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Study of the Feasibility and Energy Savings of Producing and Pre-Cooling Hydrogen with a 5-KW Ammonia Based Combined Pow...

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STUDY OF THE FEASIBILITY AND EN ERGY SAVINGS OF PRODUCING AND PRE-COOLING HYDROGEN WITH A 5-KW AMMONIA BASED COMBINED POWER/COOLING CYCLE By ROBERT JOSEPH REED A THESIS PRESENTED TO THE GRADUATE SCHOOL OF THE UNIVERSITY OF FLOR IDA IN PARTIAL FULFILLMENT OF THE REQUIREMENTS FOR THE DEGREE OF MASTER OF SCIENCE UNIVERSITY OF FLORIDA 2004

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Copyright 2004 by ROBERT JOSEPH REED

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iii ACKNOWLEDGMENTS First and foremost, I would like to thank my wife for her constant love and support during the pursuit of my degree. Her pati ence and understanding while I completed this thesis will be forever appreciate d. I would also like to tha nk my fellow graduate students for making our office an enjoyable work e nvironment and a place I looked forward going to everyday. I would like to thank my advisor, Dr. Herb ert (Skip) Ingley, fo r his guidance during my research efforts. Always willing to help, he provided much needed advice and knowledge; but he also allowed me to devel op my own ideas and solutions, providing a wonderful learning experience. I thank my committee memb ers Sherif A. Sherif, D. Yogi Goswami, and Herbert (Skip) Ingley for all of their support. Finally, I would like to thank my parent s for believing in me from the beginning and always encouraging me that I coul d accomplish anything I put my mind to.

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iv TABLE OF CONTENTS page ACKNOWLEDGMENTS.................................................................................................iii LIST OF TABLES............................................................................................................vii LIST OF FIGURES.........................................................................................................viii NOMENCLATURE............................................................................................................x ABSTRACT....................................................................................................................... xv CHAPTER 1 MOTIVATION................................................................................................................1 Current Energy Trends.................................................................................................1 Hydrogen as a Future Energy Carrier...........................................................................4 2 BACKGROUND AND THEORY..................................................................................6 Hydrogen as an Energy Carrier....................................................................................6 Characteristics.......................................................................................................6 Production Technologies.......................................................................................7 Storage Technologies............................................................................................9 Electrolysis of Water..................................................................................................13 Process Description.............................................................................................14 Energy and Efficiency.........................................................................................15 Electrolyzer Designs............................................................................................18 Hydrogen Liquefaction...............................................................................................20 Process Description.............................................................................................21 Isenthalpic vs. isentropic expansion.............................................................21 Ortho/para conversion..................................................................................24 Claude cycle........................................................................................................25 Ammonia-Water Combined Power/Cooling Cycle....................................................27 Process Description.............................................................................................28 Expander Design.................................................................................................29 Positive-displacement expanders.................................................................30 Turbo-machinery..........................................................................................30 Scroll compressor/expander.........................................................................31 5 kW Prototype....................................................................................................33

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v 3 ANALYSIS METHODOLOGIES.................................................................................35 Hydrogen Energy Requirements.................................................................................35 Electrolysis of Water...........................................................................................35 Hydrogen Liquefaction........................................................................................37 Ammonia-Water Combined Power/Cooling Cycle....................................................41 4 EXPERIMENTAL SETUP AND DESIGN..................................................................45 Scroll Machines as Expanders....................................................................................45 Testing Apparatus and Instrumentation......................................................................46 Experimental Methodology........................................................................................50 Procedure.............................................................................................................50 Data Analysis.......................................................................................................51 5 RESULTS AND DISCUSSION....................................................................................53 Hydrogen Production and Liquefaction......................................................................54 Electrolysis of Water...........................................................................................54 Hydrogen Liquefaction........................................................................................54 Ammonia-water Combined Cycle.......................................................................64 Scroll Expander Performance Study...........................................................................70 6 RECOMMENDATIONS...............................................................................................76 Analytical Study.........................................................................................................76 Scroll Expander Performance Test.............................................................................76 7 CONCLUSIONS............................................................................................................78 APPENDIX A COMPUTER PROGRAM FO R CYCLE SIMULATIONS.........................................80 Claude Cycle Simulation............................................................................................80 Thermodynamic Property Evaluation..................................................................80 Program Description............................................................................................81 Ammonia-Based Combined Power/Cooling Cycle Simulation.................................86 Thermodynamic Property Evaluation..................................................................87 Program Description............................................................................................87 B CYCLE SIMULATION OUTPUT...............................................................................99 Claude Cycle Simulation Results...............................................................................99 Combined Cycle Simulation Results........................................................................100

