Applied Science, Faculty of
Engineering, School of (Okanagan)
DSpace
UBCO
Hossain, Syed Tanjid
Master of Applied Science - MASc
University of British Columbia
The High-Pressure Direct Injection of Natural Gas in a diesel engine is one of the innovative solutions to reduce diesel engine pollutants. The quantity and quality of the fuel injected is highly depended on the spray pattern, which controls the air and NG fuel mixture in the combustion cylinder. Fuel release is regulated by a needle, that is linked to a solenoid actuator through a hydraulic system. An actuator with position feedback can control the spray pattern of a fuel injector by controlling the needle displacement of the injector. However, the relationship between the actuator displacement and injector needle displacement is unknown. Therefore, an Actuator Test Fixture (ATF) is designed that simulates the injection process of a typical hydraulically actuated fuel injector. It allows the researcher to study and ultimately model the relationship between actuator displacement and needle motion. In contrary to conventional fuel injectors, the ATF provides measurement access to all model parameters and allows their characterization. Thus, a characterized model is created that describes the relationship between the actuator displacement and ATF needle motion. The model of the ATF combines both the mechanical aspect of the components and the hydraulic mechanics of the fuel. There are two inputs to this model, the experimentally measured actuator displacement and experimentally measured supply pressure. ATF needle lift and control chamber pressures are the outputs of the model. The experimental-model comparison showed good agreement. The model is verified across several different operating points by varying the supply pressure and the actuator pulse duration to increase confidence in the experimental-model results. At high supply pressure, the model fits very well but as the pressure reduces the model starts to show discrepancy. This is likely due to un-modeled friction effects in the ATF. Nevertheless, the experimental-model control chamber pressures and needle lifts are very similar.
DESIGN AND CHARACTERIZATION OF AN ACTUATOR TEST FIXTURE FOR DIESEL FUEL INJECTORS by Syed Tanjid Hossain B.Eng. Hons., International Islamic University Malaysia, 2016 A THESIS SUBMITTED IN PARTIAL FULFILLMENT OF THE REQUIREMENTS FOR THE DEGREE OF MASTER OF APPLIED SCIENCE in THE COLLEGE OF GRADUATE STUDIES (Mechanical Engineering) THE UNIVERSITY OF BRITISH COLUMBIA (Okanagan) December 2018 © Syed Tanjid Hossain, 2018 ii The following individuals certify that they have read, and recommend to the Faculty of Graduate and Postdoctoral Studies for acceptance, the dissertation entitled: DESIGN AND CHARACTERIZATION OF AN ACTUATOR TEST FIXTURE FOR DIESEL FUEL INJECTORS submitted by Syed Tanjid Hossain in partial fulfillment of the requirements for the degree of Master of Applied Science Examining Committee: Dr Rudolf Seethaler, School of Engineering, UBC Supervisor Dr. Homayoun Najjaran, School of Engineering, UBC Supervisory Committee Member Dr. Yang Cao, School of Engineering, UBC Supervisory Committee Member Dr. Sumi Siddiqua, School of Engineering, UBC University Examiner iii Abstract The High-Pressure Direct Injection of Natural Gas in a diesel engine is one of the innovative solutions to reduce diesel engine pollutants. The quantity and quality of the fuel injected is highly depended on the spray pattern, which controls the air and NG fuel mixture in the combustion cylinder. Fuel release is regulated by a needle, that is linked to a solenoid actuator through a hydraulic system. An actuator with position feedback can control the spray pattern of a fuel injector by controlling the needle displacement of the injector. However, the relationship between the actuator displacement and injector needle displacement is unknown. Therefore, an Actuator Test Fixture (ATF) is designed that simulates the injection process of a typical hydraulically actuated fuel injector. It allows the researcher to study and ultimately model the relationship between actuator displacement and needle motion. In contrary to conventional fuel injectors, the ATF provides measurement access to all model parameters and allows their characterization. Thus, a characterized model is created that describes the relationship between the actuator displacement and ATF needle motion. The model of the ATF combines both the mechanical aspect of the components and the hydraulic mechanics of the fuel. There are two inputs to this model, the experimentally measured actuator displacement and experimentally measured supply pressure. ATF needle lift and control chamber pressures are the outputs of the model. The experimental-model comparison showed good agreement. The model is verified across several different operating points by varying the supply pressure and the actuator pulse duration to increase confidence in the experimental-model results. At high supply pressure, the model fits very well but as the pressure reduces the model starts to iv show discrepancy. This is likely due to un-modeled friction effects in the ATF. Nevertheless, the experimental-model control chamber pressures and needle lifts are very similar. v Lay Summary Due to increased regulation and awareness of the negative health effect of diesel engine emission among people, companies are funding research to look for alternate methods of fuel injection. Natural Gas/Diesel Dual Fuel Injection systems reduces engine emissions. In order to improve their injection process, it is important to understand the injector needle dynamics, which injects fuel into the combustion chamber. The goal of this thesis is to design and fabricate an actuator test fixture(ATF), which can be used to characterize the injection parameters. This thesis also presents a model that lays the groundwork for better controlling the injector needle lift. The ATF resembles a typical hydraulically actuated fuel injector. The model can predict not only full lifts, but also partial lifts. Therefore, new actuators can be modelled and designed, and can be validated using this ATF model. vi Preface This thesis is based on the research work completed in the School of Engineering at the University of British Columbia, Okanagan under the supervision and guidance of Dr Rudolf Seethaler. vii Table of Contents Abstract ......................................................................................................................................... iii Lay Summary .................................................................................................................................v Preface ........................................................................................................................................... vi Table of Contents ........................................................................................................................ vii List of Figures .............................................................................................................................. xii List of Symbols ........................................................................................................................... xvi List of Abbreviations ................................................................................................................. xix Acknowledgements ......................................................................................................................xx Dedication ................................................................................................................................... xxi Chapter 1: Introduction ................................................................................................................1 1.1 Diesel Engine Emissions................................................................................................. 1 Emission Reduction .................................................................................................... 2 1.2 Natural Gas as Fuel Source ............................................................................................. 3 Natural Gas in Compression Ignition Engines............................................................ 4 1.3 Objective and Scope ....................................................................................................... 5 1.4 Procedures to Achieve Objectives .................................................................................. 6 1.5 Thesis Structure .............................................................................................................. 8 Chapter 2: Literature Review .......................................................................................................9 2.1 Natural Gas Fuel Injection Systems ................................................................................ 9 2.2 Fuel Injectors ................................................................................................................ 10 2.3 Actuators ....................................................................................................................... 12 viii Actuation by Camshaft ............................................................................................. 12 Solenoid Actuator ..................................................................................................... 13 Working Principle of the Solenoid Actuator .................................................... 14 Piezoelectric Actuator ............................................................................................... 14 Working Principle of the Piezoelectric Actuator .............................................. 15 The Hydraulically Actuated Fuel Injector ................................................................ 16 2.4 Experimental Characterization of Fuel Injectors .......................................................... 19 2.5 Mathematical Modelling of Fuel Injectors ................................................................... 21 Chapter 3: ATF Design, Analysis and Apparatus ....................................................................26 3.1 Mechanical Design of ATF ........................................................................................... 29 ATF Body ................................................................................................................. 30 Solenoid Actuator ............................................................................................. 31 Brass Sleeve .............................................................................................................. 32 Plunger and Needle ................................................................................................... 33 End Stop .................................................................................................................... 35 Compression Nut ...................................................................................................... 36 Custom Hydraulic Fitting ......................................................................................... 37 Actuator Preload Screw ............................................................................................ 37 3.2 Design Analysis ............................................................................................................ 39 ATF and Actuator Clamping .................................................................................... 40 3.3 Apparatus ...................................................................................................................... 43 ATF Test Setup ......................................................................................................... 44 Hydraulic Components ..................................................................................... 45 ix Viscor Hydraulic Fluid ..................................................................................... 46 Relief Valve ...................................................................................................... 47 In-line Filter ...................................................................................................... 48 3.4 Measurements ............................................................................................................... 48 High-Speed Laser Sensor .......................................................................................... 48 Piezoelectric Pressure Sensor ................................................................................... 49 3.5 Control System and Data Acquisition........................................................................... 50 3.6 Safety Review ............................................................................................................... 51 Air Compressor ......................................................................................................... 52 Hydraulic Pump ........................................................................................................ 52 Failure of Hydraulic Components. ............................................................................ 52 Chapter 4: Mathematical Modelling of Actuator Test Fixture (ATF) and Experimental Characterization ..........................................................................................................................54 4.1 Mechanical Modelling of the ATF ............................................................................... 55 The equation of Motion of the Needle ...................................................................... 56 Lifting Force Spring Stiffness ................................................................................... 58 End Stop Modelling .................................................................................................. 59 Closed End Stop ................................................................................................ 61 Open End Stop .......................................................................................................... 64 Identification of the Lifting Force Spring Preload .................................................... 67 4.2 Hydraulic Modelling ..................................................................................................... 68 Inlet Orifice Modelling ............................................................................................. 68 Outlet Orifice Modelling........................................................................................... 69 x Leakage Modelling ................................................................................................... 70 Identification of Flow Constants and Leakage ......................................................... 71 Outlet and Inlet Flow Characterization ............................................................. 71 Leakage Characterization: ................................................................................ 72 Verification of the Inlet, Outlet, and Leakage Flow Type ........................................ 72 Hydraulic Subsystem Modelling............................................................................... 74 4.3 System Model ............................................................................................................... 75 Characterized Model Result ...................................................................................... 75 Chapter 5: Conclusion .................................................................................................................90 5.1 Limitations of ATF ....................................................................................................... 