, such as the piston andthe valves for the second stage. All single stage parts specified in the partslist are attached inAppendix B1 .ASSEMBLY DRAWINGI	DRAWING NUMBERTWO-STAGE INTENSIFIER (Version 2 .1)	AD2-1 (1)10 11 1237 36 35DEPARTMENT OF MECHANICAL ENGINEERINGTHE UNIVERSITY OF IRITISH COLUMBIASCALE :	 1 :1	IAEEEOVED BY :	(DRAWN BY :ChrMOphAlChklperDATE : December 1992	MATERIAL: STdOIOSS Steel220Figure B2 .1: Assembly Drawing of Two-Stage Intensifier (Version 2.1)221Two-Stage $ens ler Version 2 .1DG#. . . . . . . . . .. . . . . . . . .. . . . . . . . ..~». MA °COMl	NT1 Lock Washer	3/8" 4 Steel2 Hex Head Nut	3/8" UNF Steel3 Valve Spring (5 psi)	CH8 Series 1 Steel NUPRO4 Valve Components DWG # : VC-1 (1) 1 Steel, Alu, Viton NUPRO5 Lock Washer	1/4" 12 Steel6 Hex Head Screw 1/4" UNC 6 Steel7 Valve Cover	DWG # : VC-1 (1) 1 Stainless Steel Bar 0 2" x 0 .5"8 0-Ring	AS 568 - 118 1 Viton9 Top Cap	DWG #: TC-1 (1) 1 Stainless Steel Bar 0 4"x 3"10 Valve Poppet	CH8 Series 1 Viton 0-Ring NUPRO11 Poppet Stop	CH8 Series 1 Stainless Steel NUPRO12 Valve Spring (1 psi)	CH8 Series 1 Stainless Steel NUPRO13 Valve Body	CH8 Series 1 Stainless Steel NUPRO14 Swagelok Nut 1/2" pipe 2 Stainless Steel NUPRO15 O-Ring	AS 568 - 131 2 Viton16 Piston (upper part) DWG # : WP-1 (3) 1 Stainless Steel Bar 0 1.250"x 1"17 Piston (middle part) DWG #: WP-2 (3) 1 Stainless Steel Bar 0 1.250"x 1"18 Piston (lower part) DWG # : WP-3 (3) 1 Stainless Steel Bar 0 1.250"x 1"19 _-_pacer 1 PTFE Bar 0 1" x 1/4"20 Inlet Valve Body DWG # : VB-1 2 Stainless Steel Bar 0 1 .25" x 2.5"21 0-Ring	V19-OR-0908 2 Viton Swagelok22 Female Connector SS-400-7-4 4 Stainless Steel Swagelok23 Swagelok Nut 1/4" pipe 4 Stainless Steel NUPRO24 Valve Poppet	CH4 Series 4 Viton 0-Ring NUPRO25 Poppet Stop	CH4 Series 4 Stainless Steel NUPRO26 Lubrication Device 1 Steel 1/8" Pipe, Elibow27 Bottom Cap	DWG # : BC-1 1 Stainless Steel Bar 0 4" x 2 .5"28 Rod Seal V-Packing I.D . : 1", CS: 1/4" 1 PTFE Power-Seal Corp.29 Veit Ring CS: 1/4" 1 Woolvelt Sheet30 Permaglide® Bushing PAPZ 1616 P10 1 Teflon INA BearingAll dimensions in inches22231 Lock Washer 3/8" 6 Steel32 Allen Head Screw 3/8" x 2.5 UNC 6 Steel33 Lock Nut 3/4" UNF thread 1 Steel Bar 0 1" x 1"34 Allen Head Screw 1/4" x 1 .5 UNC 6 Steel35 Rod Extension	DWG # : RF-1 1 Steel Bar 0 3 .5" x 1 .1"36 Crosshead 1 Aluminum Bar 0 4" x 6"37 Crank Case	KOHLER Model K361 1 Cast Iron KOHLER38 Crosshead Flange DWG #: RF-1 1 Steel Bar 0 3 .5" x 1 .1"39 Center Ring	DWG # : CR-1 1 Stainless Steel Bar 04.5" x 0.75"40 Lubrication Device 1 Steel 1/8" pipe, ellbow41 0-Ring	AS 568 - 137 1 Viton42 Valve Body	CH4 Series 2 Stainless Steel NUPRO43 Valve Spring	CH4 Series 2 Stainless Steel NUPRO44 Piston Rod	DWG #: PR-1 1 Stainless Steel Bar 0 1 .25"x10 .5"45 Cylinder	DWG # : IC-1 1 Stainless Steel Bar 0 3" x 5"46 Tie Rod	DWG #: TR-1 4 Stainless Steel Bar 0 0 .5" x 8"47 PolyPak Piston Seal 25000750-375 2 PTFE PARKER48 0-Ring	AS 568 - 018 2 Viton49 Male Connector SS-810-1-8 1 Stainless Steel SwagelokTable B2: Parts List for Two-Stage Intensifier (Version 2 .1)223aW-EN6 '01W0Q00t Nmm00$}OSzl-0oo'T00z0.,OS90	oS7022430.00	o	o	oo0-4,	,o	8,‘V_ a0\'724	\N/A:_!	\ \•Fr-Z0Sn& '0ifcoooz''I?-GI225226St'►aro-app ~a+o-o+47J2J xLLU-zLS1J00°GIaCOemincU.HQZ!W00-4022700C-OMC04Q000.1	\\\Lz	000Io'.	0048-228LJ_xU...YGZwz0c a0a9-o0C00w<	0_EO	>wCl)	0ofo.e.QDCnlZrNW0CGawDaVzI-c0C.)U,00Z'0c 9.0a229APPENDIX B3DETERMINATION OF CLEARANCE VOLUMESINGLE STAGE INTENSIFIER Version 1 .2All clearance spaces are described and listed Table B3.1) Intensifier Dimensions:Bore :	3 .175 cm	1.25"Stroke:	8 .255 cm	3.25"Rod :	1 .270 cm	0.50"Displacement UP-stroke:	65.36 cm3Displacement DOWN-stroke :	54.90 cm3Displacement TOTAL:	120.26 cm32) Measured Clearance Spaces:a) Stroke Clearance (UP-stroke):This is the clearance volume between the upper piston surface at TDC and the top cap . Itis adjustable between 0 and 0.89 mm. The clearance has been adjusted to Cstri = 0.635 mm(measured with a dial gauge) which is a total stroke clearance volume of'CV, =B2 4=0.503cm3b) Stroke Clearance (DOWN-stroke):This is the clearance volume between the lower piston surface at BDC and the bottom cap. Theminimum distance between piston and bottom cap (Cstr2) is 0 .89 mm minus the UP-strokeclearance (Cstr2 = 0.255 mm) . The total stroke clearance volume is230CV 2,2 = (B2 - R2 ) • n • C,,r2 = 0.169 cm3c) Valve Clearance (Inlet Valve):The clearance space caused by the inlet valve was measured with a liquid as assembled on theintensifier.CVii, = 0.9 cm3d) Valve Clearance (Discharge Valve):liquid measured as assembledCVO1, = 0.0695 cm3e) Rod Clearance:This is the clearance space due to the radial slot between rod and bottom cap . Thedimensions of the slot are :	bore : 1 .424 cmlength: 2.032 cmThe total clearance volume is : CR = (B2 — R2) . n •74 • L = 0.655 cm3f) Crevice Clearance :The crevice clearance is defined as the radial slot due to the tolerance between piston andcylinder surface . The length of the slot is the axial distance from the horizontal piston surface tothe point of contact between seal lip and cylinder . The crevice clearance is equal for UP andDOWN stroke. Following are the dimensions of the slot : cylinder bore:	3.175 cmpiston diameter : 3 .160 cmrelevant length : 0.635 cmThe total crevice clearance per piston side is : Cc = (B2 - P2 ) •4L = 0.048 cm32313) Neglected Clearance Spaces:• the hex shaped socket of the set screws• the radial crevice between the cylinder and the top cap• the radial crevice between the cylinder and the bottom cap• the space due to pressure loading of the piston seals• the space due to pressure loading of the rod V-packing• clearance spaces inside the piston4) Total Clearance Volume:The clearance volume must be regarded seperately for each stroke . The total clearance spacesare summed up in the following table :. . . . . . . .. . . . . . . . . . . .. . . . . . . ..ARANCE SPACR. . .