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vi C EXPERIMENTAL COMPONENT LIST...................................................................103 LIST OF REFERENCES.................................................................................................105 BIOGRAPHICAL SKETCH...........................................................................................108

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vii LIST OF TABLES Table page 2.1 Heating values of hydrogen and other common fuels at STP.......................................7 2.2 Projected hydrogen costs of various production methods............................................8 2.3 Mass and energy density of select fuels........................................................................9 2.4 Advantages and disadvantages of monopolar and bipolar electrolyzers....................19 5.1 Specific energy requirements of the IMET electrolyzer...........................................54 5.2 Claude cycle simulation results for e xpander isentropic efficiency variation............57 5.3 Claude cycle simulation results for comp ressor isentropic efficiency variation........59 5.4 Claude cycle simulation results for compressor inlet pressure variation....................60 5.5 Claude cycle simulation results for compressor inlet temperature variation..............62 5.6 Claude cycle performance parameters for normal and optimum configuration.........64 A.1 Critical properties and coefficients contained within the “gas.dat” file....................86

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viii LIST OF FIGURES Figure page 1.1 World energy consumption sinc e 1970 with projections to 2025.................................1 1.2 US energy consumption by sector in 2002...................................................................2 1.3 Foreign oil imported as a percentage of the total oil consumed in the U.S..................3 2.1 Hydrogen production tec hnologies by energy source...................................................8 2.2 Fuel and total weight of several hydrogen storage systems........................................10 2.3 Process diagram of a simple alkaline electrolyzer......................................................15 2.5 T-S diagram comparing isenth alpic and isentropic expansion...................................23 2.6 Claude cycle with liquid nitrogen pr e-cooling and ortho/ para catalyzation...............26 2.7 Combined cycle flow diagram....................................................................................28 2.8 Flow path of a single fluid pocket through a scroll compressor.................................32 3.1 T-S diagram of ideal liquefaction process..................................................................38 4.1 Sanden TRS-90 automotive scroll compressor and test stand....................................47 4.2 Piston compressor with in tegrated tank and regulator................................................48 4.3 Thermocouple locations and flow meter.....................................................................48 4.4 Pony brake and back pressure gauge and valve..........................................................49 4.5 View of expander pulley showing the brake pads used as frictional surfaces............50 5.1 Sample output showing the optimum expander mass flow ratio, xe...........................56 5.2 Specific liquid yield and expander mass flow ratio as functions of the expander efficiency..................................................................................................................57 5.3 Required liquid nitrogen vs. expander efficiency.......................................................58 5.4 Specific work vs. expander mass flow ratio for varied e..........................................58

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ix 5.5 Specific work vs. expander mass flow ratio for varied c..........................................59 5.6 Impact of compressor and expander efficiencies on Claude cycle FOM...................60 5.7 Effect of compressor inlet pressure on the specific work...........................................61 5.8 Liquid nitrogen requirement vs. compressor inlet temperature..................................62 5.9 Specific work requirement vs. compressor inlet temperature.....................................63 5.10 Comparison of inlet pr essure and temperature a ffect on the cycle FOM.................64 5.11 Mass flow rate depende nce on expander efficiency.................................................65 5.12 Pump work variation with expander efficiency........................................................66 5.13 Boiler heat input and absorber heat rejection vs. expander efficiency.....................67 5.14 Cycle cooling capacity as a function of expander efficiency...................................67 5.15 Cycle thermal efficiency vs expander efficiency......................................................68 5.16 Effect of trace amounts of water within in the expander inlet stream on cycle cooling capacity........................................................................................................69 5.17 Expander exhaust and dew point temper ature at several water concentrations........69 5.18 Repeatability analysis applied to shaft power output at 65 psig...............................70 5.19 Shaft power vs. rotational speed at 60, 70, and 80 psig inlet pressure.....................71 5.20 Scroll expander isentropic efficiency........................................................................72 5.21 Volumetric efficiency variati on with expander rotational speed..............................73 5.22 Expander exit temperature and rotational speed relationship...................................73 5.23 Comparison of optimum geometries of a scroll compressor (left) and expander (right)........................................................................................................................ 75