91 5.2 Recommendation for Future Work ............................................................................... 92 ATF Improvement .................................................................................................... 92 Model Improvement.................................................................................................. 93 Bibliography .................................................................................................................................94 Appendices ....................................................................................................................................99 Appendix A - Mechanical Drawings ........................................................................................ 99 A.1 ATF Body Front View Cross-Section....................................................................... 99 A.2 ATF Body ............................................................................................................... 100 A.3 Plunger .................................................................................................................... 101 A.4 Needle ..................................................................................................................... 102 A.5 Brass Sleeve ............................................................................................................ 103 A.6 Closed End Stop ...................................................................................................... 104 A.7 Compression Nut .................................................................................................... 105 xi A.8 Lock Nut ................................................................................................................. 106 Appendix B - Experimental Procedures ................................................................................. 107 Appendix C - Calculations ...................................................................................................... 108 C.1 Thickness of ATF Body .......................................................................................... 108 C.2 Thread Stripping Calculation .................................................................................. 108 C.3 ATF and Actuator Clamping Force ........................................................................ 109 C.4 Clamping Force ....................................................................................................... 110 Appendix D - Matlab Code ..................................................................................................... 110 xii List of Figures Figure 1.1: Westport HPDI Injector Tip Assembly. ....................................................................... 5 Figure 2.1: Schematic of Port Fuel Injector. ................................................................................. 11 Figure 2.2: Cam Plunger Arrangement. ........................................................................................ 12 Figure 2.3: Traditional Hydraulically Actuated Fuel Injector. ..................................................... 13 Figure 2.4: Piezoelectric Fuel Injector. ......................................................................................... 15 Figure 2.5: Traditional Hydraulically Actuated Fuel Injector. ..................................................... 18 Figure 2.6: a) Diesel Injector Assembly. b) Diesel Needle Free Body Diagram. ......................... 23 Figure 2.7: a) Fuel Injector. b) Schematic diagram of the control chamber. ................................ 24 Figure 3.1: a) Exploded View of the ATF Design. b) Fabricated Prototype. ............................... 26 Figure 3.2: a) Cross-section of the ATF. b) Cross-section of Diesel Fuel Injector. ..................... 28 Figure 3.3: Components of the ATF. ........................................................................................... 29 Figure 3.4: Cross section of the ATF. ........................................................................................... 31 Figure 3.5: Solenoid Actuator. ...................................................................................................... 32 Figure 3.6: Brass Sleeve. .............................................................................................................. 33 Figure 3.7: a) Plunger. b) Needle. c) Assembly of Plunger and Needle, ...................................... 34 Figure 3.8: Match Fit between Plunger and Brass Sleeve. ........................................................... 34 Figure 3.9: Closed-End Stop. ........................................................................................................ 35 Figure 3.10: Application of lock nut. ............................................................................................ 36 Figure 3.11: Compression Nut. ..................................................................................................... 37 Figure 3.12: Custom Hydraulic Fittings. a) Inlet and Return Line Fitting. b) Pressure Sensor Fitting. ........................................................................................................................................... 37 xiii Figure 3.13: Schematic of the Actuator ........................................................................................ 38 Figure 3.14: Preload Actuator Screw. a) Solenoid Actuator. b) Preload Actuator Screw Assembly....................................................................................................................................... 39 Figure 3.15: Modified Actuator Preload Screw. ........................................................................... 39 Figure 3.16: a) Actuator and ATF Assembly. b) O-ring on top of the ATF. ................................ 40 Figure 3.17: a) Schematic of the ATF and Actuator in Closed Position. b) Schematic of the Actuator in Open Position. ............................................................................................................ 41 Figure 3.18: a) Copper Gasket attached to the ATF. b) Copper Gasket Design. .......................... 41 Figure 3.19: ATF and Actuator Clamp. a) Clamp Design. b) Clamp Prototype. ......................... 42 Figure 3.20: Schematic of the Setup ............................................................................................. 44 Figure 3.21: ATF Test Setup. ....................................................................................................... 45 Figure 3.22: Hydraulic Fittings. .................................................................................................... 46 Figure 3.23: a) Air Compressor. b) Hydraulic Pump .................................................................... 47 Figure 3.24: High-Pressure Relief Valve. ..................................................................................... 47 Figure 3.25: In-line Filter. ............................................................................................................. 48 Figure 3.26: HSV 2002 High-Speed Vibrometer. a) Laser beam steering rail and Sensor Head. b) Signal conditioning unit ................................................................................................................ 49 Figure 3.27: a) Piezoelectric Pressure Sensor. b) Controller ........................................................ 50 Figure 3.28: 20A Current Generated Using Simulink .................................................................. 50 Figure 3.29: Schematic of the Electronics .................................................................................... 51 Figure 3.30: Safety Enclosure. ...................................................................................................... 53 Figure 4.1: Simulink Model. a) Overall ATF Model. b) The subsystem of the ATF. .................. 55 Figure 4.2: Simplified drawing of the ATF .................................................................................. 56 xiv Figure 4.3: a) ATF Needle in Closed Position b) ATF Needle in Open Position ......................... 57 Figure 4.4: Mechanical Subsystem ............................................................................................... 58 Figure 4.5: Lifting Force Spring Stiffness. .................................................................................. 59 Figure 4.6: Stiffness and Damping Model. ................................................................................... 60 Figure 4.7: Closed End Stop Model of the ATF. .......................................................................... 62 Figure 4.8: a) Experimental Validation of Closed End Stop Deformation. b) Equivalent ATF Needle Lift vs Control Chamber Force. ........................................................................................ 63 Figure 4.9: FEM Analysis of Closed End Stop. ........................................................................... 64 Figure 4.10: a) Experimental Validation of Open End Stop Deformation. b) ATF Needle Lift vs Control Chamber Pressure Graph ................................................................................................. 65 Figure 4.11: FEM Analysis of Plunger ......................................................................................... 66 Figure 4.12: a) Experimental Verification of Lifting Force Spring Preload. b) Control Chamber Pressure vs ATF Needle Lift......................................................................................................... 67 Figure 4.13: Hydraulic Subsystem. ............................................................................................... 68 Figure 4.14: Experimental – Model Comparisons for Various Pressures at 2ms Pulse Duration and Respective Flow Rates. a) 25.6MPa. b) 24.6MPa. c) 22.6MPa. d) 21.6MPa. e) 20.6MPa. .. 80 Figure 4.15: Experimental and Simulated Injector Needle Lift. ................................................... 81 Figure 4.16: Experimental – Model Comparisons for Various Pressures at 1.5ms Pulse Duration and Respective Flow Rates. a) 25.6MPa. b) 24.6MPa. c) 22.6MPa. d) 21.6MPa. e) 20.6MPa. .. 83 Figure 4.17: Experimental – Model Comparisons for Various Pressures at 1ms Pulse Duration and Respective Flow Rates. a) 25.6MPa. b) 24.6MPa. c) 22.6MPa. d) 21.6MPa. e) 20.6MPa. .. 85 Figure 4.18: Experimental – Model Comparisons for Various Pressures at 0.8ms Pulse Duration and Respective Flow Rates. a) 25.6MPa. b) 24.6MPa. c) 22.6MPa. d) 21.6MPa. e) 20.6MPa. .. 87 xv Figure 4.19: Experimental – Model Comparisons for Various Pressures at 0.5ms Pulse Duration and Respective Flow Rates. a) 25.6MPa. b) 24.6MPa. c) 22.6MPa. d) 21.6MPa. e) 20.6MPa. .. 89 xvi List of Symbols  : Spring cavity area  : Needle differential Area : Area of contact on the needle  : Inlet orifice area of the control chamber  : Inlet orifice area  : Needle valve area in one end  : Needle valve area in other end  : Outlet orifice area of cothe ntrol chamber  : Servo piston area  : Needle area  : Actuator outlet orifice area  : Damping coefficient of the needle valve  : Preload spring damping  : Inlet and outlet flow coefficient  : Inlet flow coefficient of the control chamber  : Outlet flow coefficient of the control chamber  : Damping coefficient of spring : Friction force : Preload of lifting force spring  : Gravitional acceleration xvii ℎ : Spring displacement  : Lifting force spring stiffness  : Nozzle spring constant  : Spring constant  : Mass of needle valve  : Mass of needle and plunger : Mass of needle  : Mass of servo piston  : Atmospheric pressure  : Control chamber pressure  : Spring cavity pressure  : Hollow chamber pressure  : Supply pressure  : Outlet drain pressure  : Pressure at primary needle volume  : Common-rail pressure  : Sac volume pressure  : Fuel flow rate into control chamber  : Inlet flow rate  : Leakage flow rate  : Fuel flow rate out from the control chamber  : Outlet flow rate xviii  : Leakage resistance  : Velocity of needle  : Effective chamber volume  : Control chamber volume  : Initial volume of the control chamber  : Needle valve displacement  : Initial setting of needle valve spring  : free length of needle valve spring  : Piston displacement __: Maximum needle lift  : Bulk modulus of elasticity in the control chamber .  : Bulk modulus of elasticity  : Installation angle of the injector  : Fuel density ∑  : Net flow rate xix List of Abbreviations ATF : Actuator Test Fixture CI : Compression Ignition CNC : Compressed Natural Gas CO : Carbon Monoxide CO2 : Carbon Dioxide EGR : Exhaust Gas Recirculation HC : Hydrocarbon HCCI : Homogenous Charge Compression Ignition HPDI : High-Pressure Direct-Injection HSDI : High Speed Direct Injection NG : Natural Gas NOx : Oxides of Nitrogen PM : Particulate Matter SI : Spark Ignition SMO : Sliding Mode Observer xx Acknowledgements I offer my enduring gratitude to the faculty, staff and my fellow students at UBC, who have inspired me to continue my work in this field. I put forward my sincere appreciation and gratitude towards my supervisor Dr Rudolf Seethaler for his invaluable support, encouragement and sharing his excellent knowledge on this field. I thank him for providing coherent answers to my endless questions. He has been a great motivation for me. I would also like to thank my lab colleague Brad Reinholz for his assistance throughout my curriculum. I would also like to offer my gratitude to the UBCO machine shop technicians especially Raymond Seida for his assistance in fabricating the Actuator Test Fixture. Special thanks are owed to my parents, who have supported me throughout my years of education, both morally and financially. xxi Dedication DEDICATED TO MY MOM1 Chapter 1: Introduction Diesel-powered engines have been the most common mode of transportation for goods and services by the rail, automobile, and marine industries. Its high thermal efficiency, durability, and desirable load characteristics have made it a popular choice of fuel, but its high carbon footprint, along with high NOx and PM emissions, have subjected it to increasingly more stringent emission standards and regulations. According to US EIA (2013) consumption of diesel has increased by a significant amount from the 1980s - consumption has doubled in 30 years. Thus, governments all over the world have introduced new standards and tight regulations to control the emission of NOx and PM. Over the last few years, people have been made aware of all these harmful externalities of diesel emissions. Therefore, many ideas are being developed to replace the main fuel in diesel engines and use alternate and more environment-friendly fuels. Dongiovanni and Coppo (2010) and Sekens & Somers (2005) concluded that manufacturers keep the following criteria as their top priorities: power output, fuel efficiency, reliability, and maintenance cost. Fuel injection plays a key role in achieving high performance, as well as fuel and engine efficiency in diesel engines. 