ANU Big Dish Steam Engine Analysis, Review and Construction

James Shepherd u4663314

Supervised by Prof. Keith Lovegrove September 2010

A thesis submitted in part fulfilment of the degree of Bachelor of Engineering Department of Engineering Australian National University

Commercial In Confidence: This report contains information and descriptions that are not to be disclosed without prior consent from the Australian National University Solar Thermal Group.

This thesis contains no material which has been accepted for the award of any other degree or diploma in any university. To the best of the author’s knowledge, it contains no material previously published or written by another person, except where due reference is made in the text. James Shepherd 24 September 2010

Acknowledgements

I would like to thank Professor Keith Lovegrove for offering me the opportunity to contribute in a very small way towards the research at The Australian national University’s Solar Thermal Group.

To all staff and students currently working within the ANU Solar Thermal Group, thank you for your assistance and the pleasant conversations had during lunch. Special mention should go to Geoff Major who was eager to provide assistance with even the most minor requests in what was to me, a new environment.

I would also like to thank Ben Nash, Rob Gresham and Ljiljana Argy. Manufacturing of components required for the spare custom Lister Diesel Steam Engine at the Solar Thermal Group would not have been possible within the time constraints had Ben not assisted; especially as the ANU School of Engineering did not have a student workshop during this period due construction work.

Finally to friends and family, who have always been there, thanks for your patience and support during this busy period in my life.

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Abstract

This report investigates the condition and configuration of the current Lister HL4 Big Dish steam engine in the context of the inlet valve. The sources of failure within the mechanical pin actuating method of the inlet system are investigated along with one potential method to alleviate the potential for increased internal engine failure resulting from the complete deterioration of these pins.

Analysis of the peak stresses on the valve bash pins indicate that it is unlikely that a material will be found that possesses the strength, price and machinability to withstand the harsh conditions that these components are exposed to within the cylinder of the engine. As such these components will only ever be a wear part or consumable.

It is recommended that consideration be given to the replacement of the valve bash pins with solenoid actuated valve pins. It is hoped that this suggestion will lead to reduced potential for damage within the cylinder as a result of the complete failure of the current valve pins.

Initial calculations indicate that the bash pins could be replaced with solenoid valves operating at 40A and around 25V. The size of an appropriate solenoid coil appears to be capable of fitting within the space limitations of the steam chest of the HL4 Big Dish steam engine, and it could probably be implemented without major engine redesigns or costs incurred. Benefits to the implementation include increase motor control through variable timing and the ability to optimise the overall efficiency of the engine, through direct data analysis; as the expansion ratio of the engine could be controlled by crank angle rather than by ad hoc methods such as, deactivation of cylinders or installation of spacer plates.

However, it has been determined that running 12 solenoids of the size suggested will consume around 1kW of electrical energy or 2% of the total engine output if running at full capacity.

Should the solenoid valve concept be considered reasonable by the ANU Solar Thermal Group, there is plenty of scope within the implementation of a solenoid valve to the current steam engine, along with a control system and the subsequent analysis of the steam cycle within the engine for a PHD thesis.

Contents

Chapter 1 Introduction ...... 1 1.1 Thesis Impetus and Objectives ...... 1 1.1.1 Current Energy Sector Environment and Issues ...... 1 1.1.2 History of the ANU Solar Thermal Group ...... 2 1.1.3 Reciprocating Steam Engines and Energy Conversion ...... 2 1.1.4 STG’s Reciprocating Steam Engine ...... 3 1.2 Thesis Scope and Outline ...... 3 Chapter 2 Background and Previous Work ...... 5 2.1 Introduction ...... 5 2.2 Rankine Cycle Theory ...... 5 2.2.1 Feedwater Pump...... 7 2.2.2 Receiver Isobaric Heating ...... 7 2.2.3 Adiabatic Expansion ...... 8 2.2.4 Isobaric and Isothermal Condensation ...... 9 2.2.5 System Analysis ...... 9 2.3 Reciprocating Engine Theory ...... 10 2.4 The ANU Big Dish System ...... 12 2.4.1 Paraboloidal Dish Collector ...... 13 2.4.2 Solar Receiver ...... 14 2.4.3 STG’s Lister HL4 Diesel Conversion ...... 14 2.5 Prior Research and Conclusions ...... 16 2.6 Current Engine Conditions ...... 18 2.6.1 Valve Actuator Pins ...... 19 2.6.2 Head Gaskets ...... 20 2.6.3 Valve Ball-bearing and Valve Seats ...... 21 2.6.4 Cylinders ...... 21 2.7 Summary ...... 24 Chapter 3 Investigation into Valve Pin Failure ...... 26 3.1 Introduction ...... 26 3.2 Pin Failure Characteristics ...... 26 3.3 Valve Mechanical Operation ...... 27 v

3.3.1 Peak Contact Stresses ...... 28 3.3.2 Buckling Critical Loads ...... 33 3.4 Valve Bash Pin Geometric Analysis and Results ...... 34 Chapter 4 Valve Pin Material Analysis ...... 36 4.1 Introduction ...... 36 4.2 Material Selection Parameters ...... 36 4.3 Material Selection ...... 37 4.4 Conclusion ...... 40 Chapter 5 Valve Actuator Concept Review ...... 41 5.1 Introduction ...... 41 5.2 Current Bash Pin Design Issues and Potential Solutions ...... 41 5.3 Solenoid Actuation as a Potential Replacement ...... 42 5.3.1 Solenoid Operation and Advantages ...... 43 5.3.2 Solenoid Disadvantages ...... 43 5.3.3 Solenoid Valve Equations and Optimisation ...... 44 5.3.4 Summary ...... 49 Chapter 6 Conclusions and Further Work ...... 51 Appendix A: Worn Bash Pin Dimensional Analysis ...... 52 Appendix B: Bash Pin Material Analysis ...... 53 B.1 Stress and Deflection Coefficient ...... 53 B.2 Valve Inlet Pressure Balance ...... 53 Appendix C: Solenoid Calculations ...... 56 Appendix D: Consultations Conducted ...... 59 Appendix E: Manufacturing Cost Estimates ...... 62 Appendix F: BӦHLER S600 Material Specifications ...... 63 Appendix G: G-code for Manufacturing ...... 72 G.1 CNC Lathe G-code ...... 72 G.2 CNC Wire Cutter G-code ...... 78 Appendix H: Cylinder Bore and Piston Diameter ...... 81 Appendix I: Engine Component Blueprints ...... 83

List of Figures

Figure 1-1: ANU HL4 Customised Steam Engine Components ...... x Figure 2-1: Rankine Cycle Pv (left) and Ts (right) Diagrams ...... 6 Figure 2-2: HL4 Steam Engine PV Diagram (McIntosh, 2008) Compare to Otto Cycle ...... 10 Figure 2-3: ANU’s Big Dish Solar Thermal Power Plant (Siangsukone 2005, p.32) ...... 12 Figure 2-4: ANU’s New 500m^2 Big Dish (Lovegrove et al. 2009) ...... 13 Figure 2-5: Cross Section of SG3 Receiver Internals (Siangsukone 2005, p.40) ...... 14 Figure 2-6: STG’s Lister HL4 Steam Engine ...... 15 Figure 2-7: Cross Section of Four Processes within Two Stroke Steam Cycle ...... 16 Figure 2-8: Disassembly of Steam Engine Head Assembly ...... 19 Figure 2-9: Deterioration of Valve Lift Pins ...... 20 Figure 2-10: Crankcase Gasket Manufacturing ...... 21 Figure 2-11: Solidworks Volume Calculations of Complex Volume Beneath Valve Plate .... 23 Figure 3-1: Cylinder #1 Bash Pin Damage ...... 26 Figure 3-2: Pin-Valve Contact Coordinates ...... 28 Figure 3-3: Ball-bearing Deformed Face Area Equivalent ...... 32 Figure 3-4: Freshly Machined Valve Pins and Brass Test Piece ...... 34 Figure 4-1: Modulus of Elasticity and Endurance Limit Criteria Selection ...... 37 Figure 4-2: High Fracture Toughness and Oxidation at 500˚C ...... 38 Figure 4-3: Thermal Expansion and Hardness ...... 38 Figure 4-4: High Compressive Strength an Low Price ...... 39 Figure 4-5: Final Results of the CES Materials ...... 39 Figure 5-1: Steam Engine Solenoid Concept ...... 43 Figure 5-2: Dimension of a Solenoid Valve (Say 1964, p.22-27) ...... 46 Figure 5-3: Coil Voltage and Power Consumption as a Function of Current ...... 48 Figure 5-4: Coil Total Diameter Including the Plunger and Winding Thickness ...... 49 Figure A-1: Deformation Models for Worn Valve Pins ...... 52 Figure A-1: Copy of Piston Crown and Lift Pin Drawing with Incorrect Lift Pin Heights .... 53 Figure A-2: Stress and Deflection Coefficients for Two Bodies in Contact in a Point (Boresi and Schmidt 2003) ...... 53 Figure B-2: Pressure Differential Across Ball-bearing Prior to Inlet ...... 54 Figure F-1: Cylinder Liner Measurement Intervals ...... 81 Figure F-1: Cylinder Liner Diameter Measurements for the Soon to be Assembled Engine .. 81

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List of Tables

Figure 1-1: ANU HL4 Customised Steam Engine Components ...... x Figure 2-1: Rankine Cycle Pv (left) and Ts (right) Diagrams ...... 6 Figure 2-2: HL4 Steam Engine PV Diagram (McIntosh, 2008) Compare to Otto Cycle ...... 10 Figure 2-3: ANU’s Big Dish Solar Thermal Power Plant (Siangsukone 2005, p.32) ...... 12 Figure 2-4: ANU’s New 500m^2 Big Dish (Lovegrove et al. 2009) ...... 13 Figure 2-5: Cross Section of SG3 Receiver Internals (Siangsukone 2005, p.40) ...... 14 Figure 2-6: STG’s Lister HL4 Steam Engine ...... 15 Figure 2-7: Cross Section of Four Processes within Two Stroke Steam Cycle ...... 16 Figure 2-8: Disassembly of Steam Engine Head Assembly ...... 19 Figure 2-9: Deterioration of Valve Lift Pins ...... 20 Figure 2-10: Crankcase Gasket Manufacturing ...... 21 Figure 2-11: Solidworks Volume Calculations of Complex Volume Beneath Valve Plate .... 23 Figure 3-1: Cylinder #1 Bash Pin Damage ...... 26 Figure 3-2: Pin-Valve Contact Coordinates ...... 28 Figure 3-3: Ball-bearing Deformed Face Area Equivalent ...... 32 Figure 3-4: Freshly Machined Valve Pins and Brass Test Piece ...... 34 Figure 4-1: Modulus of Elasticity and Endurance Limit Criteria Selection ...... 37 Figure 4-2: High Fracture Toughness and Oxidation at 500˚C ...... 38 Figure 4-3: Thermal Expansion and Hardness ...... 38 Figure 4-4: High Compressive Strength an Low Price ...... 39 Figure 4-5: Final Results of the CES Materials ...... 39 Figure 5-1: Steam Engine Solenoid Concept ...... 43 Figure 5-2: Dimension of a Solenoid Valve (Say 1964, p.22-27) ...... 46 Figure 5-3: Coil Voltage and Power Consumption as a Function of Current ...... 48 Figure 5-4: Coil Total Diameter Including the Plunger and Winding Thickness ...... 49 Table A-1: Vertical Dimensions from Worn Components and Available Sources ...... 52 Table A-2: CAD Modelling Volume Comparisons for Outer Valve Pin ...... 52 Table A-3: CAD Modelling Volume Comparisons for Centre Valve Pin ...... 52 Table A-4: Effective Undeformed Bash Pin Height ...... 52 Table B-1: Resultant Force on Valve Ball-bearing due to Pressure Differential ...... 55 Table F-1: Distance Legend for Figure F-1 ...... 82

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Glossary of Terms

Terms Description

The Australian National University ANU Mechanical valve actuator situated on the pistons of the HL4 Lister steam Bash Pin engine, also called a lift pin. Bottom dead centre; indicating piston and timing position on a reciprocating BDC engine Before top dead centre; indicating piston and timing position on a BTDC reciprocating engine

CES Cambridge Engineering Selector software program

FBD Free Body Diagram

Solidworks Computer Aided Design software program

STG Solar Thermal Group, ANU

TDC Top Dead Centre

UNS Unified Numbering System

Glossary of Terms (continued)

Figure 1-1: ANU HL4 Customised Steam Engine Components

Nomenclature

Symbol Description Units

A Elliptical contact surface solution constant for the longitudinal axis B Elliptical contact surface solution constant for the transverse axis

Specific heat at constant volume kJ/kg∙K

Specific heat at constant pressure kJ/kg∙K

d Distance from the point of contact between stress elements mm

Modulus of Elasticity GPa

Enthalpy per unit mass at state i kJ/kg

Enthalpy per unit mass inclusive of irreversibilities kJ/kg

Mass flow rate kg/s

Condenser pressure kPa

Receiver pressure kPa

Heat transfer rate kW

Heat transfer at process i kW

Expansion ratio

Radius of curvature for contact element i about the mm longitudinal axis

Radius of curvature for contact element i about the the mm transverse axis

Entropy at state I per unit mass kJ/kg

Entropy of a reversible process at state i kJ/kg

Temperature at state i K

Nomenclature (continued)

Symbol Description Units

3 Cylinder volume at TDC m

3 Cylinder volume at BDC m

Work rate or power kW

Engine work output during operation W

Pump work input W

Dryness fraction

Specific heat ratio

Specific volume m3/kg

Isentropic efficiency

Otto cycle efficiency

Thermal efficiency

Angle between forces acting at the element centres of mass Radian

Poisson’s Ratio

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Chapter 1 Introduction

1.1 Thesis Impetus and Objectives

With the anticipated changes in global climate that are expected by many to result from carbon emissions, renewable energy is gaining a second wind in the political and financial markets. ANU has a long history of conducting research related to renewable energy technologies which dates back at least to the mid 1970s.