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x NOMENCLATURE A ampere [A] AC alternating current CHWS chilled water source CHWR chilled water return CWS cooling water source CWR cooling water return CO2 carbon dioxide COP coefficient of performance DC direct current E voltage [V] or energy transfer rate[Btu/hr or kW] F Faraday’s constant FOM figure of merit G Gibbs energy [Btu/lbm] GFR Gibbs free energy of reaction [Btu/lbm] H enthalpy [Btu/lbm] HHV higher heating value [Btu/lbm] HHWS heating hot water source HHWR heating hot water return HX heat exchanger IC internal combustion

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xi I.D. Inner diameter [in.] KOH potassium hydroxide L liquid LH2 liquid hydrogen LHV lower heating value [Btu/lbm] LN2 liquid nitrogen P pressure [psia] PV photovoltaic Q heat transfer rate [Btu/hr or kW] R mass specific gas constant [Btu/lbm-R] S entropy [Btu/lbm-R] SMR steam/methane reformation STP standard temperature and pressure T temperature [ R or F] V volts [V] or volumetric flow rate [cfm] volumetric flow rate [ft3/min or cfm] W work transfer rate [kW] X ammonia mass fraction cp isobaric heat capacity [Btu/R] d displacement [cm3/rev] eelectron g vapor h enthalpy [Btu/lbm] or hour [hr]

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xii m mass flow rate [lbm/hr] n number of electrons v specific volume [ft3/lbm] w specific work [kW/lbm] x mass flow ratio y liquid yield ratio z nitrogen requirement ratio Greek coefficient of thermal expansion heat exchanger effectiveness efficiency JT Joule-Thompson expansion coefficient s isentropic expansion coefficient density [lbm/ft3] rotational speed [rad/s] Subscripts C ortho/para conversion process CW cooling water Elec electrolyzer FW feed water H2 hydrogen N2 nitrogen NH3 ammonia vapor

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xiii P isobaric or pump T isothermal ab absorber act actual ad adiabatic c compressor cool cooling load e expander f liquid g electric generator h isenthalpic in expander gas inlet max maximum min minimum o standard conditions opt optimum out expander gas outlet rect rectifier s isentropic shaft expander pulley shaft strong high ammonia c oncentration stream th thermoneutral v volumetric

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xiv vg vapor generator weak low ammonia concentration stream wf working fluid

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xv Abstract of Thesis Presen ted to the Graduate School of the University of Florida in Partial Fulfillment of the Requirements for the Degree of Master of Science STUDY OF THE FEASIBILITY AND EN ERGY SAVINGS OF PRODUCING AND PRE-COOLING HYDROGEN WITH A 5-KW AMMONIA BASED COMBINED POWER/COOLING CYCLE By Robert Joseph Reed May 2005 Chair: H. A. (Skip) Ingley Major Department: Mechanic al and Aerospace Engineering This thesis presents the results of a study on hydrogen production and liquefaction and the feasibility of the 5-kW ammonia base d combined power/cooling cycle to energize these processes. Analytical models of the electrolysis, Claude liquefaction, and combined cycle processes are developed to study the effects of variable boundary conditions and component efficiencies on the hydrogen production rate and to determine the optimum operating conditions. Additionally a performance study is implemented to gauge the applicability of a scroll expander with the 5-kW combined cycle. This research is motivated by the current energy crisis and r ecent research efforts in the development of renewable energy-based hydr ogen production methods. Analytical models are adapted to com puter simulations that calculate the thermodynamic properties, heat and work inte ractions, and efficiencies of each system for variable boundary conditions and compone nt efficiencies. Data from these simulations are used to deduce the optimum configuration that results in the maximum