1.1 Diesel Engine Emissions Diesel engines emission contribute to environmental damage, global warming, and poor air quality (Lloyd and Cackette, 2001). Sydbom et al. (2001) confirm that these emissions are linked to health 2 issues like cough, nausea, and lung function failure due to the high concentration of carbon PM called soot. Soot is black smoke that comes out of the exhaust. The Hydrocarbon (HC) in fuel reacts with oxygen during the complete combustion process to form CO2, water, heat, and pressure in the cylinder. Carbon monoxide (CO) or soot is formed during incomplete combustion. When the air and fuel mixture lack oxygen during combustion, it does not burn completely. Therefore, there is insufficient oxygen to oxidize carbon atoms to carbon dioxide (CO2). Fuel is not used efficiently in this case. The size of soot can vary from the nanometer scale to micrometer range depending on the combustion process. Soot can be classified in the following sizes. PM10 can be classified as 10 µm size in diameter. If the PM is larger than this, then it can be easily filtered by our nose, but if it is less than 10 µm, it can enter the lungs and cause lung cancer. If the PM is smaller than 1 µm, it can enter the bloodstream through the lungs which can cause heart attacks (World Health Organization, 2009). One of the by-products of diesel emission is NOx, which is a proportion of 90% NO and 10% NO2 (Birger, 2010). It is capable of destroying the ozone layer, as well as causing acidic rain and harming respiratory systems (Bernard et al. 2001). NOx is produced when the temperatures and pressures are high at stoichiometric region during heavy load. NO2 reacts with the OH in the atmosphere and produces nitric acid (HNO3), which accumulates and pours as acid rain. Emission Reduction In recent years, due to increased emission regulation and awareness of the negative health effect among people, companies, and operators are funding research to look for alternate methods of fuel injection. Exhaust Gas Recirculation (EGR) reduces combustions emission by 10 – 20%, reducing 3 the NOx production by lowering the combustion temperature. However, this increases the production of soot due to reduced oxygen. There’s a trade-off in this EGR system strategy which can be further tweaked by changing the injection parameters (Kheirkhah, 2015). McTaggart-Cowan (2006) found that increasing the injection pressures can reduce PM emissions. High-pressure diesel injection leads to good atomization and fuel vaporization which reduces soot production, but this increases NO2 emissions. The paired-nozzle injection can reduce the production of soot. Increasing the paired-nozzle angle and spacing leads to better fuel atomization and vaporization compared to the single-hole nozzle. However, this technique also has a trade-off; it increases NOx production (Gao et al., 2009 and Nishida et al., 2009). Stanglmaier and Roberts (1999) found that in Homogenous Charge Compression Ignition (HCCI), soot and NOx emission are reduced since the combustion is taking place in a low temperature as lean mixture of air and fuel is produced in the very early injection. But high CO and unburned hydrocarbons are produced. 1.2 Natural Gas as Fuel Source The increasing concerns about the harmful effects of the emission of pollutant gases have created a platform for natural gas as an alternative fuel source for road vehicles due to its lower engine-out emissions and environment-friendly technology. Semin and Rosli (2008) reported that Compressed Natural Gas (CNG) has been used as a transport engine fuel for a few decades. Natural 4 Gas (NG) primarily contains methane, propane, and ethane, along with the proportion of carbon dioxide, nitrogen, helium, and water vapour (McTaggart-Cowan, Reynolds and Bushe, 2006). Natural Gas has many advantages over conventional gasoline, such as it is non-toxic, lighter than air, has a low cost per energy unit of NG, and has lower greenhouse emissions etc. (Semin and Rosli, 2008). According to Poulton (1994), a Natural Gas-powered engine will provide similar power and efficiency, but lower emissions than an equivalent diesel engine vehicle. NG can uniformly mix with combustion air and reduce PM and NOx emission due to fewer carbon bonds. CNG can be used in both spark ignition (SI) engines and compression ignition (CI) engines. In SI engines, a mixture of air and gasoline is taken into the engine, compressed and ignited whereas, in CI engines, only air is taken into the engine and compressed to a higher temperature (Cheng, 2008). However, the biggest problem with NG in a CI engine is that it requires very high temperature in the combustion chamber for auto-ignition (Naber et al., 1994). Many methods have been developed to overcome this challenge. Both spark plugs and pilot fuels have been used for initiating ignition (Korakianitis et al., 2011). The overall compression ratio of the engine can be reduced by adding a pilot diesel injection (Semin and Rosli, 2008). Natural Gas in Compression Ignition Engines To efficiently use NG in heavy duty CI engines, it needs to be injected directly in the cylinder. Since NG requires a high temperature for ignition, a pilot fuel is injected before the injection of NG. Westport Fuel Systems Inc. has developed a dual fuel injector where both NG and diesel are injected using the same injector. Figure 1.1 shows injector tip assembly of the Westport HPDI injector which uses both diesel and gas. As can be seen from Figure 1.1, there is a gas needle and 5 a diesel needle. The needles are controlled separately by different actuators. Diesel is injected through the holes from the injector body followed by gas from a separate hole. Figure 1.1: Westport HPDI Injector Tip Assembly. Reproduced from “Engine and Fuel System Integration,” by Westport Fuel Systems Inc, n.d, Retrieved from https://www.westport.com/old-pages/combustion/hpdi/integration. Copyright 2018 by Westport Fuel Systems Inc. Reprinted with permission. Dumitrescu (2000) and Duggal et al., (2004) verified that compared to conventional diesel engines, the Westport HPDI system reduces NOx emissions by 40-50%, PM by approximately 80%, and it has the same efficiency. 1.3 Objective and Scope Previous studies focused on nozzle spray patterns, co-injection strategies, and flow characteristics. The current study aims to lay out the groundwork for controlling the needle lift of diesel fuel injectors. In these injectors, a hydraulic circuit links the actuator displacement to the needle 6 displacement. However, the relationship between actuator displacement and injector needle displacement is unknown, since the parameters required for modelling this relationship are not physically accessible with conventional fuel injectors. Thus, the objectives of this thesis are as follows: i) Create a test setup that, while behaves like a typical hydraulically actuated fuel injector, provides measurement access to the parameters necessary to calibrate a model: Actuator displacement, Injector displacement, line pressure, and hydraulic chamber pressure. ii) Create and validate a model that describes the relationship between actuator and injector needle displacement. 1.4 Procedures to Achieve Objectives To achieve the above-mentioned objectives, the following steps are followed: i. Designing the Actuator Test Fixture (ATF) The ATF is designed using Solidworks and the geometry is custom made. Westport provided guidance in designing the ATF based upon their experience with designing HPDI injectors. Westport also supplied a prototype hydraulic fuel actuator to be used in the experimental setup. Characteristic parameters like chamber volume and match fit between the plunger and ATF body are matched to a typical fuel injector. FEM analysis is performed on all the components with a safety factor of 3 to check the strength of the material under 27.6MPa pressure. Thread stripping calculations are also performed where needed. The ATF is then fabricated in the UBCO in-house machine shop. 7 ii. Modelling of the ATF In this stage, knowledge of fluid and hydraulic equations are required. The model describes the relationship between actuator and injector needle displacement. There are two inputs to this model, the experimentally measured actuator needle lift and experimentally measured supply pressure. The model predicts the actuator needle lift and control chamber pressures which are the outputs. iii. Selection of the key components of the test-rig The hydraulic pressure is generated by an air driven hydraulic pump. Therefore, components such as air compressor, hydraulic pump, hydraulic fittings, and sensors need to be selected. iv. Enclosure Manufacturing A rigid frame enclosure is designed and built to house the hydraulic fittings and sensors necessary for performing experiments. v. Test-rig commissioning The test-rig is pressurized, and pressure sensors and displacement sensor readings are calibrated. vi. ATF Characterization 8 Key parameters such as the inlet, outlet, and leakage flow rates, flow constants, end stops stiffness, and preload of lifting force spring are identified and characterized. vii. Building a characterized Simulink model The parameters identified above are used to build a characterized Simulink model which represents the ATF test-rig, and its performance is evaluated and compared with the experimental results. Given supply pressure and actuator lift, this model will predict needle lift and chamber pressure. 1.5 Thesis Structure This thesis consists of six chapters and appendices. i) Chapter 1 presents background on diesel and NG engines. It also outlines the research objectives and scope. ii) Chapter 2 provides a literature review and it discusses prior work on fuel injectors, the working principle of fuel injectors and different types of actuators. iii) Chapter 3 provides the design of ATF test-rig, analysis of the design, apparatus, and safety procedures. iv) Chapter 4 outlines the ATF modelling and experimental characterization. v) Chapter 5 provides the conclusions and recommendation for future work. 9 Chapter 2: Literature Review The ATF test-rig developed in this thesis aims to simulate a typical hydraulically actuated fuel injector. This chapter describes the prior work on fuel injectors, working principles of a typical fuel injector, actuators, and the mathematical modelling of fuel injectors. 2.1 Natural Gas Fuel Injection Systems The advantage of having a dual fuel system is replacement of conventional fuel such as diesel with a much more available fuel like natural gas, often a smaller chain hydrocarbon, without sacrificing power, efficiency, and reducing emissions (Douville et al., 1998; Harrington et at., 2002; Duggal et at., 2004). In a typical direct-injection diesel engine with dual fuel capabilities, a pilot direct injection of diesel is required to assist the ignition of the NG fuel and air mixture, since it requires a temperature over 1200K (Beck et al., 2003; Aggarwal and Assanis, 1998). The downside of this method is the volumetric efficiency of the engine decreases since NG displaces charged combustion air, leading to the emission of unburned fuel and air mixture to the atmosphere. Injecting natural gas at high-pressure directly into the combustion cylinder of the engine solves this issue. Dual fuel natural gas-powered engines provide performance, reliability, and efficiency similar to diesel-powered engines with lower NOx and PM emissions (Duggal et al., 2004) for both two-stroke (Hodgins et al., 1996) and four-stroke engines (Dumitrescu et al., 2000). In dual fuel CI engines, injection takes place near the end of the compression stroke. A pilot injection of diesel takes place before injecting NG to raise the temperature of the cylinder so that the NG auto-ignites. 10 Miyake et al., (1983) tested the dual fuel capabilities of NG in a diesel engine for the first time. The study showed that the engine was 100% operational using 5% diesel-to-natural gas ratio having the same efficiency. This was further confirmed by Wakenell et al., (1987) in his study. Douville et al., (1998) found that NG dual fuel injection systems reduce NOx emissions by 20-40% and CO emissions by 15-20%. Duggal et al., (2004) examined Westport HPDI injector which had the same efficiency and power as a diesel injector, but it also had lower emissions. 2.2 Fuel Injectors The first fuel injector was developed by Herbert Akroyd Stuart using a jerk pump to meter out fuel to the injector at high-pressure, which was then integrated into a hot bulb engine in 1890 (Lutz, 1990). Later, Bosch and Clessie Cummins modified and improved the design for diesel engines. Scoltock (2011) reported that, in the mid-1920s, fuel injectors were used commercially in diesel engines. In 1925, Jonas Hesselman invented the Hesselman Engine, in which he implemented gasoline direct-injection. The injection system provides the correct amount of fuel to the combustion chamber at the right time. A fuel injector is shown in Figure 2.1. 11 Figure 2.1: Schematic of Port Fuel Injector. Reproduced from “Experimental Study of Injection Characteristics of a Multi-hole port injector on various Fuel Injection pressures and Temperatures,” by E. Movahednejad, F. Ommi, & K. Nekofar, 2013, EPJ Web of Conferences, 45(01116), 2. Copyright 2013 by E. Movahednejad, F. Ommi, & K. Nekofar. Reprinted courtesy of the Copyright Holder under a Creative Commons Attribution License 2.0. A fuel injector is usually a spring-loaded needle valve. The actuator controls the opening and closing position of the needle valve. When high-pressure fuel enters the injector, it moves down to the chamber around the needle valve. The spring in the injector body keeps the needle valve closed. A compression nut is used to set the spring pressure. When the fuel pressure exerts a sufficient force to overcome the spring compression, the needle valve opens. The fuel is sprayed into the combustion chamber. Once the high-pressure fuel supply is disconnected by the injector pump, the needle valve closes quickly under the spring compression force (machinery_spaces.com, n.d). 12 2.3 Actuators The actuation of the fuel injector can be achieved with a camshaft, with electromagnetic actuators or with piezoelectric actuators. Actuation by Camshaft The injection can be achieved mechanically by the location of cams on the camshaft. Figure 2.2 is a typical example of a cam plunger arrangement in a camshaft actuator. As the camshaft rotates, the follower moves up and down. The followers are connected to the fuel injector by a push rod and rocker arm. These linkages transfer the rotating motion of the camshaft to the reciprocating motion of the fuel injector, opening and closing them whenever needed. When the valve opens, it compresses the spring. This energy is later utilised to close the valve when the camshaft lobe rotates out from under the follower (machineries_spaces.com, n.d). Figure 2.2: Cam Plunger Arrangement. Reproduced from “Dual fuel injector modeling by finite difference method,” by C. Lim, 2000. Copyright 2000 by Clement Lim. Reprinted with permission. 13 Solenoid Actuator Traditional fuel injectors, as seen in Figure 2.3, use electromagnetic actuators like a solenoid. Solenoid actuators convert electrical energy into mechanical energy. According to Cvetkovic, Cosic, and Subic (2008), the factors which influence the operation of the solenoid are the material properties, geometric shape, air gap, and damping coefficients. Figure 2.3: Traditional Hydraulically Actuated Fuel Injector. Reproduced from “Development of a Sensorless Estimation Algorithm of the Injection Timing and Rate for a HSDI Common-Rail Injector,” by K. Nam, M. Yoon, & S. Park, 2004, JSME International Journal Series C Mechanical Systems, Machine Elements and Manufacturing, 47(3), 884. Copyright 2004 by JSME. Reprinted with Permission. 14 Working Principle of the Solenoid Actuator Cvetkovic, Cosic, and Subic (2008) further described the working principle of the solenoid actuator. Solenoid actuation is based on energized and de-energized operations. When the solenoid is energized, current flows through the wire, and a magnetic field is produced in the wire and through the air gap between the plunger and the coil. Therefore, the plunger is pulled against the return spring due to the attraction force produced in the air gap. When the solenoid is de-energized, the attraction force is removed as the magnetic field reduces, and the return spring retracts the plunger to its rest position. Piezoelectric Actuator Fuel injectors can also be actuated using piezoelectric actuators. Over the years, piezoelectric actuators have proven to be more efficient due to their faster response time, accuracy, high efficiency, ability to general larger internal forces, and enhanced combustion through multi-pulse closely spaced injections (Satkoski et al., 2011). Precise injection intervals are possible as the valve is actuated quickly, thereby, reducing emissions. Senousy et al., (2009) and Scharf (2006) discovered that the direct fuel injection system reduces fuel consumption by 15% compared to electromagnetic injection systems. In piezoelectric actuators, the needle stroke can be altered by varying the applied voltage whereas, in the case of conventional electromagnetic actuation, the needle stroke is fixed. Studies by Oh et al., (2007) showed piezoelectric actuation that uses PWM to reduce the dwell time between pulses to half of the one by conventional electromagnetic actuators. Figure 2.4 shows a typical piezoelectric fuel injector. 