ANU’s Solar Thermal Group is on the cusp of connecting all of the individual components that will generate Big Dish solar thermal power from a new 500m2 paraboloidal dish concentrator, and as a result the condition of older components needed to be surveyed. One of the initial aims of this project was to complete the assembly of a spare customised steam engine for use as a backup in case the current working engine failed. The assembly of the spare motor has been delayed as a result of internal engine component wear and the lack of spare parts required. This project will instead explore the need for an investigation into the mechanisms used to construct the engine inlet valve and required components.

1.1.1 Current Energy Sector Environment and Issues

According to IEA statistics (International Energy Agency 2009) demand for energy worldwide nearly doubled between 1971 and 2007. During this same period the IEA believes that carbon dioxide emissions from worldwide fuel consumption has also increased by the same amount. A greater understanding of the links between carbon dioxide emissions and increases in atmospheric temperature has been established since the 1970s which has lead to increase political will towards the internalisation of externalities associated with the use of fossil fuels into the price of non-renewable energy.

In the last decade increased consumption and speculative trading has also lead to spot prices for oil, natural gas and coal increasing at a rapid rate. For the ten years prior to the Global Final Crisis in 2007, spot prices for these three commodities increased by around 500%, 200% and 170% respectively.

Subsequent to the crash in 2007 oil prices plummeted. However, even at their lowest 2007 levels the increases in oil price represented a 100% increase from 1997 prices. It also appears from the most recently released IEA information (International Energy Agency 2009), that neither coal or natural gas prices were significantly affected by the GFC and that during 2009 oil prices regained the upward prices trend with an approximate increase in spot prices of 30% for the year. These statistics suggest that there is a longer term upward trend in the trading price of fossil fuels that is increasing exponentially.

As a result of the increases in demand and the immanent internalisation of the environmental costs (through taxation) involved in the predominant forms of energy production, renewable forms of energy supply are becoming increasingly viable relative to fossil fuels. Thus energy sources that are less toxic to the environment are attracting increased interest from policy makers and investors worldwide.

1.1.2 History of the ANU Solar Thermal Group

The Australia National University has a long history of research into so called alternative energy sources, with particular emphasis on solar power generation.

ANU has been designing and researching solar thermal power generation since the 1970s and during the 1980s was involved in the construction of the White Cliffs solar array power plant. The White Cliffs plant used fourteen 20m2 paraboloidal dish collectors to concentrate solar radiation into a cylindrical cavity receiver where water was converted to steams with a temperature and pressure of around 540˚C and 7MPa respectively (Bannister 1991). The steam was then fed into a three cylinder Rankine cycle reciprocating engine which was connected to an alternator with any residual electrical energy produced being stored in batteries.

The White Cliffs solar power station was subsequently converted to a photovoltaic system in 1996 once the White Cliffs community was connected to mains power (Solar Systems 2009).

In more recent times ANU’s Solar Thermal Group has focused on developing larger scale dish concentrators. In 1994 the 400m2 SG3 model was completed and used in a research capacity until 2008, at which time efforts within the group had shifted to the new 494m2 dish collector which is expected to produce “around 100kWe in electrical capacity” with the aim of operating the dish at ANU to produce “steam that is superheated up to 500˚C and 4.5MPa” (Lovegrove et al. 2009).

As of the writing of this report, ANU’s new 500m2 solar dish concentrator has had several on sun trials with the return steam line connection to the engine room, upgrading engine control systems and the recommissioning of the Lister HL4 steam engine, being some of the larger tasks to be completed prior to electrical output from the system being competed.

1.1.3 Reciprocating Steam Engines and Energy Conversion

The basic principles involved in reciprocating steam engines were largely developed throughout the industrial revolution during the 18th century .With early attempts to convert steam energy into mechanical output suffering severe inefficiencies and mechanical design shortcomings, much credit is given to James Watt for “making the first successful rotative engine” (Hills, 1989). 2

Reciprocating steam engines were used in the later part of the 18th and throughout the 19th centuries as the technology was refined. However with the advent of the steam turbine (Cohen et al. 1996) and the internal combustion engine (Lumley 1999) during the early 20th century the popularity of the reciprocating steam engine has declined significantly.

Despite there being some negative tradeoffs related to the use of a reciprocating steam engine for power generation, the relatively low costs and complexities of these engines when compared to the alternatives can make these engines an attractive option.

As noted in Section 1.1.1 the current energy production environment is undergoing a demand shift towards renewable energy resources and within this context the production of electricity from solar dish concentrators is becoming more attractive. As has been noted by Bannister (1991 , p.11) the use of a reciprocating steam engine at the White Cliffs solar power station was seen as the best option due to technical, environmental and financial issues at this remote location.

Thus whilst it is accepted that the use of steam turbines are at present the most (thermally) efficient mechanism for the conversion of thermal energy in steam to electrical energy, it is the authors belief that the use of a low cost and robust reciprocating engine at the ANU Big Dish project will remain the most feasible option in the mid to long term.

1.1.4 STG’s Reciprocating Steam Engine

As mentioned in Section 1.1.2 the ANU’s Solar Thermal Group currently use a customised Lister HL4 Diesel engine to convert steam energy into mechanical power; from which an induction generator creates electrical energy.

Over several decades much work has gone into the design and construction of the current and previous Lister HL4 conversion as is described in greater detail in Sections 2.4.3 and 2.6. Similarly, significant investigations have been conducted into the efficiency of the engine within the context of the entire steam Rankine cycle of the Big Dish system and its peripherals (See Sections 2.5).

However, in recent years with the focus on the operation and construction of the dish collectors at ANU, deterioration of internal components (Section 2.6) and the potential for design improvements have gone unattended ( Chapter 4 and 0).

1.2 Thesis Scope and Outline

Topics to be covered within this thesis are largely the result of the need for a second Lister steam engine, for use as a spare or potentially to be used in parallel with the existing steam 3 engine; to capture as much of the new 500m2 dish concentrators productive capacity as possible.

As a result the chapters in this report follow the developments of the project as the review of the spare engine component’s conditions was conducted. Ultimately this lead to the identification of issues related to the current inlet valve concept, material and design that needed to be addressed.

Chapter 2 of the report provides some context with summarised information provided on the theory of the Rankine cycle, previous research conducted on the Lister engines and the condition of the working engine and spare engine components.

Chapter 3 covers the deficiencies of the bash pin inlet valve design through the observed failure of some examples extracted from the working steam engine. Chapter 4 seeks to determine whether the material believed to be used in the original production run of these pins was the optimal choice, whilst 0 investigates whether the mechanical pin actuation method used in the current engine could be replaced with a solenoid mechanism; and the related tradeoffs involved.

Subsequently the conclusions of this report are summarised in Chapter 6 with suggestions for potential further work that might be beneficial in the near future. In addition to these chapters a lot of information is presented in the Appendix that relates to the current design of the Lister HL4 customised components, including reproduced drawings of original parts for future reference and new drawings for parts that were altered and manufactured during this project.

Chapter 2 Background and Previous Work

2.1 Introduction

Several reports have been written previously that have included chapters on the measured efficiency of the Lister HL4 steam engine at ANU. Most of this work has centred on obtaining engine specific operational data to be incorporated into a larger computer model; so that the engines contribution to the entire power plant system can be analysed.

However, there does not appear to have been much work carried out on the internal mechanics occurring during the engines operation. As the current Lister steam engine had been mothballed for the last two years and there was no chance of steam being produced by the Big Dish during the period over which this project was compiled, this seamed an opportune time to investigate these aspects of the engine.

2.2 Rankine Cycle Theory

The ANU Big Dish solar thermal power plant utilises the Rankine for the conversion of solar thermal energy to mechanical power. Much of this theory has been covered succinctly by McIntosh (2008, p.5-9) with respect to the Big Dish system, however it is reiterated here for the readers benefit.

The ideal Rankine cycle is not the most efficient theoretical cycle for the conversion of Thermodynamic energy. The Carnot cycle achieves greater thermal efficiencies over the same temperatures but the impracticalities of accurately controlling phase change proportions during expansion and pumping wet vapour make the Rankine cycle the next best option for steam power generators in general.

Furthermore, the incompressibility of supercooled vapour used in the Rankine cycle reduces the pump work required to increase the pressure of the feedwater to the desired receiver pressure.

Figure 2-1: Rankine Cycle Pv (left) and Ts (right) Diagrams

Figure 2-1 presents the Rankine cycle process diagrams plotted for the temperature of the working fluid against its entropy, and fluid pressure verses specific volume. The process path 12345 and path direction indicated by arrow heads represent fluid flow throughout the system. From the feed tank (position 1), through the feed pump (process ) to the receiver where solar thermal heat energy is transferred into the fluid (process ),back down through the steam engine where the superheated steam is expanded (process ), and finally through the condenser where heat is dissipated to the surroundings (process 5 ) completes the closed loop.

Irreversibilities associated with the actual adiabatic expansion within the steam engine are also incorporated in Figure 2-1 through the process .

Furthermore, it should be noted that the process path representing the expansion of the steam within the engine in Figure 2-1 is actually a simplification that is only indicative of the net rate of change in steam properties across the engine. The actual steam cycle within the engine was mapped during experiments conducted by McIntosh (2008) and is presented in Figure 2-2 of Section 2.3.

Process Direction Process Name 12 Isentropic Pumping 234 Isobaric Heating 45 Isentropic Expansion 45' Irreversible Adiabatic 51 Isobaric & Isothermal Condensation

Table 2-1: Rankine Cycle Process Names

Error! Reference source not found. presents the names of the subsystem processes within the ideal Rankine cycle, for which energy equations are now presented. 6

2.2.1 Feedwater Pump

Under the assumptions that the feedwater pump is frictionless, perfectly insulated and that the increases in potential and kinetic energy across the pump are negligible, the power input by the pump can be obtained from a mass and energy balance across the control volume (Moran and Shapiro 2008).

Equation 2-1: Feed Pump Work Input

Equation 2-1 presents the rate of power input by the pump per unit mass. The values for the enthalpy at States 1 and 2 can be determined by observing the inlet and outlet pump gage pressures. The entropy values at state 1 can also be obtained from steam tables at the appropriate temperature and pressure.

2.2.2 Receiver Isobaric Heating

The ideal receiver heating process is as assumed to occur at constant pressure. Again, irreversibilities are neglected and the rate of thermal energy input per unit mass can be related to enthalpy and entropy as in Equation 2-2.

Equation 2-2: Dish Receiver Heat Flow

If the temperature and pressure achieved in the dish receiver are known, the entropy at this state can be determined from steam tables for the superheated steam phase boundary at these conditions. Alternatively if the steam is not superheated the dryness fraction needs to be considered using Equation 2-5.

Thus the rate of thermal energy input at the receiver can be calculated by approximating the inlet enthalpy from Equation 2-1; if the specific volume and pressures of the liquid are known at the pump.

2.2.3 Adiabatic Expansion

To calculate the ideal work output for the engine, steam quality at exhaust needs to be considered. By determining an approximate value for the vapour dryness fraction (Equation 2-5), the enthalpy and entropy values at States 4 and 5 can be derived; it is assumed that no irreversibilities exist and thus the values for entropy and enthalpy will be the same at these states as shown in Equation 2-3 and Equation 2-4.