<	VOLUMET O	DOWNSvalve clearance (discharge valveTable B3: List of Clearance Spaces of Single Stage Intensifer (Version 1 .2)The relative clearance volume for the UP-stroke is : 1.5205 .100 = 2.33 %65.36The relative clearance volume for the DOWN-stroke is : 2.1755 .100 = 3.96 %54.90The AVERAGED relative clearance volume is : 2.33 65.36+ 3.96 .	 54.90 = 3.07 %120.26	120.26232The expected volumetric efficiency of the intensifier is illustrated in Figure B3.1 . The totalvolumetric efficiency is defined as the sum of volumetric efficiencies of UP and DOWN stroke,each corrected according to their relative volume.65.36	54.90VETOTAL = VEur 120.26 +VEDORW120.26VOLUMETRIC EFFICIENCY INTENSIFIER Version 1 .2CLEARANCE VOLUME : UP = 2 .33%, DOWN =3 .%%4	5	6	7	8	9	10	11PRESSURE RATIO (P2/P 1)Figure B3 .1: Expected volumetric efficiency of single stage intensifier (Version 1 .2)233APPENDIX B4DETERMINATION OF CLEARANCE VOLUMETWO-STAGE INTENSIFIER Version 2 .1In this section, all clearance spaces are described and listed in Table B4.1) Intensifier Dimensions:Bore : 3 .175 cm 1.25"Stroke: 8 .255 cm 3.25"Rod : 2.540 cm 1.00"Displacement Stage 1 : 65.36 cm3Displacement Stage 2 : 23.51 cm32) Measured Clearance Spaces in Stage 1:a) Stroke Clearance:This is the clearance volume between the upper piston surface at TDC and the top cap,adjustable between 0 and 1 .52 mm. The clearance has been adjusted to Cstri = 0.89 mm(measured with a dial gauge) which is a total stroke clearance volume ofC17..I = B2 .n- 4 .Ctt,l = 0.7038 cm3b) Valve Clearance (Inlet Valve):The clearance space caused by the inlet valve has been measured with a liquid asassembled on the intensifier.CV,,=0.90cm3234c) Valve Clearance (Discharge Valve):partly liquid measured and calculatedCVoy = 0.415 cm3d) Crevice Clearance due to Piston:The crevice clearance is defined as the radial slot due to the tolerance between piston andcylinder surface . The length of the slot is the axial distance from the horizontal piston surface tothe point of contact between seal lip and cylinder . The crevice clearance is equal for both stages.Following are the dimensions of the slot:cylinder bore : 3 .175 cmpiston diameter : 3 .160 cmrelevant length : 0.635 cmThe total crevice clearance per piston side is : Cc = (B2 - P2 ) • 4 . L = 0.048 cm3e) Pressure Transducer Clearance:This is the clearance space caused by the piezo electric pressure transducer.CV t = 0.075 cm33) Measured Clearance Spaces in Stage 2:a) Stroke Clearance:This is the clearance volume between the lower piston surface at BDC and the bottomcap. Its distance (Cstr2) is 1 .52 mm minus the stroke clearance of the first stage (Cstr2 = 0 .63mm). The total stroke clearance volume isCV,, ,2 = (B2 - R2 ) • 7r .4 • C„'2 = 0.181 cm3235b) Valve Clearance (Inlet Valve):The clearance space caused by the inlet valve has been measured with a liquid asassembled on the intensifier . Note that one or two inlet valves can be used for tests.CVii, = 1.025cm3 (each)A plug was made to seal one inlet valve . The clearance space using the valve plug wasmeasured with liquid:CVyp = 0.22 •7r - 0.1 = 0.003142 in3 = 0.0515 cm3c) Valve Clearance (Discharge Valve):liquid measured as assembledCVoi, = 0.125 cm3 (each)d) Rod Clearance:This is the clearance space due to the radial slot between rod and bottom cap. Itsdimensions are :	bore: 2.578 cmlength: 1 .016 cmThe rod clearance volume is : CR = (B2 — R 2 ) -7r • 74 L = 0.156cm3e) Crevice Clearance due to Piston:Total crevice clearance : Cc = (B2 — P2) . 4 • L = 0.048cm34) Neglected Clearance Spaces:Following clearance spaces have been neglected:• the hex shaped socket of the set screws• the radial crevice between the cylinder and the top cap• the radial crevice between the cylinder and the bottom cap236• the clearance space due to pressure loading of the piston seals• the clearance space due to pressure loading of the rod V-packing• clearance spaces inside the piston• the clearance space due to the pressure transducer in the second stage5) Total Clearance Volume:stroke clearance 0.7038 0.1810.390 2 @ 1 .0250.415 2 @ 0.125N/A 0 .1560.075 0N/A 0 .05150.048 0.0481 .632 2.6851.632 1.711valve clearance (inlet valve)valve clearance (discharge valve)rod clearancepressure transducer clearanceinlet valve plug' .crevice clearanceTOTAL CLEARANCE (without plug)TOTAL CLEARANCE (with plug)CLEARANCE VOLUME `(+Table B4: List of Clearance Spaces of Two-Stage Intensifier (Version 2 .1)The relative clearance volume for Stage 1 is : 1.632 .100 = 2.50 %65.36The relative clearance volume for Stage 2 is :	 ?'685 . 