15 Figure 2.4: Piezoelectric Fuel Injector. Reproduced from “Dynamic Modeling of a Piezoelectric Actuated Fuel Injector,” by C. A. Satkoski, G. M. Shaver, R. More, P. Meckl, D. Memering, S. Venkataraman, J. Syed, & J. Carmona-Valdes, 2011, Journal of dyanamic systems, measurement, and control, 133(5), 051011-2. Copyright 2011 by ASME. Reprinted with Permission. Working Principle of the Piezoelectric Actuator Satkoski et al., (2011) described the working principle of the piezoelectric actuator such that, when the stack of piezo ceramic expands, the bottom link moves down to the needle lower volume, which raises the pressure, leading to the opening of the needle valve. Therefore, high-pressure fluid leaves the needle chamber. The needle lower volume pushes the bottom link upwards as the piezo stack retracts. As a result, the needle lower volume pressure reduces, thus closing the needle valve. 16 Even though piezoelectric actuators have many advantages over solenoid actuators, the technology used in the solenoid is well known in the industry as reliable, cost-effective, and physically smaller than piezo units. Piezoelectric actuators, on the other hand, are much more expensive than solenoid units. In both cases, the principle of injection system remains the same. The Hydraulically Actuated Diesel Fuel Injector Figure 2.5 shows a schematic diagram of a diesel fuel injector. The fuel injection is controlled by a solenoid actuator. Below is a brief description of the components of a fuel injector, followed by a description of the injection process. Needle: The needle controls the flow of diesel fuel into the combustion chamber of the engine. It has an open and closed position. It gets pushed towards the closed position by a preload spring and a plunger. It gets pushed towards the open position by fuel and combustion forces. Preload Spring: The spring pushes the needle into the closed position. Plunger: The plunger mimics a hydraulic piston whose position is controlled by the pressure in the control chamber. When the pressure in the control chamber rises, it presses the plunger against the needle into the needle’s closed position. 17 Control Chamber: This chamber holds the plunger. The pressure in the control chamber controls the motion of the plunger and the needle. The control chamber has an inlet and an outlet orifice for diesel which is used as the hydraulic fluid. Inlet Orifice: Diesel enters the control chamber through the inlet orifice. It prevents rapid pressure equalization between the high-pressure diesel supply inlet and the control chamber. Outlet Orifice/Valve: Diesel fuel exits the control chamber through the outlet orifice. Flow through the orifice is controlled by a hydraulic ball valve. When it is closed, pressure builds up in the control chamber and the plunger then pushes the needle into the closed position. When the valve opens, the pressure in the control chamber drops and the fuel pressure pushes the needle into the open position. Solenoid Actuator. The hydraulic ball valve is controlled by a solenoid actuator that can rapidly open and close the hydraulic valve. Diesel Supply Line: Diesel is delivered to the injector through a high-pressure hydraulic supply line that is attached to the injector through hydraulic fittings. Diesel Return Line: Diesel returning from the injector’s outlet orifice is returned to the diesel supply system through a return line that is connected to the actuator through hydraulic fittings. 18 Figure 2.5: Traditional Hydraulically Actuated Fuel Injector. Reproduced from “Development of a Sensorless Estimation Algorithm of the Injection Timing and Rate for a HSDI Common-Rail Injector,” by K. Nam, M. Yoon, & S. Park, 2004, JSME International Journal Series C Mechanical Systems, Machine Elements and Manufacturing, 47(3), 884. Copyright 2004 by JSME. Reprinted with Permission. The injection process can be summarized as follows: In the needle closed position, the solenoid valve is closed. High-pressure diesel is fed into the injector through the fixed diameter inlet orifice and the supply pressure equalizes with the control chamber pressure, which presses the plunger and needle against the combustion force into the closed needle position. When the solenoid actuator is energized, it opens the outlet orifice valve and a small amount of fuel is released through the outlet orifice to the return line. As a result, the control chamber pressure drops, and the fuel pressure from plenum chamber and combustion pressure from sac chamber pushes the needle into the open position. Consequently, fuel is injected Supply Line - Diesel In Return Line - Diesel Out Needle Inlet Orifice Solenoid Valve And outlet orifice Preload Spring Plunger Closed End Stop Open End Stop Plenum Chamber Sac Chamber 19 into the combustion chamber. When the solenoid is de-energized, the solenoid valve closes, and the pressure in the control chamber increases back up to the supply pressure, which drives the control plunger and needle back to the closed-end stop position. The goal of this project is to develop a good model which describes the relationship between the actuator needle and injector needle displacement. This model requires an ATF for model characterization and validation. 2.4 Experimental Characterization of Fuel Injectors This project aims to develop a model to link actuator motion to needle motion. This model is derived and validated by experiments. The current section briefly reviews previous experimental campaigns aimed at characterizing fuel injectors. Inokoshi (2007) and Brown (2008) conducted experiments on flow benches and monitored the flow parameters using Coriolis mass flow meters to measure the flow rate of NG. Gravimetric analysis was used to measure the flow rate of liquid fuel. Viscor was used instead of diesel to reduce the flammability hazard. Co-injection of NG and viscor was used in the HPDI injector system. A typical HPDI injector was modified, such that viscor was injected directly into the gas reservoir instead of the combustion chamber. Viscor flow rate was calculated by taking the average of the change in mass of the supply quantity over 10 – 20mins. The same analysis is used to measure the flow rate of Viscor in this study. 20 Birger (2010) also focused on characterizing the flow rates and conducted similar experiments on a Westport HPDI Fuel Injector connected to an engine inside a visualization chamber. Birger used a Coriolis mass flow meter to measure the flow rate along with pressure sensors to measure the control chamber pressure. Birger also calculated the mass of fuel injected by measuring the mass of the supply tank. This paper focused on characterizing the flow rates only to determine the amount of fuel injected over a certain duration, not the relation between the actuator needle lift and the injector needle lift. It did not emphasize the change in control chamber pressure and how its dynamics affected the injector needle motion. The main motivation in designing the ATF in this thesis is to find the correlation between the actuator needle lift and the ATF needle lift based on the flow rates. Lim (2000) designed an injector test rig by modifying the existing conventional fuel injector and recorded the chamber pressure. The test rig simulated a Detroit 6V92 engine. Lim used a measurement rig made of strain gauges to measure the injection timing. This paper also focuses on the amount of fuel injected and correlates it with the pressure drop in the control chamber. The ATF designed in this thesis is also modified like the injector modified by Lim, to measure the chamber pressure. The relation between the actuator lift and the injector needle response is not compared by Lim, which is discussed in this thesis. Nam et al., (2004) conducted the experiments on a test bench and used a common-rail fuel injection system to measure the voltage, current, and injection rate using a commercial injection rate meter. The mathematical modelling of the ATF designed in this thesis is motivated by Nam et., since they 21 are very similar. Apart from that, Nam’s paper emphasizes on estimating the injection timing and rate with a model. Chamber pressure is not recorded along with the injector needle lifts. 2.5 Mathematical Modelling of Fuel Injectors Simulation tools like GOLD (Kheirkhah, 2015) and commercial simulation tools like AMESim (Birger, 2010) are commonly used to model fuel injectors. This thesis requires a simple analytical model to describe the relationship between needle lift and measured actuator lift, in order to develop a real-time control system. Therefore, large commercial simulations tools cannot be used, but a simple lumped parameter model is required. Thus, for validation and calibration purpose, this thesis implements the model in MATLAB/Simulink and compares it to the experimental results obtained from the ATF Modelling of fuel injector has been around for many years. In the early days, modelling was performed manually by using tables and simplifying assumptions. Burman and DeLuca (1962) for the first-time modelled fuel injectors using large equations and tedious calculations. They were able to characterize the injection pressure along with the compressibility and dynamics of the fluid. Computer simulation was introduced in the 1970s. Scullen and Hames (1978) designed a simple model of a fuel injector due to lack of computing power. The model assumed uniform pressure throughout the injector to simplify the model. Therefore, only one differential equation was used to model the pressure. 22 Kegl et al., (1992) and Marcic (1993) used a similar process to model a fuel injector. Kegl et al., and Marcic derived differential equation using equations of motion and modelled the pressure, injection rate, and needle lift in the injector. Experiments were also conducted to compare and validate the modelled values of pressure and lift with the experimental results. Lim (2000) and Nam et al., (2004) characterized an HPDI fuel injector. The models are related to the work in this thesis and reviewed in more detail below. Lim (2000) suggested that the dynamics of the fluid pressure inside the injector is a function of the geometry and movement of the components in the injector. Figure 2.6a illustrates the cam actuated diesel injector assembly. The inlet orifice and the passages inside the injector control the mass flow rate. The geometry, shape of the plunger, and needle control the pressure drop for injection. Lim used the free body diagram in Figure 2.6b to model a typical fuel injector used for co-injection. This thesis is focused on diesel injection. Therefore, the focus is on modelling of the diesel injector needle only. 23 Figure 2.6: a) Diesel Injector Assembly. b) Diesel Needle Free Body Diagram. Reproduced from “Dual fuel injector modeling by finite difference method,” by C. Lim, 2000. Copyright 2000 by Clement Lim. Reprinted with permission. In this case, the needle acceleration is controlled by the spring force, friction, and the pressure forces from the spring cavity, the primary needle volume, and the SAC volume. The needle opens when the total upward force exceeds the needle spring preload force, injecting fluid into the combustion cylinder. In a typical injector, the preload force is 175N (Lim, 2000). A similar force balance also applies to the ATF build for this thesis. The hydraulic model of the ATF is similar to the one presented by Nam et al., (2004) for a High-Speed Direct Injection (HSDI) system. It is used in the following schematic diagram of the control chamber in Figure 2.7b to model the volumetric flow rates. b) a) 24 Figure 2.7: a) Fuel Injector. b) Schematic diagram of the control chamber. Reproduced from “Development of a Sensorless Estimation Algorithm of the Injection Timing and Rate for a HSDI Common-Rail Injector,” by K. Nam, M. Yoon, & S. Park, 2004, JSME International Journal Series C Mechanical Systems, Machine Elements and Manufacturing, 47(3), 884-885. Copyright 2004 by JSME. Reprinted with Permission. The pressure in the control chamber is responsible for pushing the needle down into its closed position against the other forces in the injector. When the injector is pressurized by closing the outlet orifice, the chamber pressure and the inlet rail pressure equalizes, and the needle is pressed into the closed position. When the solenoid is actuated, the outlet orifice opens and the pressure in the control chamber falls rapidly. This removes the downward force on the needle and the needle is pushed into the open position. The following equations are used to model the flow rates and the equation of motion of the needle (Nam et al., 2004). Equation 2.1 is the continuity equation which states that the flow rate due to fluid compressibility in the control chamber, ̇, equals the differential fluid flow between the inlet and outlet orifice,  − , minus the time derivative of chamber volume, ̇ Inlet and outlet flow rates are modelled as turbulent flows in equations 2.2, a) b) Hollow Chamber Plenum Chamber 25 and 2.3. The Equation of motion of the plunger and needle is defined in Equation 2.4. Detailed Nomenclature of the variables applied in these equations are provided in the list of symbols at the beginning of the thesis. The important point for this thesis is that Equations 2.1-2.4 are very similar in nature to the ones used in the ATF. ̇ =  −  − ̇ (2.1)  = ..2| − | (2.2)  = ..2| − | (2.3) Where  =  −  and ̇ = −̇  + ̈ = −  − −  −  − ̇ + ( − )−  +  cos (2.4) 26 Chapter 3: ATF Design, Analysis and Apparatus In this chapter, the design of the ATF test-rig, analysis of the design, apparatus, and safety procedures are discussed. The experimental procedure is highlighted in Appendix B. As mentioned in chapter 2, the ATF designed and modelled in this thesis is motivated by a typical hydraulically actuated dual fuel injector. The ATF imitates its injection process but without injecting any fuel, since one of the primary objectives of this thesis, is to find the correlation between the actuator displacement and the ATF needle lift. Figure 3.1a shows the exploded view of the ATF with all the components labelled, that are designed in this section. Figure 3.1b shows the fabricated ATF prototype. Figure 3.1: a) Exploded View of the ATF Design. b) Fabricated Prototype. ATF Body Lifting Force Spring End Stop Lock Nut Compression Nut Brass Sleeve Plunger Needle Closed End Stop Pressure Sensor Hydraulic Fittings a) b) 27 The design parameters of the ATF are: i) Supply and chamber pressures, ii) Inlet and outlet orifice sizes, iii) Mass of the plunger and needle, iv) Needle lift, v) Flow rates, vi) Control chamber volume, vii) Preload of lifting force spring, Figure 3.2a shows the cross-section of the ATF designed and compares it with the cross-section of a typical dual fuel injector in Figure 3.2b. The key parameters such as the control chamber volume, mass of the needle and plunger, inlet and outlet orifices, are similar to typical dual fuel injectors so that the ATF’s needle dynamics would be realistic. 