Equation 2-3: Isentropic Expansion Entropy Equality

Equation 2-4: Isentropic Expansion Enthalpy Equality

However if the adiabatic expansion process is considered to include irreversibilities then the properties at each state (4 and 5’) need to be considered separately using the dryness fraction, values from steam tables and Equation 2-6 and Equation 2-7.

Equation 2-5: Dryness Fraction

Equation 2-6: Entropy Condenser Inlet if Condenser Pressure is Known

Equation 2-7: Entropy of Dry Saturated Vapour or at Condenser Inlet

Net energy output from the engine per unit mass of fluid can be calculated in relation to the enthalpy values at States 4 & 5’ as in Equation 2-8.

Equation 2-8: Work Output at Steam Engine

2.2.4 Isobaric and Isothermal Condensation

The condenser in an ideal Rankine cycle transfers heat from the wet vapour under assumed constant temperature and pressure until the phase change of the vapour to liquid has completed. Heat heat ejected into environment can be calculated for the condensation process using Equation 2-9.

Equation 2-9: Condenser Heat Flow

2.2.5 System Analysis

From the values for energy flow within each of the processes involved in the Rankine cycle, we can calculate the thermal efficiency of the complete system and the isentropic efficiency of the steam expansion within the engine by using Equation 2-10 and Equation 2-11 respectively.

Equation 2-10: Rankine Cycle Thermal Efficiency

Equation 2-11: Isentropic Efficiency

By observing the Pv and Ts diagrams in Figure 2-1 it becomes clear which aspects of the Rankine cycle need to be improved in order to obtain higher efficiency output from the system. For a set rate of solar thermal input and mass flow into the dish receiver, these measures would include increasing the pressure at the receiver or decreasing the temperature and pressure at which condensation occurs. 9

Research conducted by Bannister (1991) and McIntosh (2008) both experimented with measures that sought to increase the temperature and pressure of the inlet conditions to the steam engine by increasing the back pressure to the receiver dish. This research and their results are discussed in Section Error! Reference source not found..

Other measures that can help to improve the Rankine cycle efficiency include superheating, vapour reheat or supercritical cycles (Moran and Shapiro 2008); none of these are currently practical at the ANU due to material constraints within the STG’s current steam engine.

However under the assumption that the current or future reciprocating steam engine has excess thermal conversion capacity above that which the dish could supply, then implementing Preheat or Binary Vapour Cycles (Moran and Shapiro 2008) may become practical.

2.3 Reciprocating Engine Theory

The Big Dish steam engine is a customised conversion of a Lister HL4 diesel engine. The cycles of the engine are essentially the same as that of any two stroke engine; one expansion and one compression stroke. Further discussion on the operation and internal configuration of the Lister steam engine is presented in Sections 2.4.3 and 2.6.

The air standard cycle for the Big Dish steam engine is analogous to that of the Otto cycle as can be seen by the comparison in Figure 2-2; where the HL4 steam engine PV graph derived by McIntosh (2008) is compared to that for the ideal Otto Cycle.

Figure 2-2: HL4 Steam Engine PV Diagram (McIntosh, 2008) Compare to Otto Cycle

The Otto cycle is comprised of one reversible adiabatic compression (process ), one constant volume thermal increase (process ), one reversible adiabatic expansion (process 3 ) and one constant volume thermal decrease (process ). The process would correspond to the recompression and steam inlet, whilst the process would represents the expansion and exhaust of steam during the operation of the Big Dish steam engine.

As the Big Dish steam cycle is conceptually similar to the ideal Otto Cycle, the equations that relate to the efficiencies of the Otto Cycle can help to inform how the steam engines operation can be improved.

Equation 2-12 presents the equation relating Otto Cycle efficiency relative to the expansion ratio of the cylinder and the specific heat ratio (Equation 2-14).

Equation 2-12: Otto Cycle Efficiency related to the Expansion Ratio (Spencer, 2006)

Equation 2-13: Cylinder Expansion Ratio

Equation 2-14: Specific Heat Ratio

Clearly the efficiency of the Otto Cycle for any given ideal gas is dependent on the expansion ratio as the specific heat ratio is assumed to be constant. Direct experiments were carried out by Bannister (1991) to observe the effects of altering the expansion ratio on the first model of the Lister diesel to steam engine conversion; ANU’s HR3 Lister steam engine. Bannister research and results are discussed in Section Error! Reference source not found. of this report.

2.4 The ANU Big Dish System

ANU’s first Big Dish system was completed in 1994 and based around a 400m2 solar thermal dish concentrator. Since then a new 500m2 dish has been constructed but the basic system cycle has remained the same since the original SG3 version.

Figure 2-3: ANU’s Big Dish Solar Thermal Power Plant (Siangsukone 2005, p.32)

A simplified version of the system components and configuration is presented in Figure 2-3. The Big Dish system collects solar radiation and concentrates the energy into a receiver at the focal point of the paraboloidal dish using mirrors. Steam is generated from feed water pumped to the receiver as radiation bombards steel tubing within the receiver cavity.

Thermal energy is then transported to the engine room in the form of increased steam entropy as the feedwater pump pressure forces the superheated steam along the system. If the steam quality is determined by the engine control system to be high enough the engine is automatically turned over whilst pneumatic valves open to allow the steam to pass through the engine inlets. Steam is then allowed to expand within the cylinder and mechanical energy is created as a result of piston linear motion. When the steam is not of sufficient quality to operate the engine the control valve will divert the steam through the cooling tower and thus back into the feedwater tank.

The linear motion of the piston subsequently is converted into torque as it accelerates about the crankshaft within the Lister HL4 crankhouse. This torque about the longitudinal axis of the crank drives an induction generator and once the engine reaches 1200rpm a connection the mains power grid is established. It should be noted that the engine initially draws power from the grid until the point at which the engine speed reaches 1500rpm, whereby surplus energy generated is fed back into the mains system (McIntosh 2008).

As the engine reaches bottom dead centre an exhaust port is exposed allowing the majority of the steam to exit from the cylinder into the lower pressure condenser. From the condenser the steam is circulated through the cooling tower allowing the wet vapour to cool, condense and be pumped back into the feedwater tank.

2.4.1 Paraboloidal Dish Collector

The Paraboloidal dish structure acts as the support for the mirrors, receiver and piping for fluid flow up to the receiver and back down to the engine room. The position of the sun relative to the position of the dish is aligned through altitude-azimuth tracking. A summary of some of the interesting dish construction characteristics are presented in Error! Reference source not found..

Figure 2-4: ANU’s New 500m^2 Big Dish (Lovegrove et al. 2009)

Characteristic Value Aperture 494m2 Focal Length 13.4m Average Diameter 25m Average Rim Angle 50.1ο Mirror Reflectivity 93.50% Number of Mirrors 380 Mirror Size 1165mm x 1165mm Total Mass of Dish 19.1t Total mass of Base & Supports 7.3t

Table 2-2: Characteristics of 500m^2 Big Dish (Lovegrove et al. 2009)

2.4.2 Solar Receiver

The receiver for the new Big Dish is the same receiver that was used on the SG3 model dish. This receiver was mounted straight onto the new truss structure supports of the 500m2 dish after minor repairs and upgrades to electrical components were conducted.

The top hat shaped receiver encapsulates a stainless steel frame that is padded with Rockwool insulation and single length of steel tubing that is wound into a helix and bound to the inside of the steel frame (Siangsukone 2005). A cross section image of the internals of the receiver structure is presented in Figure 2-5.

Figure 2-5: Cross Section of SG3 Receiver Internals (Siangsukone 2005, p.40)

2.4.3 STG’s Lister HL4 Diesel Conversion

The current Lister HL4 based Big Dish engine was developed around the concepts of the original Lister HR3 customised steam engine. The HR3 engine concept was originally designed according to the requirement of the White Cliffs solar thermal power plant project during the 1970s and 80s.

White Cliffs is located in far eastern New South Wales at around 270km from Broken Hill. The isolated conditions of the town and limited technical experience of the town residents meant that the solution needed to be robust with moderate to low design complexity; as the local station owner was to be responsible for ongoing maintenance and repairs (Bannister 1991). As a result of the design constraints, the original HR3 steam engine was constructed with very basic inlet and exhaust mechanisms. Subsequently the HL4 version of the Big Dish engine also utilised this basic inlet setup.

The Big Dish steam engine is constructed from a Lister HL4 crank house, crankshaft, internal oil pump, main caps, connecting rods, starter motor and flywheel; the camshaft is removed and the oil pickup has been modified. The pistons, gudgeon pins and cylinder liners are parts from a Detroit Series 53 engine. Major components of the engine that have been specially designed and manufactured by ANU, including brass bushes for the small end of the connecting rods, cylinders, piston crowns, valve plates, valve guide plates and steam chests.

Figure 2-6: STG’s Lister HL4 Steam Engine

The steam inlet into the cylinder is controlled by three ball bearings in each cylinder head that are seated within a valve guide plate. Gravity, and the pressure differential across the seal that the ball bearing forms with the valve seat ensure that the expansion chamber is closed during periods when the inlet in meant to be shutoff.

Engine inlet timing is controlled by three alloy steel pins that are press fit into a crown plate which is secured to the machined piston crown by six M6 bolts with threads tapped directly into the piston crown. As the piston rises to TDC the bash pins engage the valve ball bearings, breaking the valve inlet seal and allowing steam to enter the expansion chamber.

Figure 2-7: Cross Section of Four Processes within Two Stroke Steam Cycle

2.5 Prior Research and Conclusions

Several investigations into the two models of the Lister steam engine conversion at ANU have been conducted. The major topic within these investigations that relates to the material covered in this report are the experiments conducted into efficiency gains from alterations to the volumetric displacement during expansion within the engine.

Bannister (1991) and McIntosh (2008) both experimented with a technique that McIntosh called variable displacement. However, the methods used by each individual to reduce the rate of volumetric expansion across the reciprocating steam engine were very different.

Bannister’s study aimed to investigate and model the solar thermal power plant components that had been developed by the ANU with particular attention being paid the working solar thermal power plant that had been constructed at White Cliffs by the university.

As part of this larger investigation and modelling process, Bannister attempted to vary the expansion ratio of the three cylinder HR3 Lister diesel engine. Bannister’s aim was to determine the effects of changes to the expansion ratio and incorporate the results into a numerical model of the entire Rankine cycle system with the hopes of identifying areas where the total system efficiency could be improved.

Bannister’s method employed manufactured steel plates inserted directly underneath the inlet valve plate of the engine. By reducing the clearance volume between the valve plate and the piston in this manner Bannister was able to experiment with three different expansion ratios; 1:10.1, 1:13.0 and 1:15.8 (Bannister 1991, p.100).

Whilst Bannister’s work was conducted on a different model of the Lister steam engine than the one currently being used at ANU some of his suggestions have significant relevance to the aims of this report.

Bannister notes that further investigation into the internal processes of the engine needs to be considered; with particular emphasis on the engine intake as this may not be reaching equilibrium (Bannister 1991, p.140). Whilst McIntosh later attempts to address the issue of investigating the internal processes further by using variable displacement, actually studying the effects of altering the inlet conditions in an iterative fashion with the Big Dish engine is very difficult, as timing changes cannot be deliberately altered without manufacturing specific components and disassembling the engine head each time to install them.

Secondly, Bannister considers that increasing the expansion ratio beyond some threshold value may actually reduce the engines efficiency due to the increased compression work done on the wet vapour that is not fully exhausted by the time the exhaust port has closed.

Whilst none of Bannister’s concerns could be directly addressed during the course of the investigation it is hoped that by the results of this investigation, future researchers will be able to accurately map the internal engine conditions over a continuous spectrum of intake conditions.

McIntosh’s variable displacement method involved installing a valve in the steam line to the number one engine cylinder and closing the valve during periods were steam inlet conditions were not optimal; such as during morning start-ups.

As noted by McIntosh (2008, p.9) in his report the variable displacement method can be used to increase the thermal efficiency of a Rankine cycle engine. However, this terminology is a little ambiguous as the mass flow through the system is set at the feed pump and constant during periods of steady state operation. Therefore this method of “variable displacement” does not to relate to variations in the rate of mass displacement, rather this is related to a change in the rate of thermal throughput by varying the rate of volumetric displacement. Thus this method is similar to the one employed by Bannister, in that both reduce volumetric flow across the expansion process which leads to increased back pressure to the receiver and higher steam temperatures at inlet for a set rate of solar radiation input.