100 = 11.41 % without plug23.53The relative clearance volume for Stage 2 is : 1 .711 , 100 = 7.27 % with plug23.53'.Note that if the plug is used only one inlet valve accounts for clearance volume.237APPENDIX ClEauipment Specification: TEST RIGELECTRIC MOTOR:Brand Name: HampstonType:	variable speed, DCMounting :	trunnion mountedTorque Arm: (shaft center - loadcell) 17.49"Speed:	0 - 2400 rpmPower:	upgraded to max. 25 hp @ 2400 rpmControl :	tachometer feedback speed regulationShaft 0 :	1 .600"MOTOR CONTROLLER:Brand Name: RandtronicsType:	regenerative DC motor controllerModel :	TB 750 SeriesRated Power: 15 hp (upgraded to 25 hp)Line Voltage: 240 V (3 phase)Output :	0 - 240 V DC (max. 55 Amps)Field :	100 VMode:	tachometer feedback speed regulationCRANK CASE:Brand Name : KohlerType:	single-cylinder, 4 cycle, air cooled combustion engineModel :	K361Bore :	95 .2 mm (3 .75")238Stroke:	82.6 mm (3 .25")Shaft Diameter : 17/16"Power :	13 .4 kW (18 hp) @3600 rpmTRANSMISSION:Type :	Synchronous Belt Drive (Gearbelt)Brand Name : BrowningModel : HPT 8MDriver Pulley (E-motor):Pitch Diameter: 3 .609"Number of Teeth:	36Bushing Type : SHDriven Pulley (Intensifier):Pitch Diameter:	19.248Number of Teeth:	192Bushing Type : ESpeed Ratio : 5.33 : 1Center Distance : 19.9"Belt Length : 2000 mmBelt Width : 50 mmBelt Length Factor: 1 .2Power: 28.284 hp @ 1750 rpm driver speedFLYWHEEL:Outside Diameter: 16"Width : 4"Weight : 45 kg239APPENDIX C2Equipment Specification: INSTRUMENTATIONIn this part, the measurement devices for steady state and high speed data are specified . Alltechnical data has been taken from data sheets or installation manuals that came with the devices.The accuracy of the devices as well as non-linearity and hysterisis are not specificallymentionned, rather, all experimental data presented is examined in terms of repeatability.Measurement devices which have been calibrated are presented with a calibration curve as wellas a brief explanation of the calibration procedure . Following is a list of measurement devicesdiscussed:STEADY STATE DATA:• Loadcell• Pressure Transducers• Thermocouples• Massflow Meter• TachometerHIGH SPEED DATA:• Shaft Encoder• Piezoelectric Pressure TransducerSTEADY STATE DATA INSTRUMENTS:1) LOADCELLMeasurement :	TorqueUnits:	NmModel :	Interface SM-250 (Serial # C26470)Type:	Strain Gage Force TransducerCapacity :	0 - 250 lbsOutput :	3 .218 mV/V240Excitation :	10 VDCSignal Conditioning : 5B38 analog deviceFigure C2.1 : Loadcell Calibration SetupThe loadcell was calibrated being installed on the test rig in its original test position . The testload was applied to the electric motor via load arm which was attached to the motor shaft (Fig.C2.1) .LOADCELL CALIBRATIONLOAD ARM LENGTH = 16", June-19-19921 .500 20 100 14012040	60	80TORQUE [Nm]. TEST	LINEAR FITFigure C2.2: Calibration curve of strain gage type loadcell241Motor case and motor shaft have been locked together to transmit the load directly to theloadcell via trunnion bearings.Knowing the distance of the load arm and the torque arm to shaft center as well as the appliedload, the applied and measured torque can be calculated . The load applied has been measured inpounds . The length of the load arm is 16".The voltage output signal has been read after the signal conditioning . The voltage output hasbeen plotted against the torque (Nm) which has been calculated from the applied load and theload arm length.2) PRESSURE TRANSDUCER (2 units)Measurement:Units:Model:Type:Capacity:Input:Output:Signal Conditioning :CNG inlet pressure, CNG discharge pressurebarData Sensors, Inc., PB 1000-Gstrain gage, diaphragm0 - 206 bar (0 - 3000 psi)10 VDC30 mV5B38 analog deviceThe pressure transducers have been calibrated on a dead weight gage tester where a hydraulicpressure (in psig) is applied to the transducer . The voltage output signal has been recorded aftersignal conditioning.The transducer, measuring the inlet pressure, has to provide accurate reading at pressuresbetween 10 and 100 bar, whereas the pressure transducer for the outlet pressure has to cover therange between 100 and 200 bar .242Figure C2.3: Calibration curve of inlet pressure transducerFigure C2.4: Calibration curve of outlet pressure transducerBoth curves have been devided into linear sections to produce the most accurate reading for thespecific pressure range . Following is a table showing the linear intervals for both pressuretransducers representing the linear fit in Fig . C23 and Fig . C2.4.CALIBRATION OF PRESSURE TRANSDUCERINLET PRESSURE (CHANNEL #4) OCT-925050	100	150	200PRESSURE (bar)TEST DATA	. TEST DATA	FT1 TED CURVE4CALIBRATION OF PRESSURE TRANSDUCEROUTLET PRESSURE (CHANNEL #5) OCT-9250	100	150	200	250	300PRESSURE (bar). TEST DATA	FITTED CURVE0243PRESSURE (bar) VOLTAGE (V)Inlet Pressure Tr. Outlet Pressure Tr.0.00 -0.139 -0.46513 .60 0.24027.21 0.63534 .01 0.83054 .42 1 .40068.03 1 .770 1 .265102 .04 2.635 2.180136 .05 3.085170.07 3.980187 .07 4 .680204 .08 4.850238.10 5.725Table C2: Pressure Transducer Calibration Intervals4) PRESSURE TRANSDUCER (1 unit)Measurement:Units:Model:Type:Capacity:Excitation:Output:Signal Conditioning :CNG interstage pressurebarData Instruments, model SAhigh-gain strain gage0 - 206 bar (0 - 3000 psi)10 VDC1-6 VDCN/A244The calibration of this pressure transducer has been carried out exactly the same way asdescribed before . The equation used for the linear fit is shown on the graph.Figure C2.5: Calibration curve of interstage pressure transducerNOTE: All pressure transducers exposed to gas temperatures higher than 100°C have beenequipped with an extension pipe (from the main line) to protect the transducer from damage andmalfunction due to high gas temperatures.5) THERMOCOUPLES (5 units)Measurement: CNG inlet temperature, CNG interstage temperature(before cooling), CNG interstage temperature (after cooling), CNGoutlet (end) temperature, cylinder wall temperatureUnits :	°CModel :	Omega ThermocoupleCALIBRATION OF PRESSURE TRANSDUCERINIERSTAGE PRESSURE (CHANNEL # 8) NOV-92760PRESSURE [bar] = 40.9426 * VOLTAGE [V] - 40.942650	100	150	200PRESSURE (bar). TEST DATA	. TEST DATA	FITTED CURVE0 250245Type:	J-type (iron(+) vs. constantan(-))Capacity :	0 - 760 °CExcitation :	10 VDCOutput :	0 - 40 mVSignal Conditioning : 5B37 analog device(provides the cold junction reference point)The voltage output signal (mV) of a J-type thermocouple can be translated into temperature T(°C) using a 5th order polynomial. It covers a temperature range from 0 to 760 °C (± 0 .1 °C). Tis defined as follows:T = a® + x(al + x(a2 + x(a3 + x(a4 + asx))))The coefficients for a J-type thermocouple (Seebeck Coefficient : 51 vV/°C) are as follows:ap = -0.048868252al = 19873.14503a2 = -218614.5353a3 = 11569199.78a4 -264917531 .4as = 2018441314While the thermocouple generates a low voltage signal (0 - 40 mV), the signal conditioningdevice amplifies this signal to 0 - 5 V . Therefore, the thermocouple output (TC) and the signalconditioning output (SC) need to be correlated, whereSC = (TC * a) +band TC: -100 °C <=> -0.004632 V+760 °C <=> +0.042922 VSC:-100 °C <=> 0.00 V+760 °C <_> 5.00 Vhence: a = 105 .1436246b = 0 .487025The temperature curve (polynomial) can now be plotted versus the signal conditioning output.Figure C2.6: Calibration curve of inlet temperature thermocoupleCALIBRATION OF THERMOCOUPLEINLET TEMPERATURE, SIGNAL CONDITIONING: 51337- TEMPERATURE [°C] = 183.6766 * VOLTAGE [V] - 89 .3303-100 .4	0.5	0 .6	0.7	0 .8	0.9	1	1 .1SIGNAL CONDITIONING OU l PUT [V]-5th ORDER POLYNOMIAL LINEAR NTT ERROR1009080c 7060P 5o100CALIBRATION OF THERMOCOUPLEEND-, WALL-, INTERSTAGE TEMPERATURES, SIGNAL CONDITIONING : 51337300250-TEMPERATURE [°C] = 173 .7222' VOLTAGE [V] -81 .9104loo00.5	1	1 .5	2SIGNAL WNDITIONING OUTPUT [V]-5th ORDER POLYNOMIAL LINEAR FIT ERROR • 10-5o0 2.5Figure C2.7: Calibration curve of wall, outlet and interstage temperature thermocouples247However, only a small part of the curve is of interest for the experiments . Therefore, a linearregression of the curve has been done for both the inlet temperature thermocouple (needs tocover only a small range but very accuratly) and all other thermocouples . The inlet temperaturethermocouple has been calibrated between 2 and 92 °C and the other thermocouples between 2and 265 °C to achieve maximum accuracy.6) MASSFLOW METERMeasurement:	CNG massflowUnits:Sensor Model:Transmitter Model:Type:Capacity:Pressure Rating:Temperature Rating:Signal Conditioning :barMicro Motion DH012S 100 (Serial # : 140365)Micro Motion RFT9712Coriolis meter0 - 300 kg/hr393 bar-240 - 204 °C5B32 analog deviceThe massflow meter is calibrated on-line with a special calibration computer . A range ofmassflows between 30 and 300 kg/hr can be selected . The setting can be changed according tothe maximum expected massflow to achieve maximum accuracy.7) TACHOMETERMeasurement:Units:Model:Type:motor speedrpmSingergenerator248Output :	7 V / 1000 rpmSignal Conditioning : 5B31 analog deviceSince the voltage signal sent to the signal conditioning device must not be higher than 5V, avoltage devider has been installed between tachometer and the analog device,which consists of a115 fl and a 33 SZ resistor.Figure C2.8 : Calibration curve of tachometer used on the electric motorConsequently, the output voltage has been reduced by a factor of 0 .287 (33/115) and the newoutput is 2 .008696 Volts per 1000 rpm . The signal has been measured with the data acquisitionsystem and with a voltmeter . While running at different speeds, the rpm reading has beenverified using a hand-held optical tachometer.TACHOMETER CALIBRATIONJUNE-09-1992, VOLTAGE OUTPUT : 2.0087 V / 1000 RPM, S .C. : 513310500	1000	1500	2000MOTORSPEEL [RPM]. DATA ACQUISITION READING	. VOLTMETER READING- LINEAR FIT0 3000250055' 432249HIGH SPEED DATA INSTRUMENTS:1) SHAFT ENCODERMeasurement:	intensifier speedUnits:Model:Type:Output :	360 TTL level pulses per shaft revolution1 index pulse per shaft revolution (used as positionreference point and acquisition trigger)Signal Conditioning: high speed data acquisitionA calibration of the shaft encoder is not required but the index pulse must be exactly adjusted.The index pulse has been set to where the intensifier piston is at bottom dead center . The BDCposition has been measured using a feeler gauge . Then the softpot has been rotated to find theindex pulse and lined it up with the crank angle position at BDC . After the adjustment, thesensor has been locked in this position (i .e . no movement relative to the crank case)2) PIEZOELECTRIC PRESSURE TRANSDUCERcrank angle degreeUS Digital, Softpot SP-360 IBindex bearing encoderMeasurement:Units:Model:Type:Pressure Range:Natural Frequency:Sensitivity:Linearity :dynamic cylinder pressure (single stage intensifier andfirst stage of two stage intensifier)psi (later converted to bar)PCB 112 A (Serial #: 10671)quartz transducer0- 3000 psi300 kHz1 .09 pC/psi± 1% of full scale250Capacity :	23.6 pFSignal Conditioning : high speed data acquisitionFigure C2.9: Original calibration chart for transducer # 10671, provided by PCBPiezotronics, Inc . . Calibrated on Feb-28-1992MMI1111IMMIlMMMMIMMMMIMMMIMIMOMMMMMIlMMMMMM nMBBOM	BNNM..aM  M	.M..MBN.M.M..afNN.MBNNMMMMOB.NN.MOB.M.M.M. MEMNON. M..MLN MMMMMMMMMMMMM.WN..MM...MtMIIIMMMM0•YMM.NMM..M MNMM.. M...MM. NM.MMIMNMM.OMN...n~M BB-	U	 M M....4NM...M.MM.W..MUUMUU=n i "MMMM..MMM..M...M.W.M.YMM/MM.M.tl..	