28 Figure 3.2: a) Cross-section of the ATF. b) Cross-section of Dual Fuel Injector. Reproduced from “Development of a Sensorless Estimation Algorithm of the Injection Timing and Rate for a HSDI Common-Rail Injector,” by K. Nam, M. Yoon, & S. Park, 2004, JSME International Journal Series C Mechanical Systems, Machine Elements and Manufacturing, 47(3), 884. Copyright 2004 by JSME. Reprinted with Permission. Figure 3.2 also illustrates the end stops. The plunger and needle are placed between two end stops. The closed position is down, and the open position is up. The lifting force spring pushes the needle open, since it emulates the NG fuel force, the combustion force, and the preload force. In the needle closed position, the outlet orifice is closed by the actuator and the inlet pressure equalizes with the control chamber pressure, pushing the needle into the closed end stop. When the actuator opens, the control chamber pressure drops, and the needle is pushed into the open position by the Open End Stop Closed End Stop Inlet Orifice Control Chamber Plunger Needle a) b) Outlet Orifice Lifting Force Spring Preload Spring 29 lifting force spring. When the actuator closes, the pressure in the control chamber builds up again and pushes the needle into the closed position. 3.1 Mechanical Design of ATF The ATF consists of many components like the ATF body, plunger, needle, end stops, compression nut, and custom hydraulic fittings. Figure 3.3 contains all the components designed by the author of this thesis and machined by the in-house UBCO Machine Shop. Figure 3.1a shows the exploded view of all the components designed for the ATF. All these components are assembled together and shown in Figure 3.1b which is the fabricated prototype. All the mechanical drawings for the designed parts are shown in Appendix A. Figure 3.3: Components of the ATF. Lifting Force Spring 30 ATF Body The injector body is a key component since it holds all the other components that make up the fuel injection system. The body contains all the necessary ports for the fuel supply and return, as well as a housing for the actuator. Similarly, the ATF body designed in this section also has the same necessary ports. But in a conventional fuel injector, there is no access to the control chamber. Therefore, the design is modified in such a way that the chamber pressure is monitored using a piezoelectric pressure sensor. The ATF body is made with steel 4140. Figure 3.4 shows the cross-section diagram of the ATF body. The chamber volume, pressure sensor mount, inlet orifice, and outlet orifice are labelled in the figure. The chamber volume, inlet orifice, outlet orifice, and the match fit parameters, which are the tolerances between the brass sleeve and ATF body, are kept similar to the conventional fuel injector since the needle displacement is depended on these parameters. The control chamber sees a pressure of 27.6. The ATF body has an outer radius of 27. Steel 4140 has a yield strength of 415 and a tensile strength of 655. Therefore, the ATF body must be 1.14 thick to withstand the high pressure with a safety factor of 13.6. The threads on the ATF body has a thread shear area of 4302.9, with a shear strength of 390, and it can withstand upto 1.6 of force. The force applied on the threads by the end stop is approximately 2. Therefore, the threads can also handle the high pressure. The calculations are provided in Appendix C. 31 Figure 3.4: Cross section of the ATF. Solenoid Actuator An electromechanical solenoid actuator is used with the ATF. Westport provided a prototype hydraulic actuator and an actuator block shown in Figure 3.5. Usually, there are two solenoid actuators in the block portrayed in Figure 3.5. Since only the actuator for controlling NG fuel is used, the actuator for controlling the Diesel pilot fuel is replaced with a plug. The actuator block sits on top of the ATF, and a seal is achieved using an O-ring and three M5 bolts. To achieve a better seal, a copper gasket and a clamp are used which is discussed later in Section 3.2.1. Pressure Sensor Housing Return Line Inlet Orifice Control Chamber Brass Sleeve Housing for Match Fit and Plunger End Stop Slot Retaining Ring Slot To outlet orifice 32 Figure 3.5: Solenoid Actuator. Brass Sleeve Figure 3.6 shows a housing for a brass sleeve. The sleeve is there, so that when the plunger is reciprocating through the sleeve, the friction between them is minimum, providing a good match fit, the tolerance is of the order of few microns. The brass sleeve guides the plunger made of steel in and out without binding. There are two O-rings around the plunger to prevent any leakage, as shown in Figure 3.6. Also, another O-ring is added on the top of the sleeve to achieve a face seal. The brass sleeve is held in place using a retaining ring. Even though the retaining ring keeps the brass sleeve from moving up and down, it is not guaranteed that it keeps the brass sleeve fixed, since there is a small breathing space for the ring. This needs to be addressed in the future version of the ATF. 33 Figure 3.6: Brass Sleeve. Plunger and Needle The geometry of the plunger and the needle have also been modified. Both are made separately using steel 4140. The plunger has a disc at the bottom to accommodate the lifting force spring, which can be seen in Figure 3.2a in the exploded view, and has threads at the center for the needle to screw inside. The plunger mimics a relatively standard hydraulic cylinder. Figure 3.7a and b show the geometry of the plunger and needle. Figure 3.7c shows the fabricated assembly of the plunger and needle. It can be observed that there is an O-ring attached to the plunger in Figure 3.7c. This is to prevent any sort of leakage through the match fit between the plunger and the brass sleeve. It is important to keep the mass of the plunger and needle same as the conventional fuel injector since it affects the dynamics of the needle. 34 Figure 3.7: a) Plunger. b) Needle. c) Assembly of Plunger and Needle, Figure 3.8 shows the match fit between the plunger and brass sleeve. The tolerance is very tight to avoid any leakage coming through it. The O-ring around the plunger also reduces the leakage. Figure 3.8: Match Fit between Plunger and Brass Sleeve. a) b) c) 35 End Stop There are two end stops in this design shown in Figure 3.2a. One is the open end stop which is the injector body housing, and the other one is the closed end stop show in Figure 3.9. When the actuator triggers, the needle moves upwards - which is the open position - and when the actuator is demagnetized, the needle moves downwards - which is the closed position. The open end stop is made of 4140 steel. The stiffness of the end stop should be such that, it deflects only 1% to 2% of the overall lift under load. This is not taken into account in this design, this issue needs to be addressed in the future version of the ATF. Figure 3.9: Closed-End Stop. The closed-end stop is held in place using two custom-made nuts, where one nut is used to lock the closed-end stop, and the other one is used as a jam nut, where it is jammed up against the other nut to lock the two in place. Figure 3.10 shows the application of lock nut, and how it limits the closed-end stop. 36 Figure 3.10: Application of lock nut. Compression Nut The compression nut holds the lifting force spring which is held against the disc attached to the plunger, as shown in Figure 3.1a and Figure 3.2a. The rotation in the nut determines how much preload of lifting force spring is applied. Figure 3.10 shows how the nut is screwed into the injector body and Figure 3.11 shows the design. It is made using steel 4140. The shear area of the thread is 2430.9 with a shear strength of 390. Force required to strip the threads would be 947. The maximum force the nut would see is 2. Therefore, it will not strip. The calculations are provided in Appendix C. 37 Figure 3.11: Compression Nut. Custom Hydraulic Fitting Custom hydraulic fittings are made for O-Ring Boss fitting for the inlet fuel supply and for the return line as showed in Figure 3.12a. Another fitting is also custom made for the piezoelectric pressure sensor which measures the supply pressure, as it can be seen in Figure 3.12b. Figure 3.12: Custom Hydraulic Fittings. a) Inlet and Return Line Fitting. b) Pressure Sensor Fitting. Actuator Preload Screw One of the main objectives of this thesis is to find the relationship between the actuator needle displacement and the ATF needle displacement. Figure 3.13 shows the schematic of the actuator. a) b) 38 Figure 3.14a shows she actuator preload screw and Figure 3.14b shows the assembly of the screw. The actuator is closed with the help of a preload screw and a solenoid preload spring as shown in Figure 3.13. Figure 3.14a shows the top of the solenoid actuator, where the preload screw is inserted. The actuator plunger is held in place by the ball seat valve. When the system is pressurized, the actuator preload spring prevents the actuator from opening. The screw determines the preload force. However, there’s no access to the actuator plunger. Thus, the actuator displacement cannot be measured. Therefore, the ball seat valve is modified in such a way that the actuator plunger can be extended, so that its lift can be measured using a laser sensor. The actuator preload screw is also modified, such that the needle can go through it and can be extended. Figure 3.15 shows the modified preload screw with a hole in the center, through which the needle comes out. Figure 3.13: Schematic of the Actuator  Solenoid Ball Valve Solenoid Preload Spring stiffness and damping Solenoid Actuator Preload Screw Actuator Plunger Added needle for measuring plunger position 39 Figure 3.14: Preload Actuator Screw. a) Solenoid Actuator. b) Preload Actuator Screw Assembly. Figure 3.15 shows the modified ball seat valve where a magnesium rod is screwed inside it. And it also shows the modified actuator preload screw. The preload spring is recycled from the real actuator. Figure 3.15: Modified Actuator Preload Screw. 3.2 Design Analysis During designing certain components like clamping mechanism, clamping screw, and lifting force spring stiffness, it is very important to check if these components are compatible with the above design and can withstand pressures up to 25MPa. Therefore, certain calculations are done to check if these components deform. FEM analysis of components like plunger and end stops are also done in chapter 4. a) b) Preload Screw Preload Spring Ball Seat valve 40 ATF and Actuator Clamping In a typical injector, an actuator block is attached to the injector using three M5 bolts in addition to an external clamp. A good seal is also achieved using an O-ring. Figure 3.16a shows the assembly of the injector and the ATF. Figure 3.16b portrays the O-ring in a slot on the ATF which seals the actuator to the ATF. The return line and the Viscor line of both the ATF and the actuator line up together. Figure 3.16: a) Actuator and ATF Assembly. b) O-ring on top of the ATF. Figure 3.17a shows the schematic of the ATF and the actuator with a modified preload screw in closed position. The Actuator sits on the top of the ATF and is sealed using O-ring and bolts. When the actuator is closed, the preload screw pushes the ball valve downwards, keeping it closed. When the solenoid actuates, the actuator plunger lifts along with the needle, and fuel leaves from the control chamber to the return line as shown in Figure 3.17b. The actuator needle lift is then recorded. a) b) Return Line Viscor Line 41 Figure 3.17: a) Schematic of the ATF and Actuator in Closed Position. b) Schematic of the Actuator in Open Position. The initial ATF design attempted to use the provided 3 M5 bolts in the actuator block to create a steel on steel seal between the Actuator block and the ATF. This did not provide a reliable seal. Therefore, the copper gasket shown in Figure 3.18 was designed and placed in an engraved pocket on top of the ATF. The copper gasket is also used to block the unused NG line for the second actuator. Figure 3.18: a) Copper Gasket attached to the ATF. b) Copper Gasket Design. a) b) a) b) 42 The copper gasket is designed to provide sealing through plastic deformation. The Area of the copper gasket = 318.27 , and Yield strength of copper = 80. So, the force required to get a good seal 25,461.6 . Therefore, each of the three M5 bolt needs to provide a force of 8487.2. However, with a tensile yield strength of 450 and a tensile stress area 14.18 for M5 bolts, it can withstand 6,381 of force. A safety factor of less than 0.7 indicates that the three M5 bolts are too weak to achieve plastic deformation of the copper seal. This calculation is shown in Appendix C. To increase the clamping force, the clamp between the actuator block and the ATF shown in Figure 3.19a was designed. It consists of two parts, the bottom part is threaded and screwed on the ATF body, whereas the top part sits on the actuator block. Both the parts are clamped together using two Grade 5, 5/16-inch bolts. Figure 3.19b shows the fabricated prototype of the clamp. Figure 3.19: ATF and Actuator Clamp. a) Clamp Design. b) Clamp Prototype. a) b) 43 The Tensile strength of Grade 5, 5/16-inch bolt = 827.4, the yield strength = 634.3, and tensile stress area = 56.6. Thus, the two bolts provide a combined tensile yield load of 46 and an ultimate tensile load 60. This is sufficient to provide the required plastic deformation of copper seals with a safety factor larger than 2. This calculation is shown in Appendix C. 3.3 Apparatus This section discusses all the components used - both electrical and mechanical, along with their operation. Figure 3.20 portrays the schematic of the test setup. The hydraulic pump is driven by a compressor. The increase in pressure can be observed in the pressure gauge. The inline filter cleans the fuel coming from the pump reservoir. There are two relief valves, one before and one after the filter. Both the relief valve set the pressure at 27.6MPa. One of the valves is there for safety. A 2-way ball valve controls the flow of the fuel. A piezoelectric sensor measures the supply pressure and sends it to the controller. The controller generates a desired current pulse profile for the actuator. This current profile is amplified by a current amplifier and sent to the actuator. When the Actuator is triggered, the laser sensors record the actuator needle displacement and the ATF needle displacement and sends the data to the controller. At the same time, the piezoelectric pressure sensor also sends the pressure measurement of the control chamber to the controller. 44 Figure 3.20: Schematic of the Setup ATF Test Setup The test setup shown in Figure 3.21 is made from wooden blocks. The combination of the wooden base and the vertical board has a dimension of 100cm x 25cm x 5cm. The vertical board is also reinforced with of 5cm x 5cm wooden studs at the back to make the frame rigid. Polytech HSV-2002 laser sensor 1 which is mounted at the top measures the actuator needle displacement. The Polytech laser sensor 2 is mounted on a wooden stud on the left, since there is no space on the board, measures the ATF needle displacement. The ATF, sensors, and all the hydraulic components are mounted on the board. 