McIntosh concluded that unfortunately the improvements gained from the increased steam inlet enthalpy were not significant enough to likely see the use of this form of variable displacement in the long term (McIntosh 2008, p.ii). 17

Both McIntosh and Bannister were able to achieve slight gains in efficiency in their investigations, but neither was really able to optimise the operation of the engine directly from data obtained across a detailed range of rates of expansion.

As the current conditions of the Big Dish engine required attention to be paid to mechanical components involved in governing steam inlet into the engine, an attempt has also been made in this report to propose a new actuating mechanism for the inlet valves which is introduced in 0. However, first the engine conditions and a clear understanding of the forces involved on the current bash pin actuating mechanism must be investigated to determine why these components are failing.

2.6 Current Engine Conditions

There is not a lot of information on the design of the steam engine other than many, but not all, of the original drawings submitted to the workshop for construction of the steam engine. This has lead to the need for a large amount of time to be invested during the course of the project to determine the critical dimension of components directly related to the inlet of the engine.

Several of the internal engine components have deteriorated to the extent that they need to be replaced in the current working engine, these include the valve lift pins and the copper head gaskets.

Drawings obtained for the dimensions of the valve bash pins did not appear to match the quantity of materials remaining in the deformed pins and replacement components could not be found immediately. The original drawing found for the lift pins and the piston crown in which they are held is presented in Appendix A: .

Whilst new pins were eventually found these were determined to have a vertical height that was too large. This incorrect length caused interference between the valve ball-bearing and the valve guide plate that prevented rotation of the entire engine assembly! As the dimensions of these unused pins were identical to those found on the workshop drawings it can only be concluded that at some stage a calculation error occurred previously.

One further problem associated with the evaluation of the worn components within the engine is that the total hours of operation could not be determined. Two workshop log books were located and it seems that between 2002 and 2010 the engine had only operated for a total of 128 hours. This made the evaluation of the material listed (BӦHLER S600 see Appendix E: ) on the lift pin drawing difficult as the magnitude of deformation could not accurately be assessed against any time dependant benchmarks.

Figure 2-8: Disassembly of Steam Engine Head Assembly

2.6.1 Valve Actuator Pins

As mentioned, one of the major problems during this project was determining the actual vertical dimension of the valve lift pins that have worn to the extent that they now need replacing. All of the bash pins removed from the first cylinder of the engine had very similar failure characteristics suggesting that the cause of the problems may have been systematic; although this can’t categorically be said until the pins in the other three cylinders have been observed.

Figure 2-9: Deterioration of Valve Lift Pins

Figure 2-9 displays the state of the valve bash pins in the number one cylinder once exposed. The extent to which these pins have deformed is a direct result of the harsh temperatures, pressures and stresses under which they have operated. Aspects that have lead to failure in the current pin design are discussed in more detail in Chapter 3.

The magnitude of the wear on the valve pins when compared to the dimensions of the original drawing found for the part suggested that the outer pins had lost 2.3mm in height whilst the central pin had lost 0.8mm during operation. At the time this was discovered, this amount of wear seemed excessive and the variation in wear between the two types could not be reconciled. Ironically, as will be seen in Section 3.4 the maximum height that the pins can be installed into the engine without causing interference between the valve ball-bearings and the valve guide plate are almost exactly the same as the deformed height of the extracted bash pins.

2.6.2 Head Gaskets

As expected the annealed head gaskets were compressed significantly during tensioning of the head bolt and their thickness was half the original 1mm dimension. Dimensions for an appropriate head gasket size were determined and drawings of these are presented in Appendix H: . Gaskets were subsequently manufactured using 0.9mm copper sheet as a stack of twenty using a CNC wire cutter to achieve an accurate bore clearance.

Figure 2-10: Crankcase Gasket Manufacturing

There was also no evidence of drawings for gaskets situated between the crankcase and the cylinders, although a tin template was found indicating that these may have been produced. Given that the cylinders on the complete engine were not going to be removed and head gaskets were already being produced it was decided that crankcase gaskets should be manufactured and installed on the spare engine.

2.6.3 Valve Ball-bearing and Valve Seats

The valve ball-bearings and valve seat display no evidence of significant deterioration. There are two reasons why this would probably be the case:

 The valve bash pin has a lower hardness than the hardened stainless steel ball-bearing and thus act as a sacrificial component when peak forces occur.  As the valve pins retract into the cylinder the ball-bearing remains in contact with the top of the bash pins due to gravity and fluid drag forces. This may result in the severity of the force exerted on the valve, by the valve seat, being reduced as the ball- bearing’s momentum suddenly disappears.

Therefore, at present there does not appear to be any need to alter or investigated these aspects of the current engine design.

2.6.4 Cylinders

As part of the process of inspecting the condition of the inlet system within the Lister steam engine the head plates of the engine were removed. Again there was little information on the methodology used in the original development of the engine.

In addition to the removal of the head plates an estimate of the torque used to tighten the bolts needed to be conducted. Measurements derived for the torques of the head bolts from the use of a lever arm and spring gage are displayed in Error! Reference source not found.. The 21 maximum torque required to release one of the head bolt nuts was around 100N. Subsequent verification of this figure was achieved by re-tensioning the remaining untouched head bolts at this level and only a slight amount of twist (if any) occurred before the torque wrench ratchet disengaged.

Head Bolt Torque Calculations Lever Arm Length (m): 0.9

Force Spring Gage Readings: (lbs) (kg) (N) Torque (Nm) 1 18 8.16 80.10 72.09 2 16 7.26 71.20 64.08 3 21 9.53 93.44 84.10 4 20.5 9.30 91.22 82.10 5 25 11.34 111.24 100.12 6 19 8.62 84.55 76.09

MAX 25 11.34 111.24 100.12

Table 2-3: Determination of Head Bolt Torque Measurements with a Spring Gage

Unfortunately, inspection of the cylinder liners appeared to indicate vertical scuff marks periodically spaced around the circumference of the cylinder liners. Measurement of the liner bore using tri point callipers did not suggest that the bore of the cylinders were now oversized and as the removal of the cylinders to inspect the piston skirt was not advised, no further investigation into the cylinder wear was undertaken on the complete engine.

Values for the size of the spare cylinder bores and piston diameters have been recorded and are include in Appendix H: . From the measurements made it appeared that the press fit liners had a marginally smaller diameter in the middle of the cylinder, probably resulting from this region being under the most compressive strain subsequent to being shrunk fit into the undersize cylinder diameter. This slight radius of curvature along the liner length meant that whilst the majority of the measurements were within the maximum and minimum values of 3.8725” and 3.8767” respectively (Detroit Diesel Engines 1980), the diameter at the centre of the measured cylinders were usually a little undersized.

During the measurement of the spare piston liners it was notices that the oil gallery holes (See Appendix Figure H-1) that were drilled in a radial pattern around the liners before they were press fit into the cylinders had burs remaining from the drills that had been used to make the holes. It was determined that this would probably be the cause of the vertical scuff marks noted earlier, as the piston skirts would initially been marked and rubbing unevenly against 22 the liner surface. By hand polishing the liners using 300 Grit wet and dry abrasive paper the burs were removed from the liners surface. Damage to the cylinder liners in the manner described may have lead to a decrease in the mechanical seal formed by the piston rings against the liners.

Section 2.2 covered a lot of theory in relation to the thermodynamic cycle involved in the Big Dish engine, within which a brief discussion of how the expansion ratio efficiency can be derived from the Otto air standard cycle equations. Much work has gone into the determination and optimisation of the expansion ratios for the two models of Lister conversion constructed ANU (See Section 2.5). However as this report is the most recent and probably most the pedantic coverage of the Big Dish engine since Bannister (1991), the magnitude of the current expansion ratio was explored again to ensure that its use in computational models was not based on erroneous information in future. Siangukone (2005, p.56) reports the expansion ratio of the current engine to be 1:10.8 from data on the measurements of the engine internals that he had been given; almost exactly the expansion ratio that Bannister (1991, p.173) had recommended at the conclusion of his PHD Thesis.

However, during the course of this project all of the customised components of the Big Dish engine were modelled in Solidworks to the best of the information available and within reasonable error margins and tolerances (including component wear), a new measure of the expansion ratio was determined and is presented in Error! Reference source not found.. Some of the heavy lifting in this case was performed using Solidworks to obtain volumes for complex regions such as underneath the valve plate to determine a value of around 1:12.1

Figure 2-11: Solidworks Volume Calculations of Complex Volume Beneath Valve Plate

Component Name mm mm^3 Valve Plate Cavity 27120.43

Distance Between Crown and Cavity (A) 5.68 Cylinder Bore (B) 98.4 Clearance Volume = (A*π*B^2)/4 (a) 43194.5

Outer Valve Pin Volume Above Crown (b) 811.5 Centre Valve Pin Volume Above Crown (c) 809.73

Outer diameter ( C ) 98.35 Inner Diameter (D) 77.5 Height (E) 3.5 Cavity Around Piston Crown = E*((π*C^2)/4 - (π*D^2)/4) (d) 10078.74

Total Volume at TDC = (a) - 2*(b) - ( c ) + (d) 77960.94

Stroke (F) 114.3 Cylinder Bore (B) 98.4 Stroke Additional Volume = (F*π*B^2)/4 (a) 869213.2

Expansion Ratio (Vfin/Vini) 12.14934

Table 2-4: Expansion Ratio Calculations

Clearly the expansion ratio is only an indicative measure and the returns in obtaining increasing accuracy in this area are diminishing rapidly. However as a result of engine inlet timing being directly coupled to the speed of the motor prior studies on engine efficiency were somewhat dependent on quantifying a measurement that in all reality changes very slightly over the life of an engine as components wear. If only the current engine had some mechanism for control over the engine inlet with reference to the crank angle instead……

2.7 Summary

Thermodynamics theory indicates that efficient operation of many thermodynamic processes depends on control of the system between states. ANU’s HL4 steam conversion has previously demonstrated that it is capable of operating reasonably well when the level of solar thermal power delivered is sufficient.

Despite, these successes the engine has also got some serious weak links in its design. These are not only related to the mechanical failure of components, they also relate to the best feature of the engine; it’s simplicity. Unfortunately the basic bash pin design constrains to a large degree control of the inlet angle or any aspect of engine timing.

It is possible that an alternative may be capable of resolving both the issues of bash valve deterioration and inlet timing efficiencies. Most probably this would involve the installation of a mechanical or electromagnetically actuated valve that has the required timing capabilities.

Chapter 3 Investigation into Valve Pin Failure

3.1 Introduction

Prior to determining whether another material or actuator design might be more appropriate than the current valve bash pins, an understanding of why the current configuration has undergone what appears to be systematic failure, needs to be obtained.

3.2 Pin Failure Characteristics

Due to the size of the bash pins, presenting close up images of the damage they have encountered is difficult. Figure 3-1 is a representative drawing of the measured dimensions of the worn pins and the damage that resulted from their service life.

Figure 3-1: Cylinder #1 Bash Pin Damage

The pin tips have clear evidence of significant plastic deformation through the mushrooming effect at their tops. Intermittent cracks are also present although they cannot be seen in the in photographs (Figure 2-9) there are multiple on each pin tip. As a result of the bash pin tips having a definite edge rather than being machined to have a radius, high stress concentrations occur in this region. It is likely that as a result of these high stress concentrations and the strain hardening in the material under cyclic compressive loading that cracks fractures began to appear around the lip of the pin.

These cracks would then progress inwards as cyclic loading continues, resulting in the mushrooming of the component tip. Progressive Fracture or Fatigue is related to the Fatigue Strength (Boresi and Schmidt 2003) of a material which is considered to be the stress at which a material can handle cyclic loading without the affects of fatigue failure. Increasing the Fracture Strength alone will not be able to resolve the problem of peak stresses causing cracks to develop in the material.

There is also evidence of corrosion that can be seen on the top of the crown and around the surface area exposed of the alloy steel pins. The blackening of the surface of these pins occurs as a result of oxidisation at high temperatures even though the material has alloying elements that increase its corrosion resistance (Section 4.1). This form of corrosion in alloys is also seen when materials are intentionally steam treated at temperature in the range of 340˚C to 650˚C to increase atmospheric corrosion resistance (Budinski and Budinski 2005); for example the blackened appearance of some pistols and other firearms.

Early indications of local buckling can be seen at base of two of the valve pins. This is of particular concern and the state of the uninspected cylinder valve pins should be viewed prior to the engine being operated again.