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Calibrated on Feb-28-1992Measurement :	dynamic cylinder pressure (second stage of two stageintensifier)Units :	psi (later converted to bar)Model :	PCB 112 A (Serial #: 10672)10015002000250030003500500000Calbration of Instrumentation used to certifythis sensor is traceable to NISTPCB PMzotronl s' eaibratlon procedure Is ineompilance with MIL-STD-4566242000	2500	3000252Type:	quartz transducerPressure Range:	0 - 3000 psiNatural Frequency : 300 kHzSensitivity :	1 .13 pC/psiLinearity :	± 1% of full scaleCapacity :	24.1 pFSignal Conditioning : high speed data acquisition253APPENDIX C3Equipment Specification: DATA ACOUISITION,The purpose of this section is to provide information about the components of the dataacquisition system . It is aimed to specify the most important parameters and settings, rather thana detailed description of the equipment . If further information abuot a specific model isrequired, the companies should be consulted ..The data acquisition consists of following parts:1) Personal ComputerBrand: ANOType: IBM compatibleChip: INTEL 286The PC is equipped with an A/D board and a general purpose interface bus (GPIB) tocommunicate with the ISAAC computer.2) AID Board:Model: ADVANTECH, PCL-818, PC-LabCard SeriesType: high speed, high performance I/O card for IBM A/D conversionAnalog Input:• Channels : 16 single-ended or 8 differential• Resolution : 12 bits• Converter: ADC-774• Conversion Time: 8 microseconds• Input Range bipolar: ± 10V, 5V, 2 .5V, 1V, 0 .5V•Input Range unipolar : 0-10V, 0-5V, 0-2V, 0-1V• Input Range Selection : programmable• Trigger Mode : by software254Digital Output:• Channels: 16A/D Pacer and Counter:• Device: INTEL 8254• Pacer: 32-bit with a 10 MHz or 1 MHz time base• Pacer max . Rate: 2 .5 MHz• Counter : 16-bit counter with 100 kHz time base3) Charge Amplifier:Model: Kistler Instrument Corporation 5004Type: Dual mode amplifierSetting :	Time constant: "Medium" (1 - 5000)Engineering Units/Volt : 500Sensitivity: depending on pressure transducer used (see PCBcalibration information)4) ISAAC 2000:Company: Cyborg CorporationOn-Board Features : • 68000 microprocessor which controls all systemfunctions• 2 megabytes RAM• ICL (ISAAC Command Language)255APPENDIX C4TEST PROCEDURE:Following is the typical sequence of a performance test:1) STEADY STATE DATA:Start UL• All measurement devices are turned on approximately 1 hour prior to tests• Inlet and receiver tanks are selected• The pressure regulator is adjusted to minimum position (i .e. zero flow)• The line pressure is adjusted to atmospheric pressure and the bypass valve(metering valve) is opened• The motor is turned on and adjusted to a selected speed• The intensifier operates approximately 10 minutes at atmospheric pressure towarm up• The inlet pressure is slowly increased up to the desired level (gas is circledthrough the bypass line at inlet pressure)• The instrumentation and data acquisition is initiated• Then the bypass valve is closed and gas is compressed• The gas pressure rises until it reaches the previously adjusted cracking pressureof the relief valve• The discharge pressure (i .e . cracking pressure) is readjusted to the exact settingTest• Once a steady state condition has been established, data is acquired• The inlet pressure is slowly increased to cover a range of pressure ratios• At certain conditions, high speed pressure data are acquired2562) HIGH SPEED DATA:• Allow 2 hours warm-up time for data acquisition system•	The pressure transducers are cleaned and all grease and oil is removed with analcohol solution• The setting of the charge amplifier is adjusted according to PCB specifications• The data acquisition hardware and software is set up•	The intensifier is started and a steady state condition is selected (see steadystate data)• Shortly before data is acquired, both ISAAC and charge amplifier are reset•	Inlet pressure, discharge pressure and filename are recorded for furtherreference• High speed cylinder pressure data acquisition is initiated via PC•	After having finished the high speed data acquisition (–20 sec), the inletpressure is recorded againNOTE: It is essential to know the exact inlet pressure as this is the only reference as to how toadjust the offset of the cylinder pressure data . If, during the time of cylinder pressureacquisition, the inlet pressure varies more than 5% (due to malfunction of the metering valvewhich is likely at very low inlet pressures because the orifice freezes as a result of gas expansionfrom tank pressure to adjusted inlet pressure), the data is not valid .257APPENDIX C5Calculations for Single Stage Intensifier Performance:* *************************************************** Subroutine to calculate performance values ***************************************************************Set constant values:rho.air = 1 .205'	kg/m"3 Density of air @ 200C, 1 .0133 barRGEN = 8314 .34 '	J/Kmol*K ' GENERAL GASCONSTANTSTANDARD VALUES FOR NATURAL GAS:Density :rho.STD = .713142	'kg/m"3Molecular weight:M.STD = 17 .1456999#	'kg/KmolGasconstant:R.STD = RGEN / M.STD	'J/kg*KB.C. HYDRO VALUES FOR NATURAL GAS:Density:rho.BCH = .690351	'kg/m"3Molecular weight:M.BCH = 16.597737#	'kg/KmolGasconstant:R.BCH = RGEN / M.BCH	'J/kg*KKelvin = 273.15	Temperature Conversion Factork.isentr = 1 .3	Isentropic Coefficient of Natural GasSet variables: Signals from instrumentsTorque = EngData(VAL(chan$(1)))'	N-mRPM = EngData(VAL(chan$(2)))'	Rev/minin.temp = EngData(VAL(chan$(3)))'	°Cout.temp = EngData(VAL(chan$(6)))'	°Cin.press = EngData(VAL(chan$(4)))'	barout.press = EngData(VAL(chan$(7)))'	barwall .temp = EngData(VAL(chan$(8)))'	°CCNG.mass = EngData(VAL(chan$(5)))' kg/hrCorrected Torque:torque.corr - torque - (1 .371 + 0 .000394 * RPM) 'Intensifier Speed: (speed.rat = speed ratio)intens .RPM = RPM / speed.ratPressure Ratio:p .ratio = out.press / in.pressPressure Difference:p .diff = out.press - in.press '	bar258Temperature