45 Figure 3.21: ATF Test Setup. Hydraulic Components Figure 3.22 illustrates all the hydraulic fittings such as pressure gauge, 2-way high-pressure ball valve, hydraulic hose, and ¼ inch NPT fittings. All the components are rated at 34.5MPa. The working pressure for this test setup is 27.6MPa. All the components are connected using Teflon tapes to seal the pipe threads and avoid any leakage. Standard pipe clamps are used to attach all the hydraulic fittings to the board. Polytec Laser Sensor 1 ATF Pressure Sensor Return Line Relief Valve Filter Actuator Vertical Board Wooden Base Polytec Laser Sensor 2 46 Figure 3.22: Hydraulic Fittings. Viscor Hydraulic Fluid The hydraulic pressure is generated by an air driven Power Fist hydraulic pump that can provide up to 68.9 pressure, with a flow rate of 2.34/ℎ, and an oil capacity of 2.3. The pump is driven by a 2 , 5.6 , Power Fist air compressor with max air pressure of 0.79, air flow of 3.6 at 0.62, and 5 at 0.28. The pump comes with hydraulic fluid inside. Initial tests were performed using the hydraulic fluid to check the functionality of the test rig. Then the hydraulic fluid was replaced with Viscor Calibration fluid which has the same density and viscosity as diesel but is not flammable. The pump has a return port for any excess fuel in the hydraulic line to return to the tank. Figure 3.23 shows the air compressor and the hydraulic pump used in this experiment. 47 Figure 3.23: a) Air Compressor. b) Hydraulic Pump Relief Valve The pump supplies Viscor at 68.9MPa. Therefore, a Hy-Lok RV2 series relief valve is used to regulate the pressure at 27.6MPa. Its maximum working pressure is 34.4MPa, and cracking pressure range is 27.6 to 34.4MPa. The cracking pressure is adjusted using a nut at the top which is spring loaded. Figure 3.24 shows the relief valve attached to the test setup. Figure 3.24: High-Pressure Relief Valve. 48 In-line Filter Figure 3.25 shows an Enerpac FL 2101, 10-micron high-pressure in-line filter to filter out the Viscor supplied to the injector. Since the pump uses a piston to create a reciprocating motion along an axis, it creates ripples in the Viscor. The filter is also believed to slightly dampen those ripples. Figure 3.25: In-line Filter. 3.4 Measurements There are four sensors used in this test setup to measure the parameters. Two high-speed laser sensors measure the actuator and ATF needle displacement. Two piezoelectric pressure sensor measures the supply and control chamber pressure. High-Speed Laser Sensor A Polytec HSV-2002 high-speed vibrometer is used to measure the displacement of the actuator needle and the ATF needle. The laser vibrometer provides a high-resolution signal output with low noise for precise results. The laser vibrometer has two output channels - one for displacement and the other for velocity. It has a frequency bandwidth of 250, resolution of 0.32µ and 49 15µ/, and sensitivity of 0.32/ and 5//. For high accuracy, the ATF sensor head is connected to the displacement channel, and the actuator sensor head is connected to the velocity channel. The actuator velocity is integrated into Matlab to determine the displacement of the actuator needle. Figure 3.26a shows a laser beam steering rail and a sensor head, and Figure 3.26b shows the laser signal conditioning unit. Figure 3.26: HSV 2002 High-Speed Vibrometer. a) Laser beam steering rail and Sensor Head. b) Signal conditioning unit Piezoelectric Pressure Sensor PCB Piezotronics 113B22, high-frequency ICP pressure sensors are used to measure the chamber pressure and supply pressure. They can measure maximum pressures up to 15 with a resolution of 20. The pressure sensors are calibrated to 1/. Figure 3.27 shows the piezoelectric pressure sensor and its controller. a) b) 50 Figure 3.27: a) Piezoelectric Pressure Sensor. b) Controller 3.5 Control System and Data Acquisition A schematic diagram of the Hardware in the loop (HIL) system controlling the ATF test-rig is shown in Figure 3.29. A pulse of 20 A current, shown in figure 3.28, is generated using a dSpace 1104 HIL system connected to a Maccon SWM048 current amplifier. This pulse is fed to the actuator that opens the outlet orifice valve, which in turn triggers the needle to lift. Pulse parameters are controlled by dSpace software interface, which allows the user to input desired amplitude and pulse duration, to trigger the actuator. The laser sensors and the pressure sensors are also interfaced to the dSpace HIL system and the results are displayed on the software interface in a computer. Data is collected at 40 . Figure 3.28: 20A Current Generated Using Simulink a) b) 51 Figure 3.29: Schematic of the Electronics 3.6 Safety Review The experimental setup is based in EME 2212 laboratory of The University of British Columbia, Okanagan. The experiment requires high pressure. Even though all the hydraulic fittings are rated at 34.4MPa, but if any of the fittings are not tightened, the operator could get hurt. The air compressor and pump might also lead to failure if precautions are not taken. This section discusses all the precautionary measures taken to avoid any fatal accidents. 52 Air Compressor Operation of any equipment with loose bolts and/or fittings will lead to excessive vibration and the premature failure of the compressor system control components. Safeguard: The air compressor is equipped with a pressure relief valve that is designed to discharge tank pressure at the predetermined setting when a systems failure occurs. Hydraulic Pump A hydraulic pump failure can be caused by using low quality fluid and dust. Safeguard: High-quality oil is used along with high-pressure filter to remove any dust particles. Failure of Hydraulic Components. Catastrophic ruptures or even small leaks in the hydraulic circuit can lead to life-threatening projectiles or razor-sharp sprays of hydraulic fluid. Safeguard: i) An enclosure made of Lexan polycarbonate sheet is placed on the top of the test rig for safety. Figure 3.30 shows the safety enclosure, and it has a dimension of 110cm x 30cm x 40cm. ii) The hydraulic relief valve is designed to keep the pressure below rated values of all components. iii) The hydraulic pump will only engage when its foot pedal is actively pressed by the operator. In case of failure, the operator can immediately stop the pump by releasing the foot pedal. 53 iv) The control software monitors the absolute pressure in the hydraulic circuit. If an overpressure is detected the software will try to relieve pressure by opening the actuator fully. Figure 3.30: Safety Enclosure. 54 Chapter 4: Mathematical Modelling of Actuator Test Fixture (ATF) and Experimental Characterization Based on the mathematic modelling methods shown in Section 2.5 (Nam et al., 2004), the ATF presented in this thesis is modelled. The mathematical model of the ATF is divided into two subsystems: mechanical and hydraulics. The assumptions for modelling the injectors are as follows: i) Fuel is compressible, ii) Friction and hydraulic losses through orifices are neglected, iii) The supply pressure pulsation is neglected, iv) Temperature is constant, v) Ambient pressure is constant at 100, vi) Fluid inertia is neglected. Figure 4.1a shows the general ATF model which has two inputs: the measured actuator needle lift and measured supply pressure. ATF needle lift and the chamber pressure are the output from this model. The ATF model has been divided into two subsystems in Figure 4.1b, mechanical and hydraulic model. 55 Figure 4.1: Simulink Model. a) Overall ATF Model. b) The subsystem of the ATF. 4.1 Mechanical Modelling of the ATF Figure 4.2 shows the simplified model of the ATF. The solenoid actuator sits on top of the ATF. The lifting force spring simulates the preload force, the NG fuel pressure force and the combustion force in the cylinder. Viscor is supplied to the ATF using the inlet orifice which has a fixed diameter, and the solenoid valve controls the size of the outlet orifice area. a) b) 56 Figure 4.2: Simplified drawing of the ATF The equation of Motion of the Needle ̈ = −  −  ( − ) (4.1) Where, i)  () is the needle spring force. The spring force consists of the net force acting on the plunger and needle as discussed in Section 4.1.2, and it also incorporates the end stops which is shown in Figure 4.6.    =     =     =     =  k c      =  Open End Stop Closed End Stop Solenoid Ball Valve Lifting Force Spring Stiffness, k Lifting Force Spring Damping, c Mass of needle and plunger Control Chamber  = displacement of needle Solenoid Preload Spring stiffness and damping  Actuator Preload Screw Actuator Needle 57 ii)  () is the needle damping force, and it is determined by model matching. iii)  () is the mass of needle and plunger. Figure 4.3: a) ATF Needle in Closed Position b) ATF Needle in Open Position In the needle closed position as shown in Figure 4.3a, the solenoid valve is closed, the supply pressure equals the chamber pressure, which presses the needle against the lifting force spring into the closed end stop (lower needle position). When the solenoid actuates, the solenoid valve opens, the chamber pressure drops. The spring force then overcomes the force from the chamber pressure and drives the needle towards the open end stop as shown in Figure 4.3b. The process is reversed, when the solenoid valve closes and the pressure in the chamber increases back up to the supply pressure, driving the needle back towards the closed end stop. Figure 4.4 shows the mechanical subsystem of the ATF Model. The details of this model are discussed in the remaining subsections of section 4.1. a) b) Closed End Stop Open End Stop 58 Figure 4.4: Mechanical Subsystem Lifting Force Spring Stiffness The experiments are not conducted in an engine. Therefore, there is no combustion. The lifting force spring simulates the combustion force and the Natural Gas force. This is designed to provide an almost constant force throughout the lift of the ATF. The net force in a typical fuel injector is 1233.6 . The spring is designed in a way such that the preload of lifting force spring is approximately constant throughout the desired needle lift. The lifting force spring purchased has a stiffness of 432 /. For instance, in order to reach 1233.6  the spring must undergo 2.85 of deflection. And then to achieve a desired needle lift the spring deflects further 0.2 which provides 86.4 of force. The increase in preload is just 7% which shows that the preload remains almost constant throughout the needle lift. That is why the spring chosen is very stiff. The spring stiffness is tested using a load frame according to the graph in Figure 4.5. 59 Figure 4.5: Lifting Force Spring Stiffness. Therefore,  =  = 432∗2.9 = 1234 (4.1) The compression nut with a pitch of 0.705 / requires 4.05 turns to achieve the desired combustion and fuel force. However, since it is very difficult to determine the zero position of the spring in the setup, the actual force needs to be identified and matched in the Simulink model as described in Section 4.1.4. End Stop Modelling There are two end stops in the ATF as shown in Figure 4.2 and Figure 4.3. When the ATF is fully pressurized, the plunger disc rest against the closed end stop, and the needle is in closed position. When the actuator is energized, the plunger is pulled up, and the disc pushes against the open end stop, which is the ATF body. y = 432.31x - 137.8R² = 0.998902004006008001000120014000 0.5 1 1.5 2 2.5 3 3.5Force (N)Displacement (mm)Load (N)60 The end stops are modelled as springs and dampers as shown in Figure 4.6. When the control chamber is fully pressurized, the plunger disc hits the closed end stop, where the control chamber force is ,. The disc deforms up to −0.1 at a chamber pressure of 25. The spring stiffness and damping for the closed end stop are  and . When the solenoid actuates, the chamber pressure starts to decrease. When the chamber pressure is less than the preload of the lifting force spring , the needle starts to open. Ultimately, the plunger disc hits the open end stop, where the stiffness and damping are  and . When the solenoid closes, the plunger returns to the closed end stop. Figure 4.6: Stiffness and Damping Model. 61 The parameters used on these models are listed below: i) , is the initial force on the plunger when the control chamber is fully pressurized, ii)  is preload of the lifting force spring emulating combustion forces and NG fuel forces, iii)  is stiffness of the lifting force spring,  is the closed end stop stiffness, and is the open end stop stiffness, iv)  is the damping coefficient between the end stops,  is the closed end stop damping coefficient, and is the open end stop damping coefficient, v)  is the deflection of the closed end stop when the actuator is closed, and the needle/plunger is pressed into closed end stop. Closed End Stop The stiffness of the closed end stop is tested by removing the lifting force spring from the ATF and adjusting the setpoint block (closed end stop) such that there is zero ATF needle lift as shown in Figure 4.7. The supply pressure is slowly increased by turning the relief valve, and the pressure and deformation readings are recorded as shown in Figure 4.8. 62 Figure 4.7: Closed End Stop Model of the ATF. The top diagram of Figure 4.8a shows the ATF needle displacement (blue), starting form zero, against time. The bottom portion of Figure 4.8a shows the control chamber pressure (blue) and supply pressure (red). As pressure is slowly increased, the end stop starts deforming. The closed end stop deforms up to 0.09 at 25 pressure which corresponds to 2050, by multiplying the area of the plunger in contact with the fuel in the control chamber. Figure 4.8b shows the relationship between ATF needle lift and the control chamber force. Therefore, the observed end stop stiffness from Figure 4.8b is around .= 24,117.6/. Setpoint Block (Closed End Stop) Open End Stop Plunger Disc 63 Figure 4.8: a) Experimental Validation of Closed End Stop Deformation. b) Equivalent ATF Needle Lift vs Control Chamber Force. Figure 4.9 shows a FEM validation of the closed end stop. When the test fixture is closed, the chamber pressure from the top presses against the spring from the bottom. The differential force acts on the ears of the end stop block. The green arrow at the bottom of the ears of the end stop block shows that the fixture is fixed in that region, and the purple arrow on the top of the ring shows that 2050 of force is acting from the control chamber. The end stop stiffness observed from Figure 4.9 in the FEM analysis is .= 6,743.4/, which is almost 3.5 times less than the measured stiffness (24,117.6/) in Figure 4.8b. Stiffness of 6,743.4/ is used in the Simulink model because the stiffness measured from the experiment in Figure 4.8b is very much dependent on how much the lock nut is tightened into the end stop and the correct torque for the locknut is difficult to achieve experimentally. b) a) 64 Figure 4.9: FEM Analysis of Closed End Stop. Since damping is difficult to measure, it is modelled by trial and error when fitting the system simulation to experiments in Section 4.3.1.  is set as 550 and the closed end stop damping is modelled as 550∗3. The model matching for damping coefficient is verified in Section 4.3.1. Open End Stop The stiffness of the open end stop is tested both experimentally and using a FEM Analysis. There are two effects influencing the open end stop stiffness: a) Elastic deformation of the plunger disc against the ATF body. b) A Viscor film between the plunger disc against the ATF body. The open end stop stiffness is identified experimentally by adjusting the relief valve, starting from zero supply pressure with the actuator open. Figure 4.10a shows the experimental verification of the open end stop stiffness, where the supply pressure is gradually increased, and the Fixed Ears 65 corresponding ATF needle displacement is recorded. Figure 4.10b shows the relationship between the ATF Needle lift and control chamber pressure. The needle slightly lifts until the pressure overcomes the preload of the lifting force spring at 18 which is equivalent to 1450 of force. This measurement measures the spring pressure into the open end stop. Therefore, the open end stop stiffness is estimated at .