The valve pin material has exhibited the ability to yield minimally at high stresses within the engine without completely failing apart. This would indicate that the BӦHLER S600 is quite durable, or capable one absorbing a large amount of strain energy between the yield limit and ultimate limit of the material. However this characteristic of the material may be significant;y different if the heat treatments are applied.

Whilst this durability has saved the cylinder bores from valve pin debris falling freely into the cylinder, the indications of local buckling could cause major damage if left unchecked.

3.3 Valve Mechanical Operation

This section of the report seeks to gain approximate measures for the peak stresses encountered as a result of a combination of load and geometric considerations. Peak stress approximations are calculated using equations to determine the maximum principle compressive and shear stress resulting from contact between one or more curved surfaces

(Boresi and Schmidt 2003) whilst buckling considerations are determined using Euler’s formula for columns with one end fastened.

3.3.1 Peak Contact Stresses

Figure 3-2 presents a representation of the contact surface between the valve ball-bearing and the valve lift pin. For simplicity the stress analysis is carried out as though the forces were applied statically. It should be noted that the actual forces may be somewhat larger as a result of the momentum of the piston assembly when the interaction occurs between the surfaces.

Assumptions included in the model are that the materials obeys Hooke’s Law and that each material is isotropic and homogeneous, though not necessarily both the same material (Boresi and Schmidt 2003).

Figure 3-2: Pin-Valve Contact Coordinates

Geometric and load characteristics are presented in Error! Reference source not found. with the normal force between the pin and ball being calculated by resolving the pressure differences across the ball-bearing just prior to the moment of contact between the two (Appendix B.2 ).

The magnitude of pressure on the inlet side of the ball-bearing was obtained from Lovegrove et al. (2009) report in which it stated that “produces steam that is superheated at up to 500˚C at 4.5Mpa”. It was then decided to use a value of 5.00MPa during the report for conservatism.

The pressure in the cylinder after recompression of the wet vapour that was not exhausted entirely, was obtained from McIntosh’s (2008, p.36) work. The recompression pressure prior

28 to the opening of the inlet valve determined by McIntosh was approximately 800kPa as can be seen in Figure 2-2. This recompression pressure value is highly dependent on the system conditions at the time of measurement and are expected to vary greatly according to parameters such as the mass flow through the circuit. Again in the spirit of conservative analysis it was decide that a 200kPa recompression pressure would be used in order to ensure that no underestimation occurred during the analysis.

Name Nomenclature Value Ball bearing radius of curvature 5.5mm

Pin face radius of curvature

Angle between normal forces 0 Radians Modulus of Elasticity 217GPa Poisson’s Ratio 0.29 Force P 463N

Table 3-1: Lift Pin and Ball-bearing Properties

One further simplification was required because the exact class of stainless steel and hardening methods used to produce the ball bearings was not readily available. By referring to a material properties guide, it was discovered that the majority of stainless steels appeared to have a Young’s Modulus of around 200GPa. This value is very close to the value given by BӦHLER for their S600 alloy of 217GPa. Thus it was decided that using equations that assume both materials are the same is not an unacceptable departure from the characteristics of the situation.

The general method employed in the determination of the principle stresses for the area of contact between any two curved surfaces involves the solutions of all the points at coordinates (x,y) for which d in Equation 3-1 equals zero with constants A and B that are related to the Radii of Curvature in coordinates (x,y) for each object . Thus the general solution to determine the stresses resulting from the forces over the point of contact involves an integral over the area for which the distance between the deformed faces is zero. The derivation of the general equations used is not included here as it is rather long and complex, please consult Boresi and Schmidt (2003, p.589) for a complete explanation of the development of these equations.

Equation 3-1: Approximate Distance Between Two Curved Surfaces at Coordinates (x,y)

Equation 3-2: Root Solution B to Eq:3-1

Equation 3-3: Root Solution A to Eq:3-1

For a sphere contacting a plane the equations for B and A simplify giving the result for our pin and ball-bearing contact in Equation 3-5. Where the radius of curvature for the valve pin head is considered to be infinite whilst for the ball-bearing it is the radius of the sphere; which is equal in both the x and y axis.

Equation 3-4: Contact Between a Sphere and Plane

Equation 3-5: Ratio of Solutions to the Quadratic Equation in Eq:3.1

Where E and ν represent the Young’s Modulus and Poisson’s ratio respectively for the material under consideration in Equation 3-6.

Equation 3-6: Constant of Proportionality, Delta

From the value obtained in Equation 3-5 we can determine the stress and deflection coefficients from Figure B-2 in Appendix B: . The values obtained from the figure are

30 presented in Error! Reference source not found. from which the principle stresses and strain are determined.

Coefficient Value

2.2 1.00

0.90

0.65 0.47

0.21

Table 3-2: Stress and Deflection Coefficients for Bash Pin-Valve Contact

Equation 3-7: Semiminor Axis of Area of Contact

Equation 3-8: Maximum Principle Compressive Stress

Equation 3-9: Maximum Shear Stress at a Distance zs from the original surface of the Material

Equation 3-10: Point of Maximum Stress Below the Surface of the Original Contact

Equation 3-11: Total Deflection between the Two Bodies

Equation 3-8 to Equation 3-11 predict the maximum compressive stress, maximum shear stress magnitude and position relative to the plane of contact and the total displacement of the two contact faces relative to one another. Interestingly the force experienced at the contact point of the lift pin and ball-bearing is predicted by this model to be 150 times the equivalent stress on the face of the lift pin if the force was applied uniformly across it.

By way of a reverse checking the formula results given by Boresi and Schmidt’ equations (2003, p.589-608), we can calculate the equivalent area of the ball-bearing face in contact with the lift pin if we assume that all of the point load deformation occurs within the ball- bearing.

Figure 3-3: Ball-bearing Deformed Face Area Equivalent

Equation 3-12: Radius of an Equivalent Area of Deformation within the Ball-bearing

Equation 3-13: Equivalent Contact Area on Ball-bearing Face

Equation 3-14: Stress Encountered on the Equivalent Surface Area of the Deformed Valve

Equation 3-14 predicts a value of around one third of that predicted by the Boresi and Schmidt’s (2003, p.589-608) equations. Whilst it is a little presumptuous to assume that the check used here is valid, this approach seems to add credibility to the result that the stresses resulting from the point contact are probably in the order of Giga Pascals. 32

The equations used in this section to gain an understanding of the peak forces that result from bash pin actuation of the valve are quite technical and require a sufficient level of knowledge to understand the concepts correctly. No guarantee can ever be given that an idealised model can accurately predict the behaviour of complex physical systems. However, it seems that these equations indicate that the stresses at the valve pin are at an order of magnitude that would cause most non-alloy steels to yield.

As for whether the values obtained in the analysis of peak stresses indicate that the BӦHLER S600 steel has insufficient strength to operate under these conditions the best evidence is the deterioration of the pins themselves.

3.3.2 Buckling Critical Loads

An estimate of the critical force that will cause the piston lift pins to buckle upon application can be determined using Euler’s Formula. From the pressure balance conducted in Appendix B.2 we are able to determine an approximate value for the critical load that may cause buckling of the valve bash pins.

Equation 3-15: Euler’s Formula

By rearranging Equation 3-15 we can calculate the radius of a valve pin made from the same material as those currently in the engine with a safety factor of 2.5 for the 464N load with the Modulus of Elasticity for BӦHLER S600 being 217GPa (BӦHLER UDDEHOLM 2005).

Equation 3-16: Critical Buckling Moment of Inertia

Equation 3-17: Moment of Inertia for a Circular Cross-Section

From Equation 3-16 and Equation 3-17, with le being an equivalent length equal to two times the length of the valve pins above the piston crown (24.7mm) when installed, we can calculate the radius of a similar pin that will not buckle under the normal load of the valve. Results indicated that failure of this material would probably occur for a circular cross section with a radius of 1.14mm at the anticipated loads; well below the pin radius of 2.5mm. Thus even 33 though the valve pins appear to have suffered from local buckling these calculations would indicate that the materials is capable of withstanding the required forces.

3.4 Valve Bash Pin Geometric Analysis and Results

Many methods were attempted to try and determine within a reasonable level of accuracy the original length of the worn valve pins. Ultimately the only method that was successful proved to be direct measurement of the maximum valve for valve lift relative to the pins at TDC.

The actual processes attempted are not all that exciting and little is to be gained by going in to detail here. Some tables of values calculated early on in the process are included in Appendix A: . It is recommended though that if the same problems occur in the future, by direct measurement of the valve lift pin top relative to the top of the valve plate at TDC, followed then by measuring the height of ball-bearing surface protruding above the underside of an inverted valve guide plate, a good measure of whether the installed pins need more material can be achieved; using the upper face of the valve plate and low face of the valve guide plate as a reference as the two are flush.

The maximum safe valve travel appears to be around 2.8mm at TDC. This corresponds to an outer pin height of around 40.9mm for the outer pins and 36.05 for the centre pins with maximum valve lifts at around 2.7mm for both the inner and outer pins respectively.

Subsequent to the determination of appropriate valve pin heights a batch of twenty new pins were produced to for the construction of the spare engine and in anticipation of the need for the replacement of pins in the uninspected engine.

Figure 3-4: Freshly Machined Valve Pins and Brass Test Piece

The most interesting development taken from the determination of the pin height’s, was that the pins that had been removed from the assembled were almost exactly the same height as the calculated value. This leaves open the possibility that at some stage these pins were installed with their length’s too long. Had this occurred it would be remarkable if the only damage that resulted was the compression of the pins to a point where the excess height was deformed out in a similar fashion to that seen in Figure 2-9. However, if this scenario was the case then there is the chance that the material used to construct the pins is appropriate and can with stand the stresses; in which case this investigation may all have been in vain and someone has a lot of explaining to do.

Chapter 4 Valve Pin Material Analysis

4.1 Introduction

It is believed from the drawings obtained for the valve bash pins and the statements made by Robert Whelan (See Appendix C: Professor Keith Lovegrove) that the material used for the pins was BӦHLER S600 High Speed Steel. This steel is advertised by BӦHLER as a tool steel and the alloy can be heat treated to achieve high strength and toughness.

Although it should be noted that the high toughness values for the S600 alloy steel is usually achieved after heat treatment has occurred. Values for Rockwell-C hardness tests of the used pin material and new bar obtained from BӦHLER both had similarly low values for the material hardness at between 15 and 20.

Elements within the steel include Manganese, Chromium, Molybdenum, Vanadium and Tungsten (Appendix E: . Most of these elements are included to increase the hardnenability of the alloy, whilst Chromium also serves to increase corrosion resistance and strength at high temperatures (Budinski and Budinski 2005).

The UNS standard number for the S600 is T11302 (BӦHLER UDDEHOLM 2005) for which unsurprisingly the ~T designates a tool steel with the standard quality of the material having a Hardness Value of 630 and a Machinability Index of 340, making the alloy able to achieve very high hardness and consequently making it relatively difficult to machine to high tolerances (Budinski and Budinski 2005).

4.2 Material Selection Parameters

To assess the material currently being used against other prospective candidates the Cambridge Engineering Selector computer package was used. Parameters were developed against which materials that were not suited to the harsh conditions within the engine could be excluded. These criteria included:

 High Endurance Limit; As discussed the valve pins are subjected to cyclic loading at 25 cycles per second. Thus a high endurance limit is critical to the reliability of the component.

 High Young’s Modulus; A slight amount of elastic deformation is usually a good things, and helps materials to absorb the energy in the load force, but in the case of the valve pins minimal strain is seen as advantageous as this will serve to open the valve as wide as it can be during inlet.

 High Fracture Toughness; required to arrest fractures if they occur in the material.

 Very Good Oxidation Resistance at 500˚C; All materials that are used in connection with superheated steam must exhibit this characteristic.

 Low Thermal Expansion Coefficient; for the reasons outlined in the argument above for the Young’s Modulus of the material.

 High Hardness; Required for wear resistance.

 High Compressive Strength; Valve pin loading is primarily through compression and as such a high compressive strength is imperative.

 Low to Moderate Price.

4.3 Material Selection

Figure 4-1: Modulus of Elasticity and Endurance Limit Criteria Selection

The high requirement for the Modulus of Elasticity and more importantly the endurance limit excluded the majority of materials listed in the CES database as shown in Figure 4-1.

Figure 4-2: High Fracture Toughness and Oxidation at 500˚C

Oxidation above 500˚C and relatively high fracture toughness reduced the set of possible candidates even further.

Figure 4-3: Thermal Expansion and Hardness

Figure 4-4: High Compressive Strength an Low Price

Figure 4-5: Final Results of the CES Materials

Error! Reference source not found. to Figure 4-5 presents the process of elimination graphically on the CES computer program. Values for the selection criteria are presented below in Error! Reference source not found. and were guided by the results obtained in Section 3.3.