Conversion to Kelvin:in.tempK = in .temp + Kelvin'	Kout .tempK = out .temp + Kelvin '	Kwall.tempK = wall .temp + Kelvin'	KTemperature Difference:temp.diff = out .temp - in .temp '	°CTemperature Ratiotemp.ratio = out .tempK / in .tempKForce on piston rod:rod.force = out .press * 10000 * bore " 2 * pi / 4 / 1000000'	NPolytropic Coefficient:poly.coeff = LOG(p.ratio) / (LOG(p.ratio) - LOG(temp.ratio))Piston speed max.pist .speed = ((intens .RPM * stroke * pi / 60) / 100)'	m/secVolume flow:vol.flow = CNG .mass * R.BCH * in .tempK / (1 .013 * 100000)'	Normal Cubic Meters / hrscfm = vol .flow * 35 .3146 / 60'	standard cubic feet per minuteTheoretical Temperatures:theoret.out .tempK = in.tempK * p.ratio " ((k .isentr - 1) / k.isentr)'	Kelvintheoret .out .temp = theoret.out .tempK - Kelvin'	°CDisplacement:disp.up = bore " 2 * pi / 4 * stroke / 1000000 * ncylinders'	m3 for UP strokedisp.down = disp .up - rod " 2 * pi / 4 * stroke / 1000000 * ncylinders'	m3 for DOWN strokedisp.total = disp.up + disp .down'	m3 for whole intensifier revolution (UP + DOWN)Mass:ind.mass = in .press * 100000 * disp .total / R.BCH / in.tempK * intens .RPM * 60 'kg/hr for UP and DOWN strokemass .up = in.press * 100000 * disp .up / R.BCH / in.tempK * intens .RPM * 60'kg/hr for UP strokemass .down = in.press * 100000 * disp .down / R.BCH / in.tempK * intens.RPM * 60'kg/hr for DOWNstrokeVolumetric Efficiency:vol .eff.total = CNG.mass / ind .mass * 100'	% for UP and DOWN strokevol .eff.up = CNG.mass * disp.up / disp .total / mass .up * 100 '	% for UP strokevol .eff.down = CNG.mass * disp.down / disp.total / mass .down * 100' % for DOWN strokePower:ind .power = (poly .coeff / (poly.coeff - 1)) * CNG .mass * R .BCH * in.tempK * (temp.ratio - 1) / 3600 / 1000'kW theoretical (polytropic) powertheoret.power = (k.isentr / (k.isentr - 1)) * CNG.mass * R.BCH * in .tempK * ((p .ratio " ((k .isentr - 1) / k.isentr)) -1) / 3600 / 1000'	kW theoretical (isentropic) power consumptionact.power = torque * RPM * 2 * pi / 60 / 1000	kW actual power consumptionact.power.corr = torque .corr * RPM * 2 * pi / 60 / 1000	kW corrected power consumption259Mechanical Efficiency:mech.eff = theoret.power / act .power.corr * 100	% overall (isentropic) efficiencyPower Consumption per Unit Mass:power.per.mass = act .power .corr / CNG .mass	(kW/kg/hr)CALIBRATION FILE ("intens.cal"):Single Stage Intensifier (Version 1 .2) ,CALIBRATION FILE FOR INTENSIFIER DATA ACQUISITIONLast modified 11-30-1992Intensifier Version 1 .02, Mechanical Drive, E-MotorTorque	Measurement0	Voltage Range (# defined in software)N-m	Units-5,-.02271, 1 .369, 5, 0, 0, 0, 0, 0, 0	Voltage Signal-456.56, 0, 128 .11, 456 .56, 0, 0, 0, 0, 0, 0	Corresponding Engineering UnitsWall Temperature6eC0, .5, 2, 0, 0, 0, 0, 0, 0, 00, 4 .59, 265.5, 0, 0, 0, 0, 0, 0, 0Tachometer5rpm.002, 3 .75, 4 .083, 0, 0, 0, 0, 0, 0, 00, 2106, 2298, 0, 0, 0, 0, 0, 0, 0Inlet Pressure0bar- .139, .24, .635, .83, 1 .4, 1 .77, 2 .635, 4.68, 0, 00, 13 .605, 27.21, 34 .013, 54 .421, 68.027, 102 .04, 187.075, 0, 0Outlet Pressure8bar-.465, 1 .265, 2 .18, 3 .085, 3 .98, 4 .85, 5 .725, 0, 0, 00, 68 .03, 102.04, 136 .05, 170 .07, 204.08, 238 .1, 0, 0, 0Inlet Temperature7eC0, .5, 1, 0, 0, 0, 0, 0, 0, 00, 2 .508, 94.34, 0, 0, 0, 0, 0, 0, 0Outlet Temperature6260eC0, .5, 2, 0, 0, 0, 0, 0, 0, 00, 4.59067, 265 .5, 0, 0, 0, 0, 0, 0, 0Massflow0kg/hr-5, 0, 5, 0, 0, 0, 0, 0, 0, 0-100, 0,100, 0, 0, 0, 0, 0, 0, 0CONFIGURATION FILE ("intens.cfg"):Single Stage Intensifier (Version 1 .2)CONFIGURATION FILE FOR INTENSIFIER DATA ACQUISITIONLast modified 12-02-1992Intensifier Version 1 .01 Mechanical Drive, E-MotorMode:	DOUBLE ACTINGBORE :	3 .175 cmSTROKE :	8 .255 cmROD :	1 .27 cm# of Cylinders :	1Speed Ratio :	5 .33 : 1# Variable Name Channel Constant Value1 Torque (N-m) 0 none2 Engine Speed (RPM) 2 none3 CNG Inlet Temp . (eC) 5 none4 CNG Inlet Pressure (bar) 3 none5 CNG Massflow dm/dt (kg/hr) 7 none6 CNG Outlet Temp . (eC) 6 none7 CNG Outlet Pressure (bar) 4 none8 Cyl. Wall Temp . (eC) 1 noneNumber of display columns: 8# Variable Title 1 Title 2 Units Display1 3 Inlet Temp. eC ######.#2 6 Outlet Temp eC #####.##3 4 Inlet Pressure bar ######.#4 7 Outlet Pressure bar ######.#5 11 Power kW #####.##6 15 Volum . Effic . (%) #####.##7 5 measur . Massflow kg/hr #####.##8 1 Torque (Nm) #####.##ISAAC Configuration:Number of Channels :	2Channel Label Units Slope Offset0 BDC 88 1 .000 0.0001 Pressure (PSI) 1 .000 0.000Crank angle pulses/rev =	360Data points required/rev =	360External clock divide =	1Number of consecutive cycles =	1261APPENDIX C6Calculations for Two-State Intensifier Performance:SUB PerfCalc**************************************************** Subroutine to calculate performance values ***************************************************************Set constant values:rho.air = 1 .205'	kg/m"3 Density of air @ 20eC, 1 .0133 barRGEN = 8314.34 '	J/Kmol*K GENERAL GASCONSTANTSTANDARD VALUES FOR NATURAL GAS:Density:rho.STD = 0.713142	kg/m"3Molecular weight:M.STD = 17 .1456999#	kg/KmolGasconstant:RSTD = RGEN / M .STD	J/kg*KB.C. HYDRO VALUES FOR NATURAL GAS:Density:rho.BCH = .690351	kg/m"3Molecular weight:M.BCH = 16 .597737#	kg/KmolGasconstant:R.BCH = R.GEN / M.BCH	J/kg*KKelvin = 273.15	conversion factork.isentr = 1 .3	isentropic coefficient of natural gasSet Variables : Signals from instrumentsTorque = EngData(VAL(chan$(1)))RPM - EngData(VAL(chan$(2)))in.temp = EngData(VAL(chan$(3)))out.temp = EngData(VAL(chan$(6)))in.press = EngData(VAL(chan$(4)))out .press = EngData(VAL(chan$(7)))wall.temp = EngData(VAL(chan$(8)))CNG.mass - EngData(VAL(chan$(5)))inter.temp.BC = EngData(VAL(chan$(9)))inter.temp.AC = EngData(VAL(chan$(10)))inter.press = EngData(VAL(chan$( 1 1)))BC: Before Cooling	AC: After CoolingIntensifier Speed: (speed.rat = speed ratio)intens .RPM = RPM / speed.ratN-mMotor Speed in RPMCNG inlet temperature in eCCNG end temperature (after 2nd stage) in eCCNG inlet pressure in barCNG end pressure (after 2nd stage) in barwall temperature of intensifier cylinder in eCCNG massflow in kg/hrinterstage temperature after 1st stage in eCinterstage temperature before 2nd stage in eCinterstage pressure in bar262Torque Correction:Torque .corr = Torque - (1 .371 - 0 .000394 * RPM)Pressure Ratios:p.ratio .1 = inter .press / in.press	first stagep.ratio .2 = out .press / inter .press	second stagep.ratio .total = out.press / in.press	overall pressure ratioConversion of Temperatures to Kelvin:in.tempK = in.temp + Kelvin	inlet temperature conversion to Kelvinout.tempK = out.temp + Kelvin	end temperature conversion to Kelvininter .tempK .BC = inter.temp.BC + Kelvin	interstage temperature before cooling conversion to Kelvininter .tempK .AC = inter .temp.AC + Kelvin	interstage temperature after cooling conversion to Kelvinwall .tempK = wall.temp + Kelvin	wall temperature conversion to KelvinTemperature Differences:temp.diff.1 = inter .temp.BC - in.temp	first stage in eCtemp.diff.2 = out.temp - inter.temp.AC	second stage in eCtemp.diff.cool = inter.temp.BC - inter.temp.AC	intercooling in eCtemp.diff.total = out.temp - in .temp	total temperature difference in eCTemperature Ratios:temp .ratio.1 = inter .tempK .BC / in.tempK	first stagetemp .ratio.2 = out .tempK / inter.tempK .AC	second stageVolume Flow:vol.flow = CNG .mass * R.BCH * in .tempK / (1 .013 * 100000)	Normal Cubic Meters per Hourscfm = vol .flow * 35 .3146 / 60	Standard Cubic Feet per MinuteTheoretical Discharge Temperatures:theoret.out.tempK.1 = in.tempK * p.ratio .1 " ((k.isentr - 1) / k.isentr)	first stage (in Kelvin)theoret .out.tempK .2 = inter.tempK .AC * p .ratio .2 " ((k .isentr - 1) / k .isentr) second stage (in Kelvin)theoret.out.temp.1 = theoret .out.tempK .1 - Kelvin	first stage (in Celsius)theoret.out.temp.2 = theoret .out.tempK .2 - Kelvin	second stage (in Celsius)Polytropic Coefficient:poly.coeff.1 = LOG(p.ratio.1) / (LOG(p.ratio .1) - LOG(temp.ratio.l))	first stagepoly.coeff.2 = LOG(p.ratio .2) / (LOG(p.ratio.2) - LOG(temp.ratio.2))	second stageIntensifier Displacement:FIRST STAGE:disp. l = ((bore .1 " 2) - (rod . l " 2)) * pi / 4 * stroke.1 * ncylinders .1 (in cm3)SECOND STAGE:disp.2 = ((bore .2 " 2) - (rod.2 " 2)) * pi / 4 * stroke.1 * ncylinders .2 (in cm3)Theoretical Massflow:FIRST STAGE:mass.l = in.press * 100000 * disp.l / 1000000 / R.BCH / in.tempK * intens.RPM * 60	(in kg/hr)SECOND STAGE:mass .2 = inter.press * 100000 * disp .2 / 1000000 / R.BCH / inter .tempK .AC * intens .RPM * 60 (in kg/hr)NOTE: The massflow of the second stage does not necessarily have to match with the massflow of the first stagebecause the second stage might be partially bypassedVolumetric Efficiency:FIRST STAGE:vol.eff.1 = CNG .mass / mass .l * 100	(in %)SECOND STAGE :263vol .eff.2 = CNG .mass / mass .2 * 100 (in %)OVERALL:vol .eff.total = vol .eff.1 (in %)NOTE: The volumetric efficiency for the second stage is only valid if the interstage pressure is lower than the endpressure (because in this case there is no more bypassing)Theoretical Power Consumption:FIRST STAGE:poly.power .l = (poly .coeff.1 / (poly .coeff.1 - 1)) * CNG .mass * R.BCH * (inter.temp.BC - in .tempK) / 3600 /1000	(in kW) polytropic power consumptionisentr.power.1 = (k.isentr / (k.isentr - 1)) * CNG.mass * R .BCH * in .tempK * ((p .ratio .1 " ((k .isentr - 1) /k.isentr)) - 1) / 3600 / 1000	(in kW) isentropic power consumptionSECOND STAGE:poly.power.2 = (poly.coeff.2 / (poly .coeff.2 - 1)) * CNG.mass * R.BCH * (out.tempK - inter .temp.AC) / 3600 /1000	(in kW) polytropic power consumptionisentr .power.2 = (k .isentr / (k .isentr - 1)) * CNG.mass * R.BCH * inter.tempK.AC * ((p.ratio.2 " ((k.isentr - 1) /k.isentr)) - 1) / 3600 / 1000	(in kW) isentropic power consumptionOVERALL:poly .power.tot = poly .power .1 + poly .power.2	polytropic power consumptionisentr .power .tot = isentr.power .1 + isentr.power .2	isentropic power consumptionMeasured Power Consumption:act.power = Torque .corr * RPM * 2 * pi / 60 / 1000	(in kW)Power Consumption per Unit Mass:power.per.mass - act.power / CNG.mass	(in kW/(kg/hr))Efficiencies:isentr.eff = isentr .power .tot / act.power * 100	(in %) isentropic efficiencypoly.eff = poly .power.tot / act.power * 100	(in %) polytropic efficiencyCALIBRATION FILE ("2stage.cal"):Two-Stage Intensifier (Version 2 .1)CALIBRATION FILE FOR INTENSIFIER DATA ACQUISITIONLast modified 11-02-1993Intensifier Version 2.01, 2-Stage, Mechanical Drive, E-MotorTorque	Measurement0	Voltage Range (# defined in software)N-m	Units-5,- .0134, 1 .3783, 5, 0, 0, 0, 0, 0, 0	Voltage Signal-456.56, 0, 128 .11, 456 .56, 0, 0, 0, 0, 0, 0	Corresponding Engineering UnitsWall Temperature6BC0, .5, 2, 0, 0, 0, 0, 0, 0, 00, 4.59, 265 .5, 0, 0, 0, 0, 0, 0, 0Tachometer5264rpm.002, 3 .75, 4 .083, 0, 0, 0, 0, 0, 0, 00, 2106, 2298, 0, 0, 0, 0, 0, 0, 0Inlet Pressure0bar-.139, .24, .635, .83, 1 .4, 1 .77, 2 .635, 4 .68, 0, 00, 13 .605, 27.21, 34 .013, 54.41, 68 .027, 102 .04, 187.075, 0, 0Outlet Pressure8bar-.465, 1 .265, 2.18, 3 .085, 3 .98, 4 .85, 5 .725, 0, 0, 00, 68 .03, 102.04, 136.05, 170.07, 204.08, 238.1, 0, 0, 0Inlet Temperature700, .5, 1,0,0,0,0,0,0,00, 2 .508, 94.34, 0, 0, 0, 0, 0, 0, 0Outlet Temperature6eC0, .5, 