= 69,047.6/. Figure 4.10: a) Experimental Validation of Open End Stop Deformation. b) ATF Needle Lift vs Control Chamber Pressure Graph Figure 4.11 shows the FEM analysis of the closed end stop. When the ATF opens, the lifting force spring applies a preload of 1450 onto the bottom of the plunger disc holding the spring and the top side of the disc is pressed against the test fixture body as shown in Figure 4.11. A small amount of force (720) is applied on the plunger from the chamber. The disc is fixed around the sides. a) b) 66 Figure 4.11: FEM Analysis of Plunger The observed deflection at the bottom of the disc in Figure 4.11, is 0.000014 with a stiffness of .= 1.4 ∗10/. The plunger stiffness is much larger than the closed end stop stiffness and can be neglected. This value fits the model better than the value obtained from Figure 4.10b. Therefore, the stiffness of the open end stop is 1.4 ∗10/. The damping is very difficult to measure. Therefore, it is also identified when matching the system model to experiments in Section 4.3.1, similar to the closed end stop. The damping is modelled as 550∗2. The model matching for damping coefficient is verified in Section 4.3.1. 67 Identification of the Lifting Force Spring Preload As discussed in section 4.1.1, the lifting force spring simulates the combustion force and the force due to fuel pressure in the cylinder. Based on the calculation shown in section 4.1.1, the compression nut was set to 4 turns to achieve 1250 of preload by the lifting force spring for the measured spring stiffness of 432/. However, since it is very difficult to determine the zero position of the spring in the setup, the actual preload of the lifting force spring needs to be identified and matched in the Simulink model. The preload of the lifting force spring is identified experimentally by slowly increasing the pressure in the control chamber from zero pressure by adjusting the relief valve. Figure 4.12a a plot of the resulting control chamber pressure corresponding ATF needle displacement. From Figure 4.12b, it is observed that ATF Needle opens at 18 which corresponds to 1450. Figure 4.12: a) Experimental Verification of Lifting Force Spring Preload. b) Control Chamber Pressure vs ATF Needle Lift a) b) 68 4.2 Hydraulic Modelling The hydraulic model consists of a variable volume control chamber, that is fed through the inlet orifice and that bleeds out through the outlet orifice and leakage flows. Figure 4.13 illustrates the hydraulic subsystem in the ATF model. The remainder of this section describes the components and the associated parameter identifications of the hydraulic model. Figure 4.13: Hydraulic Subsystem. Inlet Orifice Modelling Turbulent flow is assumed to occur in the inlet orifice and the outlet orifice since the high-pressure fluid is supplied. Therefore, according to according to Nam et al., (2004), the flow rate through the orifice is proportional to the square root of the pressure drop. 69  =  ∗2( − ) (4.3) Where, i)  () is the volumetric flow rate into the orifice. It is a measurement parameter, ii)  () is the supply pressure entering the inlet of the injector which then passes through the orifice, iii)  () is the area of the inlet orifice determined from measurements. It is an important parameter, iv)  is the flow constant of the inlet flow obtained by performing characterization experiments described in section 4.2.4, v)  () is the chamber pressure, vi)  () is the density of the diesel. Outlet Orifice Modelling  =  ∗2( − ) (4.4) Where, i) () is the volumetric flow rate out of the injector through the solenoid valve, ii)  () is the diesel drain pressure, iii)  () is the actuator orifice. It is an important parameter for the opening of the valve. A large  and small  reduces the control chamber pressure quickly. If  is too large then pressure cannot be reduced, and the pump keeps on supplying. 70 Therefore,  must be always smaller then , otherwise the ATF needle will open very slowly, vii) is the flow constant of the outlet flow obtained by performing characterization experiments described in section 4.2.4. Leakage Modelling The leakage is through the match fit between the plunger and the ATF body as shown in Figure 4.2. The flow is laminar since the fluid moves in a smooth path.  = −  (4.5) Where, i)  () is the leakage flow between the gas plunger and the injector body which is in relative motion and it is modelled as laminar flow, ii)  =  () where  () is the kinematic viscosity, iii)  () is the needle diameter, iv)  () is the diametric clearance, v)  () is the length of leakage. Geometrically it is the length of the plunger that is in contact with the ATF body through which leakage takes place, vi)  is the atmospheric pressure. 71 Identification of Flow Constants and Leakage The flow constants mentioned in Section 4.2.1, Section 4.2.2, and Section 4.2.3, , , and  are obtained by characterization experiments. Outlet and Inlet Flow Characterization For instance, fuel coming out of the outlet orifice is measured after 10 injections, and its volume is calculated. Equation 4.4 can be rewritten as Equation 4.6 and 4.7. Where,  is the actuator needle lift and  ℎ is the solenoid valve width.  =  ∗ ℎ ∗2( − ) (4.6)  =  ∗ ℎ ∗ ∗2∗√ −  (4.7) Volume of fluid collected:  =   = 1.1 ∗10  (4.8)  =  ℎ ∗ ∗2∗ ∗√ −   (4.9) Model Identification  = ℎ ∗2 ∗∫  ∗√ −   (4.10) Therefore, numerical integration is performed on Equation 4.10 using Matlab to calculate  =6.6. Similarly, the same experiment is performed to characterize the inlet flow constant, = 5.5. 72 Leakage Characterization: The leakage coming out of the ATF needle is collected for 2 minutes, and its volume is calculated. Equation 4.5 can be rewritten as  =1 −  (4.11) Volume of fluid collected:  =1 −   (4.12) Model Identification  =1 −   (4.13)  = −  (4.14) Typical value for  = 3.3 ∗10 . From Equation 4.13, = 7.7 ∗10  Verification of the Inlet, Outlet, and Leakage Flow Type In Section 4.2.1 and Section 4.2.2, it is assumed that the inlet and outlet flows are turbulent. To prove this theoretically, the orifices are assumed to be a narrow pipe and their respective Reynold’s numbers are calculated using the following equation (Seethaler, n.d):  =2 (4.15) 73 Where, i) is the equation of Reynold’s number of flow through a pipe, ii) is the volumetric flow rate and its typical value for injector is,  = 5.77 ∗10 and  = 4.23∗10 , iii)  is the kinematic viscosity of Viscor Calibration Fluid and its value being 3.0545∗10, iv)  is the radius of the orifices. The Reynold’s number for the inlet orifice is 3.33∗10, and for the outlet orifice it is 2.75∗10. Both Reynold’s numbers are greater than 2300. Hence, the orifices have turbulent flow. In Section 4.2.3 it is assumed that the leakage flow is laminar. To prove this theoretically, the flow is assumed to pass through a slot and the respective Reynold’s number is calculated using the following equation (Seethaler, n.d):  =2( + ) (4.16) Where, i) is the equation of Reynold’s number of flow through a slot, ii)  is the leakage volumetric flow rate and its typical value for injector is 3.3 ∗10, iii)  () is the needle diameter, iv)  () is the diametric clearance, The Reynold’s number for Leakage is 23.73. Therefore, the leakage flow is laminar. 74 Hydraulic Subsystem Modelling All the parameters modelled in Section 4.2 are combined in the subsystem shown in Figure 4.13. using continuity and compressibility equations for the control chamber:  = 2+  (4.17) Where, i) ∑  () is the net flow rate into the chamber, ii)  () is the rate of increase of fluid volume in the chamber due to fluid compressibility, iii)  () is the volumetric flow rate created by the displacement of the needle, iv)  () is the contact area of the fluid on the needle (Satkoski et al., 2011), Equation 4.4 is rewritten as: = +  ∗2( − )−  ∗2( − ) ∗2 (4.18) Where,  =  −__  (4.19) is the changing volume due to the plunger motion,  is the fixed volume of the chamber, and__ is the maximum needle lift. 75 4.3 System Model The components of the ATF such as the inlet, outlet flow rates, equation of motion of the needle, end stop stiffness, damping, and the lifting force spring are modelled in Sections 4.1 and 4.2. These models are combined, and the Simulink system model is shown in Figure 4.1a. There are two inputs to this model, the experimentally measured actuator needle lift and experimentally measured supply pressure. ATF needle lift and control chamber pressures are the outputs of the model. In an actual engine, it is not possible to access the actuator and ATF needle displacements, and chamber pressures. But the rail pressure in an engine can be measured easily. That is why the measured line pressure and actuator displacement can predict the chamber pressure and ATF needle displacements accurately. The following section discusses how these simulated model outputs compare to experiments. Characterized Model Result The simulation uses a fourth-order Runge-Kutta solver with a fixed step size of 5e-8s to solve the differential equations. The model is fed with measured values of inlet pressure and actuator displacement to calculate the simulated outputs, ATF needle lift and control chamber pressure. These outputs are compared to the measured outputs in this section. The model is also verified across a number of different operating points by varying the supply pressure and the actuator pulse duration. Figures 4.14, 4.16 - 4.19 show experimental model comparisons for actuator pulse durations of 2,1.5,1,0.8, and 0.5 respectively. Within each figure, roman labels a-e depict subfigures with supply pressures of 25.6,24.6,22.6,21.6, and 20.6 76 respectively. Within each roman labeled subfigure, subfigures α depict measured actuator lift in yellow (input for Simulink model), measured needle lift in blue, and simulated needle lift in red (output of Simulink model). Subfigures β depict measured supply pressure (input for the Simulink model) in dashed yellow, the measured chamber pressure in blue, and simulated chamber pressure in red (output of Simulink model). Due to the low spring stiffness of the closing end stop, lift increases with rising supply pressure, since the needle is pressed further into the closing end stop. This is indicated by a negative initial valve position of up to −0.1. The distance between end stops is approximately 0.1 and the open end stop deforms approximately 0.05 due to the preload of the lifting force spring. Overall this makes for a maximum lift of up to 0.25. It is observed that, for 2,1.5, and 1 pulse duration, the needle spring is able to force the needle to the full 0.05 into the open end stop. At the shorter pulse durations of 0.8 and 0.5, the pressure in the control chamber does not drop sufficiently to allow for the full needle lift, and the needle is driven only partway into the open end stop, reducing the overall lift. Figure 4.14 a. shows the simulation results for 25.6 supply pressure and 2 actuator pulse duration. At time zero, the actuator is closed, and the chamber pressure sees the full line pressure. This pressure drives the plunger and needle up to −0.1 into the closed end stop. When the actuator solenoid valve opens, pressure in the control chamber drops, since fluid exits through the output orifice. Once the spring force exceeds the force of the hydraulic fluid on the needle in the control chamber, the ATF needle starts to open, since the lifting force spring contracts. This is observed at 1.4, when the control chamber pressure remains constant until 2, after which it 77 starts to drop again. The flat pressure region is indicative of when the valve has started moving and the plunger replaces part of the volume of fluid exiting through the exit orifice. Once the plunger is slowed down by the open end stop, the control chamber pressure drops further to the point where the line pressure is used entirely to overcome friction losses in the input and output orifices. Velocity of the plunger and needle are constant during opening. The ATF needle is fully open with a lift of 0.18, when the chamber pressure is minimum. The opening rate of the experimental and the simulated ATF needle is almost identical, along with the maximum lift of about 0.18. When the actuator valve is closed, the pressure in the control chamber starts to build up. The ATF needle starts to close only when the chamber pressure exceeds the preload of the lifting force spring. This is observed in the chamber pressure at 3.6, where the pressure remains constant until 4, after which it starts to build up slowly. The flat pressure region starts when the needle leaves the opening end stop and it ends when the needle reaches the closing end stop. At this point the needle slows down, since the closing end stop is much stiffer than the lifting force spring and it takes more pressure buildup to compress. The simulated needle starts to close a bit earlier than the experimental needle, but the experimental needle comes back to the closed position earlier than the simulated needle. This can also be correlated with the control chamber pressure, where the measured chamber pressure increased faster back up to 25.6. These differences between model and experiment are likely due to inaccurate friction modelling. The viscous friction model for the end stops should likely be improved by adding dry friction terms similar to a Stribeck model. Such a model would also be able to explain that for all line pressures, the chamber pressure needs to drop/rise by 5 before any needle motion takes place during the opening/closing motion. Nevertheless, the experimental-model control chamber pressures and needle lifts are very similar. 78 Figure 4.14b – e show experiments with the same actuator pulse duration of 2, but progressively lower supply pressure. Both simulation and experiment indicate that with reduced supply pressure, opening needle velocities increase, and closing velocities decrease. This is not surprising since a lower line pressure leads to lower chamber pressure, which then speeds up opening and slows down closing. It is also observed that with reduced supply pressure, the simulated ATF needle starts to open and close a bit faster than the measured ATF needle, along with a bit higher lift. At the same time, the simulated chamber pressure also drops more than the measured chamber pressure. This is likely due to the simplified friction model as described in the previous paragraph. The sticking friction is quite high as it can be seen in Figure 4.14 a.α, the pressure drop between opening and closing is 2.5MPa in either direction. When the differences between supply pressure and spring lifting force gets close to that limit, the simulation shows large errors. As expected this discrepancy increases, as the line pressure drops and the unmodeled constant dry friction term becomes more significant. In Section 4.1.3, it was discussed that the end stop damping coefficient was set by model matching. It can be observed from Figure 4.14 that the response is consistent with the type of damping modeled and it is also validated across different pressure range and actuator pulse duration. The ATF needle lift and the deflection in the end stop can be compared with the results obtained by Seykens et al., (2005) as shown in Figure 4.15. From the figure, it can be observed that at a high-pressure the closed end stop deflects up to −0.15, and with reducing pressure the deflection reduces. The injector model is underdamped compared to the ATF model as it can be 79 observed from the needle lifts. The needle lift is approximately 0.24 which is also very similar to the ATF needle lift (0.18) obtained in Figure 4.14. a.α) a.β) b.α) b.β) 80 Figure 4.14: Experimental – Model Comparisons for Various Pressures at 2ms Pulse Duration and Respective Flow Rates. a) 25.6MPa. b) 24.6MPa. c) 22.6MPa. d) 21.6MPa. e) 20.6MPa. d.α) e.α) e.β) d.β) c.α) c.