Stage Criteria >/< Value 1 Young's Modulus > 0.55GPa 1 Endurance Limit > 27MPa 2 Fracture Toughness > 10MPa(m^0.5) 2 Oxidation at 500˚C Very Good 3 Hardness > 10.0MPa 3 Thermal Expansion < 6.00E-06 4 Price < $50 Compressive 2.00GPa 4 Strength >

Table 4-1: CES Parameter Values

It appears that the only commercially available material in the CES database that has the characteristics required to withstand the conditions within the engines operating conditions is Tungsten. Tungsten is considered to have the highest hardness, melting point and modulus of elasticity of all metals p.(Budinski and Budinski 2005, p.698). As a result Tungsten has a high machinabilty index which is a measure of the relative difficulty to machine the material; for which high values indicate expensive and time consuming production methods.

Due to the tight tolerances required in the production of the valve lift pins and resulting cost for the part to be produced if Tungsten alloys were used instead it is unlikely that this would be a feasible option.

These results to a degree vindicate the use of BӦHLER’s S600 by the engine developers as this alloy has around 6.4% Tungsten (Appendix F: ) content by weight and would be considerably easier to machine in this form than if Tungsten itself was the predominant element in the alloy.

4.4 Conclusion

The Cambridge Engineering Selection software is a good tool to use for an initial indication of which elements would be best suited to the characteristics required but not too much significance should be attached to the results.

The software is a slightly older version and it is known that CES have had problems attracting suppliers to list the materials on the system. Thus the results can sometimes be a little narrow and not a complete representation of all the options available.

What is probably the most significant outcome from the results is that they confirm that the use of BӦHLER S600 with a tungsten content of 6.4% Tungsten to be a reasonable choice. 40

Chapter 5 Valve Actuator Concept Review

5.1 Introduction

Solenoids valves are implemented within many different devices and environments, from washing machines to diesel fuel injection. The principles behind their operation are well known but the design and testing required in the construction of actual solenoids for specific applications, is quite complex and beyond the scope of this project.

Therefore the analysis of the initial solenoid concept for the steam engine is constrained to a first estimate of the size, voltage and current requirements for a coil and core that are capable of providing the required actuation force to open the inlet valve over the valve lift requirements.

The majority of the problems encountered by the lift pins on the Lister steam engine result from the high stresses due to repetitive point loading as the pins come in contact with the ball bearings. It is hoped a solenoid valve may be a viable solution to the mechanical wear that is occurring at the tops of the valve pins.

5.2 Current Bash Pin Design Issues and Potential Solutions

The Lister Diesel steam engine uses ball-bearings seated within an enclosed region to form a seal across the valve seat by way of a pressure gradient across the valve plate itself. The processes and stresses involved in the operation of the bash pin valves is cover in Section 3.3.

The issue of pin wear is most probably the result of high peak contact stresses at the point when the ball-bearing and the bash pin surfaces collide. A survey of materials through a software database using important material property requirements to find a suitable material provided only one material as the solution.

Tungsten is known to be potentially toxic and can be difficult to machine to tight tolerances due the high hardness values; which can be seen through the relative machinability index of 520 (Budinski and Budinski 2005, p.834). Thus, as the only material identified within the materials selection process that could achieve the required reliability, it does not appear that the conditions within the engine are conducive to the bash pin concept; no matter which material is chosen.

In order to alleviate the high peak stresses within the valve operation two things need to be changed in the design. The first is that the contact between the ball-bearing and any actuating shaft needs to be maximised so that the peak stresses encountered are kept to a minimum. The second is that the relative velocities of the ball-bearing and the actuating pin should if possible 41 be zero on the same axis as the centre of force application to reduce the additional stresses between the two resulting from material momentum changes.

With these two suggestions in mind ideal solutions to this contact geometry problem would come in the form of a camshaft and follower configuration commonly seen in many reciprocating engines or the use of a valve in which the actuating shaft itself forms the seal with the valve seat.

There is evidence that the designers of the HL4 steam engine considered building a mechanically actuated steam inlet valve that would be controlled by either a custom timing gear linked to the crankshaft or by reinstalling the factory camshaft of the HL4 engine. Only drawings for such a valve have been found and it is not known whether the idea progressed into the prototyping phase.

As the reinstallation of the factory camshaft and designing the timing gear would be quite time consuming and costly it is not expected that approval for such a radical change to the current engine design would be given. In addition using direct gear linkages between a camshaft and valve actuator does not automatically come with the ability to easily vary inlet angle on the engine. Thus any form of mechanically driven actuation method is unlikely to have the added benefit of variable inlet timing from which researchers could assess the optimal performance characteristics of the motor.

5.3 Solenoid Actuation as a Potential Replacement

It may be possible for a solenoid valve to achieve the required reduction in material stress during periods of actuation as well as the ability to have variable inlet timing.

Peak stresses between the ball-bearing and actuating shaft can be minimised by increasing the contact surface area. By eliminating the need for a two piece valve/actuator and combining the two into a single shaft the issues of internal stress are then dominated by the materials cross-section and shape stress concentration factors. Furthermore the peak loads would be reduced by having an actuating shaft that does not have any velocity relative to the actual valve.

It is with these potential benefits in mind that a solenoid valve arrangement on the current HL4 steam engine components is presented in Figure 5-1. The idea is that due to budgetary restrictions on the development of the engine itself there are very limited resources available for the implementation of a solenoid valve. By using old and damaged HL4 engine components and making slight modifications, such as machining holes in the steam chest plate and valve guide plate, a configuration similar to the in the picture may be possible at

42 limited expense. This concept is not without several issues though, such as increased heat losses as it dissipates through the solenoid body into the environment.

Figure 5-1: Steam Engine Solenoid Concept

5.3.1 Solenoid Operation and Advantages

As mentioned the imagined solenoid valve solution has several perceived advantages which include the ability to map the engine inlet angle through a computer program and reduced peak stresses resulting through the removal of momentum energy transfer between the lift pin and piston. Other potential advantages with using a solenoid rather than the current bash pin include:

 The reduction of risk that bash pin material or one of the crown bolt comes loose within the engine cylinder during operation, which would undoubtedly cause major damage to the piston, cylinder line and valve plate.  Elimination of the need to manufacture piston crowns, and the top of the piston itself.

5.3.2 Solenoid Disadvantages

Disadvantages associated with the solenoid concept include:

 Increased heat dissipation through the solenoid coil body. 43

 The electrical energy consumption required to operate the solenoid.  Market availability and cost.  Solenoid failure through overheating; this is the most common form of solenoid failure.

5.3.3 Solenoid Valve Equations and Optimisation

The calculations used to conduct a first assessment into the power and size requirements for a solenoid capable of lifting the valve ball-bearings during operation, largely follows the design equations given by Say (1964, 22-5). The method employed calculates the required Amp- turns for a solenoid given the required force on the plunger and construction of the core, casing and stroke length.

Our aim is to determine whether a relatively small solenoid can produce enough force to break open the valves of the steam engine under a pressure gradient of around 5MPa. Factors such as the required current, voltage and subsequently power must also be considered for various configurations to determine whether the implementation will result in large losses to the efficiency of the engine as a whole.

Many of the constraints incorporated in the design process are presented in Error! Reference source not found.. The force requirement is a result of the pressure gradient across the valve ball-bearing, whilst the voltage was chosen to be 24V because it is the same potential used by the motor control systems within the engine room as well as for safety, cost and accreditation requirements. The majority of the other constraints are based on the dimension of the steam chest and position of the valves within the engine.

Parameter Value Force 463.5N Voltage 24V Plunger Diameter 15mm Plunger Air Gap 3mm Coil Height 80mm Wire Diameter 0.8mm Max Coil Temperature 450DegC Throat Height 20mm

Table 5-1: Constraints on Solenoid Design

Minimising the amount of air gap within a solenoid is of great importance, as the relative permeability of air is unity and tends to make up the largest proportion of required Amp-turns in the system. Valve displacement during inlet is currently at a maximum of 2.7mm therefore 44 the plunger air gap was set to 3mm with a plunger diameter and the required force known the solution to the flux density across the cross-sectional area of the air gap can be calculated by Equation 5-1. By dividing both sides of Equation 5-2 by the air gap area and substituting in the value for the flux density from Equation 5-1 as shown by Equation 5-3 for a given air gap -7 length, and knowing the permeability of air is μ0=4*π*10 , we can calculate the number of amp turns required to produce the required flux density across the plunger air gap.

Equation 5-1: Pull Force between a Magnetised Coil and Ferromagnetic Core Plunger

Equation 5-2: Magnetic Flux Lines

Equation 5-3: Flux Density

The total number of amp turns for the coils is the sum of all the individual requirements for each component within of the coil flux path as shown in Equation 5-4.

Equation 5-4: Total Ampere Turns Required

The width of the coil winding within a solenoid as displayed in Figure 5-2 can be calculated using an approximation for the space-factor within the coil windings of 0.5 (Say 1964, p.22- 27), along with the coil current density and a given coil height as in Equation 5-5 and Equation 5-6.

Equation 5-5: Current Density through a Conductor

Equation 5-6: Approximate Coil Winding Width

Figure 5-2: Dimension of a Solenoid Valve (Say 1964, p.22-27)

Once the dimensions in Figure 5-2 have been obtained the total length of wire used in the coil can be calculated by multiplying the diameter of the coil be the number of windings, these values then lead to the calculation of voltage required in the coil.

Using Copper for the coil windings means the resistivity (ρ) in Equation 5-7 would be -8 ρ0=1.72*10 Ωm at 20˚C (Young and Freedman 2004, p.948). In the case of the solenoid being within close proximity to the engine, increases in conductor resistance needs to be included in the calculations. To calculate the resistivity of the wire at the maximum anticipated operating temperature of 450˚C we need the Temperature Coefficient of Resistivity, α=0.00393(˚C)-1 (Young and Freedman 2004, p.949).

Once the total Amp turns have been calculated for each of the elements within the solenoid coil, at a given current we can easily calculate the Voltage and Power loss associated with the required plunger force through Equation 5-7 and Equation 5-9.

Equation 5-7: Voltage In Terms of Current and Conductor

Equation 5-8: Relative Resistivity at a Temperature with Reference to a Datum Temp

Equation 5-9: Power Consumed by Conductor

Figure 5-3 displays the Voltage and Power consumption for the over a range of currents up to 200A whilst Figure 5-4 displays the total coil thickness for the copper wire windings and the plunger core. The equations used to generate these figure in the Matlab software program are included in Appendix C: .

The voltage and coil diameter both decay rapidly with increasing current but the power consumption of the coil increases linearly with any increase in current. Thus there is a trade- off between obtaining a desired voltage level and coil size, with the need to keep power consumption down.

Given that the solenoid idea was conceived to facilitate increases in efficiency by varying the inlet timing of the steam engine, power requirements should be kept to a minimum. Even though we can expect an increase in engine output resulting from the removal of the bash pins, this is likely to be offset entirely by the electrical power requirement for a solenoid employed on the engine.

Figure 5-3: Coil Voltage and Power Consumption as a Function of Current

Figure 5-3 would appear to indicate that for the forces, dimensional and voltage requirements of the solenoid a current should be chosen in the range of 20-40A, beyond this value the power consumption for all the solenoid valves on the engine cylinders would be in excess of 1kW. If we assume that the engine is running at 50kW the use this particular solenoid would result in an output loss of around 2%. Whilst this loss may appear to be minor in the context of an engine that struggles to achieve 15% thermal efficiency, losses in excess of 1kW should be avoided if possible.

The solenoid diameter is constrained by the need for one solenoid coil per valve. The area available on the top of the steam chambers is not sufficient to have three shorter style solenoids with larger diameters. Thus it was decided to set one of the constraints in the analysis to solenoid coils 80mm in height.

Again there is a trade-off between the possible coil diameters and the current flow through the coil. Figure 5-4 has the upper bound of coil widths set to 25mm which explains why the graph does not include the results for coil diameters above this value. As the equations used to determine the graphs do not include the Amp turns required for any other solenoid components other than the wire and core, it is expected that an appropriate coil should be less than 20mm total diameter to allow some area for the coil case that will be required to withstand the pressures involved in the steam engine.

Figure 5-4: Coil Total Diameter Including the Plunger and Winding Thickness

For the constraints listed in Error! Reference source not found. and the considerations mentioned above regarding power losses and the special requirements of an appropriate material to fill as the coil core a current value of around 40A appeared to be the best selection.