2, 0, 0, 0, 0, 0, 0, 00, 4 .5907, 265 .5, 0, 0, 0, 0, 0, 0, 0Massflow0kg/hr-5, 0, 5, 0, 0, 0, 0, 0, 0, 0-100, 0, 100, 0, 0, 0, 0, 0, 0, 0Interstage Press.4bar1,6 .4,0,0, 0, 0, 0, 0, 0, 00,221 .1,0,0, 0, 0, 0, 0, 0, 0none00,0,0,0,0,0,0,0,0,00, 0, 0, 0, 0, 0, 0, 0, 0, 0Interst . Temp. BC6aC0, .5, 2, 0, 0, 0, 0, 0, 0, 00, 4 .5907, 265 .5, 0, 0, 0, 0, 0, 0, 0Interst . Temp. AC6eC0, .5, 2, 0, 0, 0, 0, 0, 0, 00, 4 .5907, 265 .5, 0, 0, 0, 0, 0, 0, 0265CONFIGURATION FILE ("2stage.cfg"):Two-Stage Intensifier (Version 2.1)CONFIGURATION FILE FOR INTENSIFIER DATA ACQUISITIONLast modified 11-02-1993Intensifier Version 2.01 Mechanical Drive, E-MotorDimensions of the 2-Stage Intensifier:FIRST STAGE:BORE: 3 .175 cmSTROKE: 8 .255 cmROD: 0 cm# of Cylinders : 1SECOND STAGE:BORE : 3.175 cmSTROKE: 8 .255 cmROD : 2.54 cm# of Cylinders : 1Speed Ratio: 5 .33 : 1# Variable Name Channel Constant Value1 Torque (N-m) 0 none2 Engine Speed (RPM) 2 none3 Inlet Temp . (eC) 5 none4 Inlet Pressure (bar) 3 none5 Massflow dm/dt (kg/hr) 7 none6 Outlet Temp. (eC) 6 none7 Outlet Pressure (bar) 4 none8 Wall Temp . (eC) 1 none9 Interstage. Temp . BC (eC) 10 none10 Interstage. Temp. AC (eC) 11 none11 Interstage . Pressure (bar) 8 noneNumber of display columns : 8# Variable Title 1 Title 2 Units Display1 4 Inlet Press. bar ######.#2 11 Inter . Press . bar ###### .#3 7 Outlet Press . bar ###### .#4 5 Massflow kg/hr ##### .##5 1 Torque Nm #####.##6 9 Inter. Temp. BC eC ###### .#7 10 Inter . Temp. AC eC ###### .#8 6 Outlet Temp . eC ###### .#ISAAC Configuration:Number of Channels:	3Channel Label Units Slope Offset0 BDC 88 1 .000 0.0001 Pressure (PSI) 1 .000 0.0002 Pressure (PSI) 1 .000 0.000Crank angle pulses/rev = 360Data points required/rev = 360External clock divide = 1Number of consecutive cycles = 1266APPENDIX D1The Effect of CoolingNO COOLINGIntensifier Speed 150 rpm, Discharge Pressure 208 .5 barln(V)_ Inlet Pressure 20.2 bar	— THEORY 2 .5% Clearance VolumeFigure Dl .l: The ln(P)-In(V) diagram at design pressure ratio (no cooling)AIR COOLINGIntensifier Speed 150 rpm, Discharge Pressure 201 bar2.165 .554.532.5Pressure Ratio : 9 .95 : 1n = 1 .289n = 1 .191- 43 .5Inlet Pressure 20.2 bar2	3	4In(V)T THEORY 2 .5% Clearance Volume0 5Figure D1.2: The In(P)-ln(V) diagram at design pressure ratio (air cooling)26765 .554 .543.532.5WATER COOLINGIntensifier Speed 150 rpm, Discharge Pressure 205 .7 bar0	2	3	4	5ln(V)Inlet Pressure 19.5 bar	— THEORY 2.5% Clearance VolumeFigure D1.3 : The ln(P)-ln(V) diagram at design pressure ratio (water cooling)268APPENDIX D2SINGLE STAGE INTENSIFIER CAPACITYFigure D2.1: Single stage intensifier capacity (in Nm3/hr)Figure D2.2: Single stage intensifer capacity (in SCFM)269APPENDIX D3The Effect of Chan ges in Discharge PressureFigure D3.1: Volumetric efficiency at 200 bar discharge pressure and 150 rpmVOLUMETRIC EFFICIENCYDISCHARGE PRESSURE: 200 bar, INf ENSIFIER SPEED : 150 rpm90408	9	10	11PRESSURE RATIO (P2/P1). EXPERIMENTAL DATA	LINEAR REGRESSION7 1312VOLUMETRIC EFFICIENCYDISCHARGE PRESSURE : variable 170-180 bar, INTENSIFIER SPR.: I) : 150 rpm9080706050408Volumetric Efficiency at Design Pressure Ratio: 68 .1%13 149	10	11	12PRESSURE RATIO (P2/P I). EXPERIMENTAL DATA	LINEAR REGRESSIONFigure D3.2: Volumetric efficiency at 175 bar discharge pressure and 150 rpm27090408VOLUMETRIC EFFICIENCYDISCHARGE PRESSURE : variable 145-160 bar, INTENSIFIER SPEED : 150 rpm9	10	11	12PRESSURE RATIO (P2/P 1)EXPERIMENTAL DATA	LINEAR REGRESSION13VI"vim-T •.Volumetric Efficiency at Design Pressure Ratio : 70.8%Figure D3.3: Volumetric efficiency at 150 bar discharge pressure and 150 rpm271APPENDIX ElTWO-STAGE INTENSIFIER CAPACITYFigure E1 .1: Two-Stage Intensifier Capacity in Normal Cubic Meters per Hour.Figure E1.2 : Two-Stage Intensifier Capacity in Standard Cubic Feet per Minute.CNG VOLUME FLOW (Nm3/hr)DISCHARGE PRESSURE : 200 BAR2	4	6	8	10	12	14	16OVERALL PRESSURE RATIOFill 1'E1.) CURVE35302 25v2015J 1050272APPENDIX E2PRESSURE-VOLUME DIAGRAMS OF THE TWO-STAGE INTENSIFIER,Figure E2 .1: The p-V diagram at 100 rpm and 4 .6 : 1 overall pressure ratio.CYLINDER PRESSURE vs VOLUMEStage 1&2, Intensifier Speed : 100 rpmInlet Pressure: 43.3 barInterstage Pressure: 134 barDischarge Pressure : 200 barOverall Pressure Ratio 4 .6 : 1Stage 2Stage 10	to	20	30	40	5o	60	70	80VOLUME (crrr"3). Experimental Data	Theory2502005o02502005o0CYLINDER PRESSURE vs VOLUMEStage 1&2, Intensifier Speed : 100 rpm0	to	20	30	40	50	60	70	80VOLUME (cm^3). Experimental Data	TheoryInlet Pressure : 29.1 barInterstage Pressure: 100 barDischarge Pressure : 197 barOverall Pressure Ratio 6 .7 : 1Figure E2 .2 : The p-V diagram at 100 rpm and 6 .7 : 1 overall pressure ratio .273Figure E2.3 : The p-V diagram at 150 rpm and 10 .4 : 1 overall pressure ratio.CYLINDER PRESSURE vs VOLUMEStage 1&2, Intensifier Speed : 150 rpm10	20	30	40VOLUME (cm^3). E,q erimental Data	Theory706050 80CYLINDER PRESSURE vs VOLUMEStage 1X2, Intensifier Speed : 150 rpm2502005000 20to 50 so60 7030	40VOLUME (cm^3)Experimental Data	TheoryFigure E2.4 : The p-V diagram at 150 rpm and 8 .9 : 1 overall pressure ratio .274CYLINDER PRESSURE vs VOLUMEStage 1&2, Intensifier Speed : 150 rpm0	10	20	30	40	50	60	70	80VOLUME (cm^3). Experimental Data	TheoryFigure E2.5: The p-V diagram at 150 rpm and 6 .4 : 1 overall pressure ratio.CYLINDER PRESSURE vs VOLUMEStage 1&2, Intensifier Speed : 200 rpm2502005000 10 20	30	40VOLUME (cm^3). Experimental Data	Theory50 60 70 80Figure E2 .6: The p-V diagram at 200 rpm and 12 .1 : 1 overall pressure ratio .


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