β) 81 Figure 4.15: Experimental and Simulated Injector Needle Lift. Reproduced from “Detailed Modeling of Common Rail Fuel Injection Process,” by X.L.J. Seykens, L.M.T. Somers, & R.S.G. Baert, 2005, MECCA, 3(2-3), 36. Copyright 2005 by Xander (X.L.J.) Seykens, and Bart (L.M.T.) Somers. Reprinted with Permission. Figure 4.16 and Figure 4.17 show the same trends for shorter pulse durations of 1.5 and 1 respectively. Initially, the chamber pressure drives the needle −0.1 into the closed end stop. The experimental needle opens and closes a bit earlier than the experimental needle. This is also due to inaccurate friction modelling which is discussed for Figure 4.14. With decreasing pressure, the opening speeds up and the closing slows down. It is also observed that with decreasing pressure, the ATF needle lift decreases due to the low stiffness of the closing end stop. However, the experimental-model control chamber pressures and needle lifts are very similar. 82 a.α) a.β) b.α) b.β) c.α) c.β) 83 Figure 4.16: Experimental – Model Comparisons for Various Pressures at 1.5ms Pulse Duration and Respective Flow Rates. a) 25.6MPa. b) 24.6MPa. c) 22.6MPa. d) 21.6MPa. e) 20.6MPa. d.α) d.β) e.α) e.β) a.α) a.β) 84 b.α) b.β) c.α) c.β) d.α) d.β) 85 Figure 4.17: Experimental – Model Comparisons for Various Pressures at 1ms Pulse Duration and Respective Flow Rates. a) 25.6MPa. b) 24.6MPa. c) 22.6MPa. d) 21.6MPa. e) 20.6MPa. Figure 4.18 illustrates the experimental-model comparison for partial ATF needle lifts at 0.8 of actuator pulse duration for various supply pressures. One of the most important task of this thesis, is to predict partial ATF needle lifts. In case of partial lifts, the control chamber pressure does not drop sufficiently for the ATF needle to open completely, and the needle is driven only partway into the end stop, reducing the overall lift. This is observed in Figure 4.18 and Figure 4.19. Figure 4.18a shows the simulation results for 25.6 supply pressure and 0.8 actuator pulse duration. It shows a partial lift at 0.12. When the actuator valve opens, the simulated ATF needle opens a bit earlier than the experimental needle, but with the same lift. At 1.5, the ATF needle starts to open. The chamber pressure remains constant until 2. When the actuator valve is closed at 2, the pressure in the control chamber starts to build up. The ATF needle starts to close only when the chamber pressure exceeds the preload of the lifting force spring. At 2.5 the needle enters the closing end stop, and at this point the needle slows down, since the closing end stop is much stiffer than the needle spring and takes more pressure buildup to compress. e.α) e.β) 86 With reducing supply pressure from Figure 4.18b - e, it is observed that the opening rate remains the same, whereas the simulated ATF needle starts to close earlier than the experimental ATF needle. This is also evident from the control chamber pressure, as the simulated pressure builds up faster than the experimental pressure. At high pressure the model fits well, as the pressure reduces towards the preload of the lifting force spring, the model shows high discrepancy. This is likely due to the unmodeled dry friction. Similar trends are observed when the actuator pulse duration of 0.8 in Figure 4.18 is reduced to 0.5 in Figure 4.19. a.β) b.β) a.α) b.α) 87 Figure 4.18: Experimental – Model Comparisons for Various Pressures at 0.8ms Pulse Duration and Respective Flow Rates. a) 25.6MPa. b) 24.6MPa. c) 22.6MPa. d) 21.6MPa. e) 20.6MPa. e.β) e.α) c.α) c.β) d.α) d.β) 88 b.α) b.β) c.β) c.α) a.α) a.β) 89 Figure 4.19: Experimental – Model Comparisons for Various Pressures at 0.5ms Pulse Duration and Respective Flow Rates. a) 25.6MPa. b) 24.6MPa. c) 22.6MPa. d) 21.6MPa. e) 20.6MPa. d.β) e.β) d.α) e.α) 90 Chapter 5: Conclusion The objective of this research was to develop a model that would describe the relationship between actuator needle displacement and injector needle displacement for dual fuel injectors. To achieve this, an ATF, with the guidance of Westport Fuel Systems Inc., was designed and created to provide access to measurement parameters, such as actuator displacement, injector displacement, line pressure, and chamber pressure. The ATF was used to experimentally characterize and validate a model for the hydraulically actuated fuel injectors, the model was validated by comparing its performance with the experimental results. The key parameters of the ATF such as the control chamber volume, orifices, mass of the needle and plunger, were designed to be similar to typical hydraulically actuated fuel injectors so that realistic needle lifts were obtained from the ATF. An electromechanical solenoid actuator was used to trigger the ATF. The ATF test setup allowed characterizing the inlet and outlet flow rates, end stop stiffnesses, spring stiffnesses, and flow constants. The model that was developed imitates the injection process of a typical hydraulically actuated fuel injector very well. Even though the experimental-simulation comparisons were not 100% accurate, the model outputs were very close to the experimental measurements. For instance, the control chamber pressure trace from the model was similar to the experimental results, and it was reflected in the corresponding simulated ATF needle lift. The control chamber pressure also showed fluctuations around certain regions which indicated when the ATF needle starts to open and hits the end stops. 91 The model was verified across several different operating points by varying the supply pressure and the actuator pulse duration to increase confidence in the experimental-model results. For longer pulse duration, the full lift of the ATF needle was observed. Much of the needle displacement was in the closed end stop due to its low spring stiffness. However, for shorter pulse duration, partial lifts of the ATF needle were observed, since the chamber pressure did not drop sufficiently enough to achieve full lift. The chamber pressure and ATF lift data retrieved from the experiment and model at different pressures and pulse duration showed mostly consistent results with only few irregularities. The experimental-model data indicated that, with reduced supply pressure, the simulated ATF needle opened and closed a bit faster than the experimental needle due to inaccurate friction modelling. At high pressure, the model fits very well but as the pressure reduces discrepancies are observed. The tolerances involved in fabricating the ATF, especially the match fit between the plunger and brass sleeve, are large compared to a typical injector, which leads to more discrepancies. Nevertheless, the experimental-model control chamber pressures and needle lifts are very similar. 5.1 Limitations of ATF Even though the ATF simulates a typical diesel fuel injector, there are some limitations. i) The ATF is confined between the pressure range of 18MPa to 25MPa since the hydraulic fittings are rated at 27.6 MPa and the preload of lifting spring is equivalent to 18MPa. ii) The end stop stiffnesses are lot softer than a diesel fuel injector which alters results. iii) The lifting force spring can deflect up to maximum 3.6 which limits its preload at 1550. High combustion force and fuel force cannot be simulated with this spring. 92 5.2 Recommendation for Future Work The developed model of the ATF is not 100% accurate. Some parameters can be tweaked more and further characterized with better instruments. Due to time restrictions, these issues could not be addressed in this thesis. The list of recommendations are listed below to model a more accurate ATF. ATF Improvement i) Inlet and outlet flow rates of the ATF are estimated in this research. Consequently, the flow constants are also estimated. To increase the accuracy of flow measurement, Coriolis mass flow meters can be used to measure the flow rates. ii) As discussed in Section 3.1.4, the end stop should deflect by 1% or 2% of the overall lift under load. The end stop should be redesigned in future versions of the ATF. iii) The brass sleeve discussed in Section 3.1.2 guides the plunger in and out smoothly, with good tolerance. However, the brass sleeve is held in place by a retaining ring which does not ensure that it will be fixed rigidly to the ATF body. While the plunger is reciprocating under pressure, the brass sleeve could also move up and down, adding moving mass to the plunger and needle. This would highly affect the dynamics of the needle, providing inconsistent measurements. This issue can be addressed by screwing a nut against the brass sleeve instead of retaining ring. 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Retrieved from http://www.euro.who.int/__data/assets/pdf_file/0006/189051/Health-effects-of-particulate-matter-final-Eng.pdf 99 Appendices Appendix A - Mechanical Drawings This section contains all the mechanical drawings of the components used to make the ATF. A.1 ATF Body Front View Cross-Section 100 A.2 ATF Body 101 A.3 Plunger 102 A.4 Needle 103 A.5 Brass Sleeve 104 A.6 Closed End Stop 105 A.7 Compression Nut 106 A.8 Lock Nut 107 Appendix B - Experimental Procedures This section describes the experimental procedures. i) Connect the actuator to the MACCON power amplifier and the amplifier to dSpace. ii) Connect Amplifier to power supply and turn on dSpace and start dSpace software. iii) Turn on the Polytec Laser Sensor and wait for the LED lights to flash which shows that the controllers and sensors are ready to use. iv) Connect the air supply line to the hydraulic pump and the hose of the pump to the hydraulic fittings on the test rig. v) Switch on the air compressor and turn the pressure switch to the auto position. Adjust the air pressure regulator to 0.62MPa and check its air pressure gauge. vi) To increase the cracking pressure in the Relief Valve, turn the cracking pressure adjusting nut clockwise making the spring more compressed. vii) Step on pump foot pedal and check hydraulic gauge to ensure that the correct pressure is reached. viii) Make sure that dSpace reads the same pressure from the absolute pressure sensor. ix) Set the correct current (20A) and pulse duration (2ms) in dSpace. x) Enable the amplifier. xi) Send current pulse to the actuator. xii) Have dSpace record pressures and displacements. xiii) Disable the amplifier. xiv) Release the pump foot pedal and depress the release end of the pump pedal to release the pressure. xv) Switch off the sensors, amplifier and the dSpace. 108 Appendix C - Calculations This section shows all the calculations for thickness of the ATF body and thread stripping calculation for the compression nut. C.1 Thickness of ATF Body The control chamber sees a pressure of 27.6MPa. The ATF body has an outer radius of 27mm. Steel 4140 has a yield strength of 415MPa and a tensile strength of 655.  =  ∗2 ∗ (C.1) Where, i)  = thickness of the ATF steel body. ii)  = Pressure. iii)  = tensile strength of steel. Therefore, the ATF body must be 1.14mm thick to withstand the high pressure with a safety factor of 13.6. C.2 Thread Stripping Calculation  = 4( − 0.938194∗) (C.2)  = 2 ∗ (C.2) Where, i)  = Stress area. ii)  = Shear area. 109 iii)  = Pitch of the nut. iv)  = Diameter of the nut.  = ℎ ℎ ∗ (C.3) The threads on the ATF body has a thread shear area of 4302.9, with a shear strength of 390, and it can withstand upto 1.6MN of force. The force applied on the threads by the end stop is approximatly 2kN. Therefore, the threads can also handle the high pressure. C.3 ATF and Actuator Clamping Force  =  ℎ ∗ (C.4) The Area of the copper gasket = 318.27 , and Yield strength of copper = 80. So, the force required to get a good seal  = 80 ∗ 318.27  = 2,5461.6 (C.5) Therefore, each of the three M5 bolt needs to provide a force of 8,487.2.  = 450 ∗ 14.18  = 6,381 (C.6)   =6,3818,487.2= 0.7 (C.7) However, with a tensile yield strength of 450 and a tensile stress area 14.18 for M5 bolts, it can withstand 6,381 of force. A safety factor of less than 0.7 indicates that the three M5 bolts are too weak to achieve plastic deformation of the copper seal. 110 C.4 Clamping Force The Tensile strength of Grade 5, 5/16-inch bolt = 827.4, the yield strength = 634.3, and tensile stress area = 56.6.    =  ℎ ∗ = 827.4 ∗56.6 =60 (C.8)    =  ℎ ∗ = 634.3 ∗56.6 = 46 (C.8) Thus, the two bolts provide a combined tensile yield load of 46 and an ultimate tensile load 60. This is sufficient to provide the required palstic deformation of copper seals with a safety factor larger than 2. Appendix D - Matlab Code This section shows the Matlab Code. ATF Characterization clear all close all clc load('bufexp_188.mat') data=bufexp_188; tmax=0.006; index=find(data.X.Data<tmax); time_m = data.X.Data(index); P_chamber_m = data.Y(3).Data(index)*1e6; %Chamber Pressure [Pa] P_line_m = (data.Y(5).Data(index))*1e6; %Line Pressure [Pa] x_actuator_m = data.Y(7).Data(index)*1e-3; %Actuator Displacement [m] x_actuator_m = max(x_actuator_m,0); %Actuator Displacement [m] x_injector_m = (-(data.Y(8).Data(index)-data.Y(8).Data(1))*1e-3); %Injector Displacement [m] P_line_m=P_line_m-(P_line_m(1)-P_chamber_m(1)); 111 Parameters Pressure values Pin = % Diesel Supply Pressure [Pa] Pout = % Diesel Drain Pressure [Pa] Patm = 101.325e3; % Atmospheric Pressure [Pa] Hydraulic Circuit Parameters rho = 825; % Density of viscor fuel [kg/m^3] Ain = % Gas feed orifice [m^2] Aout = % Valve Area [m^2] Adiesel = % Effective area of diesel hydraulic force acting on top of gas plunger [m^2] Vc= % Nominal Chamber Volume [m^3] dVc = % Varying Chamber Volume [m^3] Beta = 1300e6; % bulk Modulus of diesel [Pa] y_needle_max = 120e-6; Fs0 = 18.0e6*Adiesel; % Combustion and Preload force [N] m = % Moving mass of gas plunger [Kg] k = 417000; % Spring constant(N/m) 547000 kec=(5e6*Adiesel)/0.0725e-3; % measured keo=kec*2; % k*10 c=550*1; % Damping Coefficient (Ns/m)750 - 500 cec=550/y_needle_max*3; % Damping Coefficient (Ns/m)1100 - 3100 ceo=550/y_needle_max*2; % Damping Coefficient (Ns/m)1100 - 3100 b1=0.04; % land length Aact = % Ball surface area [m^2] Amax= % Max valve area [m^2] xmax = % Max armature lift [m] xtravel=xmax; % Max armature travel [m] vw=Amax/xtravel; % Valve width [m] xmin = 0; y_needle_0=-((P_chamber_m(1))*Adiesel-Fs0)/kec; x_injector_m = (-(data.Y(8).Data(index)-data.Y(8).Data(1))*1e-3)-((P_chamber_m(1))*Adiesel-Fs0)/kec; Outlet Characterization dt=time_m(2)-time_m(1); deltaP = P_chamber_m - Patm; % Change of pressure [Pa] mass_measured = 0.1/1000;%0.00037 % Measured MAss [kg] /3 0.175 Vol_measured = mass_measured/rho; % Calculate volume measured [m^3] integral_act=filter(dt,[1,-1],x_actuator_m.*sqrt(deltaP)); Cdact=Vol_measured./(integral_act(end))./vw./sqrt(2/rho); % Pressure Coefficient, measured. 112 Cdact=Cdact*1.0; Q_act=Cdact.*x_actuator_m.*sqrt(deltaP).*sqrt(2/rho).*vw; Inlet Characterization integral_in=filter(dt,[1,-1],sign(P_line_m-P_chamber_m).*sqrt(abs(P_line_m-P_chamber_m))); Cdin=Vol_measured./(integral_in(end))./Ain./sqrt(2/rho); % Pressure Coefficient, measured. Cdin=Cdin*.9; % Inlet flow coefficient Q_in=Cdin.*sqrt(P_line_m-P_chamber_m).*sqrt(2/rho).*Ain; Plots sim('injector_model_overall') figure('rend','painters','pos',[100 100 600 600]) subplot(2,1,1) plot(time_m*1e3,x_injector_m*1e3,y.time*1e3,y.signals.values*1e3,x.time*1e3,x.signals.values*1e3,'linewidth',3) xlabel('Time [ms]') ylabel('Displacement [mm]') legend({'ATF Needle Lift Lift meas','ATF Needle Lift sim','Actuator Needle Lift meas'},'FontSize',7) axis([0 5 -0.15 0.21]) %set(gca,'YTick',-0.1:0.05:0.25); grid on grid on subplot(2,1,2) plot(time_m*1e3,P_chamber_m*1e-6,p_chamber.time*1e3,p_chamber.signals.values*1e-6,time_m*1e3,P_line_m*1e-6,'--','linewidth',3) xlabel('Time [ms]') ylabel('Pressure [MPa]') legend({'Control Chamber Pressure meas','Control Chamber Pressure sim','Supply Pressure meas'},'FontSize',7) axis([0 5 5 26])
Thesis/Dissertation
eng
Mechanical Engineering
Vancouver : University of British Columbia Library
Attribution-NonCommercial-NoDerivatives 4.0 International
Graduate
Design and characterization of an actuator test fixture for diesel fuel injectors
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