Results Values Coil Current 40A Voltage 24V Power Consumed 960.3W Coil and Plunger Diameter 19mm

Table 5-2: Optimised Solenoid Coil Values

5.3.4 Summary

There remains a lot of work to be done in analysing whether the bash pin actuation method could be feasibly replaced by a solenoid valve. Although it would appear that the use of an 80mm high solenoid coil with 5.7m of copper wire windings running 40A and 24V could provide the required force to open the inlet valve at pressures in the order of 5MPa.

Considerations such as the materials and geometry used to construct the case and wire diameter still need to be optimised along with determining the best material for the plunger

49 and how the plunger is to be connected to the valve stem if these are made separately rather than an one piece machined out on a lathe. In addition the costs of constructing the described solenoid valve needs to be considered along with ways to minimise the power consumption of the unit.

In addition there are complex electromechanical interactions between each solenoid coil that will have to be considered along with these effects on the surrounding control systems.

The benefits to engine reliability and the increased ability of research staff to experiment on the best inlet timing may out weight the costs associated, both financial and for net engine power output, by installing a solenoid valve to control valve actuation rather than by continuing to use the bash pin actuation method. Furthermore, the Big Dish engine design enables testing the use of a prototype solenoid setup periodically as the head plates are relatively easy to remove and reinstall.

Chapter 6 Conclusions and Further Work

By evaluating the wear characteristics and stresses involved in the deterioration of the present valve bash pin design, it can be concluded that it is not likely that there will be a feasible solution to the problem in terms of material selection alone.

The pins extracted from an assembled engine that has only operated for less than 200 hours in the last eight years show significant signs of mechanical failure. Peak compressive stress at the point of contact between the valve ball-bearing and actuating bash pin are likely to be around 1GPa to 3GPa.

Materials selection through the Cambridge Engineering Database software indicated that only Tungsten and high content Tungsten alloys could be capable of enduring the conditions within the engine. Unfortunately Tungsten is moderately expansive to purchase and very difficult to machine making replacing the existing pins with this metal a reasonably expensive option. Furthermore this may have the unintended consequence of causing the hardened stainless steel ball-bearings to wear.

Simply choosing a substitute material would not be addressing the more fundamental issues involved in the HL4 conversion and its use in a Rankine cycle power station. The greatest asset of the engine is it simplicity, however this also leads to major issues related to reliability and efficiency.

Having wear components within the cylinder that are susceptible to failure may result in significant and more costly damage to the pistons and cylinder liners. This can also be said of using bolts to secure the piston crown to the top of the piston; should one bolt come free during operation.

By using a solenoid coil such as the one recommended in Chapter 6 the issues of reliability as well as fundamental engine efficiency can be addressed concurrently.

Thus it suggested that if the recommendations of this report are accepted further work can be done in evaluating the optimal configuration of the solenoid coil and control systems.

However if the decision to continue using the engine only in its current configuration is chosen then it is recommended that BӦHLER S600 continues to be used as the material from which to construct the valve lift pins. Although it is recommended that this be accompanied by periodic evaluation of the condition of the pins so that better understanding of the wear and serviceable life can be understood.

Appendix A: Worn Bash Pin Dimensional Analysis

Figure A-1: Deformation Models for Worn Valve Pins

Pin Vertical Dimensions Valve Pin Model Name Vertical Dimension (mm) Whelan Outer Pin 43 Worn Outer Pin 40.7

Whelan Inner Pin 36.5 Worn Inner Pin 35.7

Table A-1: Vertical Dimensions from Worn Components and Available Sources

Outer Valve Pin Valve Pin Model Name Volume (mm^3) Whelan Design 2277.03 Worn Pin Upper Bound 2273.37 Worn Pin with Ball Bearing Indentation 2260.14

Table A-2: CAD Modelling Volume Comparisons for Outer Valve Pin

Centre Valve Pin Valve Pin Model Name Volume (mm^3) Whelan Design 1999.06 Worn Pin Upper Bound 2024.12 Worn Pin with Ball Bearing Indentation 2004.61 Worn Pin - Preferred Model 2012.74

Table A-3: CAD Modelling Volume Comparisons for Centre Valve Pin

Bash Pin Vertical Dimension Upper Bound Valve Pin Model Name Volume (mm^3) Height (mm) Outer Pin 2260.14 42.2 Centre Pin 2012.74 36.9

Table A-4: Effective Undeformed Bash Pin Height

Appendix B: Bash Pin Material Analysis

Figure B-1: Copy of Piston Crown and Lift Pin Drawing with Incorrect Lift Pin Heights B.1 Stress and Deflection Coefficient

Figure B-2: Stress and Deflection Coefficients for Two Bodies in Contact in a Point (Boresi and Schmidt 2003) B.2 Valve Inlet Pressure Balance

In order to determine the total net force on the valve ball-bearing at the instant prior to the valve bash pin opening the inlet valve a hydrostatic force balance was taken. The value for the inlet pressure was obtained from Bannister’s (1991, p.11) thesis, with the maximum value obtained in his report (6.00MPa) being used for conservatism. The pressure in the cylinder after recompression of the wet vapour that was not exhausted entirely, was obtained from McIntosh’s (2008, p.36) work. The recompression pressure prior to the opening of the inlet valve determined by McIntosh was approximately 800kPa as can be seen in Figure 2-2.

Figure B-3 presents a FBD for the pressure balance across the exposed surface projections for the ball-bearing valves in the Lister HL4 steam engine.

Figure B-3: Pressure Differential Across Ball-bearing Prior to Inlet

From Figure B-3 the resultant forces on each projected surface of the ball-bearing can be calculated using Equation B-1 to Equation B-3.

Equation B-1: Area exposed to Hydrostatic Force

Equation B-2: Resultant Forces due to Hydrostatic Pressure

Equation B-3: Net Force on the Valve Pin

Valve Ball-bearing Resultant Force Surface Diameter (m) Area (m^2) Pressure (Pa) (N) Inlet Side 0.011 9.50E-05 5.00E+06 4.75E+02 Cylinder Side 0.0086 5.81E-05 2.00E+05 1.16E+01

Net Force (N) -463.55 Table B-1 presents the results of the hydrostatic force balance across the inlet valve ball- bearing at the instant preceding the contact between the top face of the engine bash pin and the ball-bearing.

Valve Ball-bearing Resultant Force Surface Diameter (m) Area (m^2) Pressure (Pa) (N) Inlet Side 0.011 9.50E-05 5.00E+06 4.75E+02 Cylinder Side 0.0086 5.81E-05 2.00E+05 1.16E+01

Net Force (N) -463.55

Table B-1: Resultant Force on Valve Ball-bearing due to Pressure Differential

The net resultant force just prior to the inlet valve opening is around 524N and directed towards the low pressure cylinder side of the valve as expected. This force is equivalent to around the weight of a 53kg mass under Earth’s gravity. It was also determined that under this load the mass of the stainless steel ball-bearing was insignificant to the total value and was thus neglected.

Appendix C: Solenoid Calculations

Appendix D: Consultations Conducted

The following notes are on interviews conducted in order to obtain information on the history of the ANU Lister diesel conversion and solutions to the issues encountered during the process of the project. These details are presented in bullet form and are paraphrased from brief notes taken during each consultation.

Interview #1 Name: Mr Robert Gresham Credentials: Chief Technical Officer, ANU School of Engineering Date: 30th April 2010

Introduction: Rob Gresham has worked at The Australian National University for many years and was approach due to his previous involvement in drafting at the Research School of Physics and Engineering. During this time Rob in that area Rob was exposed to the development of the original HR3 and HL4 engines.

 Failure modes of the occurred during the development phase of the steam engines included:  Valve ball-bearings bouncing within the valve guide slots when engaged by the bash pins.  Valve pin plastic failure and cracking (similar to the conditions noted in Section 2.6.1 of this report)  Piston cracking at the crown plate weld (this method of piston crown construction in different to that seen in the current engine as bolts now hold the crown plate to the top of a machined recess in the piston.  Valve pin relative heights may have been staged or stepped.  Basic design constraints at the time of development:  Basic technical design to assist in installation and maintenance in the isolated community of White Cliffs; where the initial HR3 design was deployed as part of a NSW government initiative to provide local power supply to the town.

Interview #2 Name: Dr Keith Lovegrove Credentials: Associate Professor, ANU School of Engineering Date: 11th May 2010

Introduction: Dr Lovegrove is the leader of the ANU Solar Thermal Group for which this project was conducted (and the author’s supervisor). Keith has been involved in aspects of renewable energy research at the ANU for several decade and has intimate knowledge of the Big Dish solar thermal power plant. Much of the information Keith was able to contribute during this meeting related to the problems associated with the limited information on the 59 initial design of the Lister steam engines; as Keith had recently had a chance encounter with Bob Whelan, one of the original engine designers  The lift pins were not originally a press fit into the piston crowns.  The material used to construct the pins was probably that listed in Figure B-1.  Ensure the valve pins are not too large. If the pins are too high they will force the valve ball-bearings into the top of the valve guide plate and destroy either the pin, ball-bearing or both.  The radius at the base of the pins needs to be machined at very high precision so as not to cause stress concentrations in the material.

Interview #3 Name: Mr Ben Nash Credentials: Workshop Manager, ANU School of Engineering Date: 12th May 2010

Introduction: Ben Nash has worked as a workshop technician at ANU for over 15 years and is currently the workshop manager for the School of Engineering. Unfortunately due to the construction of a new workshop for the School of Engineering Ben was not able to provide as much help as he would have liked. Never the less his assistance was invaluable in the construction of new valve lift pins and gaskets for a spare Lister diesel engine conversion.

 High Speed Tool steels are general very malleable prior to heat treatment and the construction of the lift pins should not be very difficult on a CNC lathe once the G- Code is generated correctly.  The lift pins should not be hardened too much if at all. Under the high temperature, pressure and cyclic nature of the conditions in which the pins suffer high compressive forces they will probably shatter. However, by induction hardening the tips of the pins the rate of deformation appearing on the tops of the pins may be reduced.

Interview #4 Name: Dr Zbigniew Stachurski Credentials: Director (CSEM), ANU School of Engineering Date: 15th September 2010

Introduction: Dr Stachurski speciality is related to his research on composite materials and mechanics of materials in general. Zbigniew was very kind to provide some of his time and acquired knowledge in relation to the material used in the manufacturing of the lift pins.

 Under the operating conditions of the engine and if the pins are constructed from the BӦHLER S600 material, hardening the material is very likely to cause brittle fracture of the pins.  The BӦHLER S600 material may have some inherent hardness prior being subjected to the conditions within the steam engine. At these temperatures and pressures the 60

material may be experiencing a slight and slow diffusion which could possibly be softening the metal  Ball-bearing deterioration should be avoided or the engine valve will most likely not seal.  The constituent elements within the steel alloy would provide the material with corrosion resistance.  The material being use is probably a good choice and it is unlikely that a cheap alternative will be able to be found.

Interview #5 Name: Mr Greg Burgess Credentials: Research Officer, ANU School of Engineering Date: 21st September 2010

Introduction: Greg Burgess works primarily on the ANU Big Dish engine system and has been involved in the project since 1999. Greg has a lot of technical knowledge regarding electronics and is in charge of a lot of the control systems for the Big Dish.  The SG3 Big Dish had a lot of technical faults towards the end of its operation and as a result was not operational on a continued based between after 2002.  Logs show that the engine had only operated for 128 hours between 2002 and 2010. There is no guarantee that the engine hour clock was zeroed when the engine was constructed and there can be no way of knowing exactly how many hours the engine had operated since the SG3 dish was complete in 1994.  STG are going to be replacing the currently motor control system with a near new PLC (programmable logic controller). This new controller should easily be able to control any solenoid valve inlet control system on the engine if one was installed. The new controller system loop cycle takes 1.1ms and has inputs and outputs for digital signals as well as inputs for temperature, voltage and current.

Appendix E: Manufacturing Cost Estimates

Appendix F: BӦHLER S600 Material Specifications

Appendix G: G-code for Manufacturing

G.1 CNC Lathe G-code

G.1.1 Centre Valve Lift Pin % O0000

N100 G21 N120 G0 T0101 N110 X30. Z50. N130 G97 S1910 M03 N140 G0 G54 X20. Z0. N150 G50 S3600 N160 G96 S120 N170 G99 G1 X-1. F.1 N180 G0 Z2. N190 G50 S2000 N200 X14.822 N210 Z4.7 N220 G1 Z2.7 F.12 N230 Z-21.3 N240 X16. N250 X18.828 Z-19.886 N260 G0 Z4.7 N270 X13.644 N280 G1 Z2.7 N290 Z-21.3 N300 X15.222 N310 X18.05 Z-19.886 N320 G0 Z4.7 N330 X12.466 N340 G1 Z2.7 N350 Z-21.3 N360 X14.044 N370 X16.872 Z-19.886 N380 G0 Z4.7 N390 X11.288 N400 G1 Z2.7 N410 Z-21.3 N420 X12.866 N430 X15.694 Z-19.886

N440 G0 Z4.7 N450 X10.11 N460 G1 Z2.7 N470 Z-21.3 N480 X11.688 N490 X14.516 Z-19.886 N500 G0 Z4.7 N510 X8.932 N520 G1 Z2.7 N530 Z-21.3 N540 X10.51 N550 X13.338 Z-19.886 N560 G0 Z4.7 N570 X7.754 N580 G1 Z2.7 N590 Z-20.627 N600 X7.832 Z-21.3 N610 X9.332 N620 X12.16 Z-19.886 N630 G0 Z4.7 N640 X6.576 N650 G1 Z2.7 N660 Z-10.496 N670 X7.832 Z-21.3 N680 X8.154 N690 X10.982 Z-19.886 N700 G0 Z4.7 N710 X5.398 N720 G1 Z2.7 N730 Z-.365 N740 X6.976 Z-13.936 N750 X9.804 Z-12.522 N760 G0 X14.15 N770 Z-17. N780 G1 Z-19. N790 Z-30.7 N800 X14.2 N810 G3 X15.4 Z-31.3 R.6 N820 G1 Z-36.3 N830 X18.228 Z-34.886 N840 G0 Z-17. N850 X12.9 N860 G1 Z-19. 73

N870 Z-30.7 N880 X14.2 N890 G3 X14.55 Z-30.726 R.6 N900 G1 X17.378 Z-29.312 N910 G0 Z-17. N920 X11.65 N930 G1 Z-19. N940 Z-30.7 N950 X13.3 N960 X16.128 Z-29.286 N970 G0 Z-17. N980 X10.4 N990 G1 Z-19. N1000 Z-30.7 N1010 X12.05 N1020 X14.878 Z-29.286 N1030 G0 Z-17. N1040 X9.15 N1050 G1 Z-19. N1060 Z-30.7 N1070 X10.8 N1080 X13.628 Z-29.286 N1090 G0 Z-17. N1100 X7.9 N1110 G1 Z-19. N1120 Z-30.3 N1130 G2 X8.7 Z-30.7 R.4 N1140 G1 X9.55 N1150 X12.378 Z-29.286 N1160 G0 Z1.635 N1170 X5.398 N1180 G1 Z-.365 N1190 X7.9 Z-21.883 N1200 Z-30.3 N1210 G2 X8.7 Z-30.7 R.4 N1220 G1 X14.2 N1230 G3 X15.4 Z-31.3 R.6 N1240 G1 Z-40.9 N1250 X18.228 Z-39.486 N1260 G0 Z1.623 N1270 X4.999 N1280 G1 Z-.377 N1290 X7.5 Z-21.888 74

N1300 Z-30.3 N1310 G2 X8.7 Z-30.9 R.6 N1320 G1 X14.2 N1330 G3 X15. Z-31.3 R.4 N1340 G1 Z-40.9 N1350 X17.828 Z-39.486 N1370 T0100 N1380 M01 (TOOL - 3 OFFSET - 3) (LATHE TOOL 3 INSERT - NONE) N1400 G0 T0303 N1410 G97 S1731 M03 N1420 G0 G54 X22.068 Z-38.872 N1430 G50 S2000 N1440 G96 S120 N1450 X18.068 N1460 G1 X-.4 F.05 N1470 X3.6 N1480 G00 X18.068 N1500 T0300 N1510 M30 %

G.1.2 Outer Valve Pin % O0000 N100 G21 N120 G0 T0101 N130 G97 S2000 M03 N140 G0 G54 X17. Z0. N150 G50 S2000 N160 G96 S120 N170 G99 G1 X-.8 F.12 N180 G0 Z2. N190 G50 S3600 N200 X14.12 N210 Z4.7 N220 G1 Z2.7 F.2 N230 Z-35.54 N240 X14.2 N250 G3 X15.4 Z-36.14 I0. K-.6 75

N260 G1 Z-40.548 N270 X18.228 Z-39.134 N280 G0 Z4.7 N290 X12.24 N300 G1 Z2.7 N310 Z-35.54 N320 X14.2 N330 G3 X14.52 Z-35.562 I0. K-.6 N340 G1 X17.348 Z-34.148 N350 G0 Z4.7 N360 X10.36 N370 G1 Z2.7 N380 Z-35.54 N390 X12.64 N400 X15.468 Z-34.126 N410 G0 Z4.7 N420 X8.48 N430 G1 Z2.7 N440 Z-35.525 N450 G2 X8.7 Z-35.54 I.11 K.385 N460 G1 X10.76 N470 X13.588 Z-34.126 N480 G0 Z4.7 N490 X6.6 N500 G1 Z2.7 N510 Z-10.703 N520 X7.9 Z-21.883 N530 Z-35.14 N540 G2 X8.7 Z-35.54 I.4 K0. N550 G1 X8.88 N560 X11.708 Z-34.126 N570 G0 Z1.629 N580 X5.198 N590 G1 Z-.371 F.1 N600 X7.7 Z-21.885 N610 Z-35.14 N620 G2 X8.7 Z-35.64 I.5 K0. N630 G1 X14.2 N640 G3 X15.2 Z-36.14 I0. K-.5 N650 G1 Z-43.74 N660 X18.028 Z-42.326 N670 G0 Z1.623 N680 X4.999 76

N690 G1 Z-.377 N700 X7.5 Z-21.888 N710 Z-35.14 N720 G2 X8.7 Z-35.74 I.6 K0. N730 G1 X14.2 N740 G3 X15. Z-36.14 I0. K-.4 N750 G1 Z-43.74 N760 X17.828 Z-42.326 N770 X30.411 Z4.717 N780 T0100 N790 M01 (TOOL - 3 OFFSET - 3) (LATHE TOOL 3 INSERT - NONE) N810 G0 T0303 N820 G97 S1272 M03 N830 G0 G54 X20.016 Z-43.764 N840 G50 S2000 N850 G96 S80 N860 G1 X16.016 F.1 N870 X-.4 N880 X3.6 N890 G0 X18.016 N900 G00X41.326 Z-40.525 N910 T0300 N920 M30 %

G.2 CNC Wire Cutter G-code

G.2.1 Copper Crankcase Gasket

Please note that this head gasket was machined in two parts. The first pass was for the inner diameter of the gasket which corresponds to the outer diameter of the cylinder liner. The second pass covers the bolt holes and the outer profile of the gasket which is meant to be the same as the profile of the cylinder head.

" ( N0001 PROGRAM NAME LINERDIAMETER ) ;" " N0100 G90 ;" " N0102 G54 ;" " N0104 G29 ;" " N0106 T94 ;" " N0112 C405 ;" " N0114 G41 H160 ;" " N0114 G00 X38. Y-70. ;" " N0120 G01 X37.8726 Y-64.7357 ;" " N0122 G02 X0. Y-75. I-37.873 J64.736 ;" " N0124 X-37.873 Y-64.736 I0. J75. ;" " N0126 X-51.619 Y-70.619 I-13.746 J13.117 ;" " N0128 X-64.735 Y-37.873 I0. J19. ;" " N0130 X-69.4 Y-28.437 I64.735 J37.873 ;" " N0132 G01 X-69.4 Y28.4362 ;" " N0134 G02 X-62.772 Y41.044 I69.4 J-28.436 ;" " N0136 X-34.607 Y66.538 I13.784 J13.077 ;" " N0138 X34.607 Y66.538 I34.607 J-66.538 ;" " N0140 X62.772 Y41.044 I14.381 J-12.417 ;" " N0142 X69.4 Y28.436 I-62.772 J-41.044 ;" " N0144 G01 X69.4 Y-28.437 ;" " N0146 G02 X64.735 Y-37.873 I-69.4 J28.437 ;" " N0148 X51.619 Y-70.619 I-13.116 J-13.746 ;" " N0150 X37.873 Y-64.736 I0. J19. ;" " N0152 G40 ;" " N0154 M02 ;"

" ( N0001 PROGRAM NAME OUTERPROFILE ) ;" " N0100 G90 ;" " N0102 G54 ;" " N0104 G29 ;" " N0106 T94 ;" " N0110 G90 X38. Y-70. ;" 78

" N0112 C405 ;" " N0114 G41 H160 ;" " N0118 G01 X37.8724 Y-64.736 ;" " N0120 X37.768 Y-64.7968 ;" " N0122 G02 X0. Y-75. I-37.768 J64.797 ;" " N0124 X-37.768 Y-64.797 I0. J75. ;" " N0126 G01 X-37.8724 Y-64.736 ;" " N0128 X-37.9563 Y-64.8229 ;" " N0130 G02 X-51.619 Y-70.619 I-13.663 J13.204 ;" " N0132 X-64.822 Y-37.957 I0. J19. ;" " N0134 G01 X-64.7356 Y-37.8728 ;" " N0136 X-64.7964 Y-37.7684 ;" " N0138 G02 X-69.4 Y-28.437 I64.796 J37.768 ;" " N0140 G01 X-69.4 Y28.4362 ;" " N0142 G02 X-62.838 Y40.943 I69.4 J-28.436 ;" " N0144 G01 X-62.7725 Y41.0439 ;" " N0146 X-62.855 Y41.132 ;" " N0148 G02 X-34.686 Y66.629 I13.867 J12.989 ;" " N0150 G01 X-34.6068 Y66.5382 ;" " N0152 X-34.4995 Y66.5938 ;" " N0154 G02 X34.5 Y66.594 I34.5 J-66.594 ;" " N0156 G01 X34.6068 Y66.5382 ;" " N0158 X34.6863 Y66.6291 ;" " N0160 G02 X62.855 Y41.132 I14.302 J-12.508 ;" " N0162 G01 X62.7725 Y41.0439 ;" " N0164 X62.8384 Y40.9427 ;" " N0166 G02 X69.4 Y28.436 I-62.838 J-40.943 ;" " N0168 G01 X69.4 Y-28.437 ;" " N0170 G02 X64.796 Y-37.768 I-69.4 J28.437 ;" " N0172 G01 X64.7356 Y-37.8728 ;" " N0174 X64.8225 Y-37.9567 ;" " N0176 G02 X51.619 Y-70.619 I-13.204 J-13.662 ;" " N0178 X37.956 Y-64.823 I0. J19. ;" " N0180 G01 X37.8724 Y-64.736 ;" " N0182 G40 G50 X37.8726 Y-67.7357 ;" " N0184 M02 ;"

G.2.2 Copper Head Gasket " ( N0001 PROGRAM NAME GASKET ) ;" " N0100 G90 ;" " N0102 G54 ;" 79

" N0104 G29 ;" " N0106 T94 ;" " N0110 G92 X0. Y0. ;" " N0112 C405 ;" " N0114 G41 H160 ;" " N0118 G01 X0. Y-48.3 ;" " N0120 G03 X0. Y48.3 I0. J48.3 ;" " N0122 X0. Y-48.3 I0. J-48.3 ;" " N0124 G01 G40 G50 X0. Y0. ;" " N0126 M00 ;" " N0128 G00 X60. Y0. ;" " N0132 M00 ;" " N0134 G41 ;" " N0138 G01 X56.4 Y0. ;" " N0140 G02 X0. Y-56.4 I-56.4 J0. ;" " N0142 X56.4 Y0. I0. J56.4 ;" " N0144 G01 G40 G50 X60. Y0. ;" " N0146 M02 ;"

Appendix H: Cylinder Bore and Piston Diameter

As mentioned in Section 2.6.4 the measurements of the cylinder liners and the piston diameters of the spare components selected for assembly as another complete Lister HL4 steam engine are presented below.

Figure H-1: Cylinder Liner Measurement Intervals

Figure H-2: Cylinder Liner Diameter Measurements for the Soon to be Assembled Engine

Designation Distance (mm) A 10 B 25.4 C 44.45 D 95.25 E 127 F 177.8 G 228.6

Table H-1: Distance Legend for Figure H-2

Appendix I: Engine Component Blueprints

These pages have been removed from this version of the report due to Commercial in Confidence Restrictions.

For a copy of this report that includes the drawings produced during the project please contact Dr Keith Lovegrove at The Australian National University, School of Mechanical Engineering.

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