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xvi hydrogen production rate. The scroll expander pe rformance test was carried out with a common automotive air-conditioning scroll compressor arranged in an open-cycle configuration using air at vari able inlet pressures. Predic tions on its performance with ammonia were made based on the observed tr ends and by contrasti ng the properties of the two working fluids. The minimum specific energy required fo r electrolysis and liquefaction is 24.839 kW-h/lbm-H2 (54.76 kW-h/kg-H2) and 3.817 kW-h/lbm-H2 (8.41 kW-h/kg-H2), respectively, for a to tal of 28.656 kW-h/lbm-H2 (63.18 kW-h/kg-H2). With a 5-kW output from the combined cycle, the maxi mum liquid hydrogen production rate is 7.21 gallons (27.3 liters) per day. Experiment al measurements of the scroll expander’s performance show isentropic efficiencies of 15 to 20 percent with maximum power output of 0.368 Hp (0.274 kW) at 1460 RPM with an inlet pressure of 80 psig (653 kPa). Simulation results show pre-cooling the hydrogen prior to liquefaction does not reduce the specific energy consumption and, in fact, is detrimental to the thermal efficiency. Furthermore, pre ssurized electrolysis is found to be the most effective means of reducing the specific energy of liquefaction. The heat and work interactions of the combined cycle scale with the inverse of the expander efficiency. Additionally, isentropic expander efficiencies above 60% are required to extract any cooling from the cycle. The performance test proved that scroll tip l eakage is the major cause of poor expander performance. Improvements of the scroll de sign such as increasing the scroll wrap and introducing low-friction material s would significantly increase its efficiency and make it a suitable design for low-output applications.

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1 CHAPTER 1 MOTIVATION Current energy consumption and forecasted demand with regard to limited fossil fuel reserves is presented in this chapter to demonstrate the necessity fo r the conversion to a renewable resources-based global ener gy market. Economical, environmental, and political factors are addresse d as further motivation. The remainder of the discussion introduces hydrogen as a potential energy carrier for a renewable energy market. Current Energy Trends Approximately 85.7% of the world’s energy is currently supplied by fossil fuels, with crude oil making up 38.8% of that total. Global energy consumption is projected to increase 54% over the next 25 years (Ener gy Information Administration, 2004). 206.7 242.8 285.2 311.1 348.4 368.7 403.9 470.8 517.3 567.8 622.9 0 100 200 300 400 500 600 700 800 19701975198019851990199520012010201520202025 YearGlobal Energy Consumption (1015 BTU) Pro j ectedHistorical Figure 1.1. World energy consumption since 1970 with projections to 2025 (Energy Information Administration, 2004)

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2 This increased demand is being fed primarily from countries with rapidly industrializing and emerging economies such as India and Chin a. Proven oil reserv es are sufficient to satisfy this demand over the next 20 years, afte r which there is debate as to whether oil production will peak before 2030 or that continued technological progress and new oil discoveries will satisfy the demand we ll into this century (Ramsay, 2003). The economic effects of increasing energy demand on a limited supply are apparent today with peak 2004 oil prices near $50/barrel and average gas prices in the US near $2.00/gallon. As fossil fuel production peaks and inevitably begins to decline, and without other viable energy sources, pr ices will continue to escalate. Residential 21% Commercial 18% Industrial 34% Transportation 27% Figure 1.2. US energy consumption by s ector in 2002 (Energy Information Administration, 2003) Figure 1.2 gives an overview of how energy is consumed in the US economy. Industry is affected directly and indirectly by the cost of energy. The direct effect is to increase the cost of processing raw materials and production Fuel costs involved with transporting finished goods is the indirect effect. The natural response of industry to increasing cost is to slow production and/or reduce labor fo rces, thus slowing the entire economy.

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3 A number of adverse environmental phe nomena such as the greenhouse effect, air pollution, acid rain, and oil spills are attributed to the use of fossil fu els. The burning of all fossil fuels produces carbon dioxide, a greenhouse gas. The Energy Information Administration reports that carbon dioxide contributes over 84 % to the total of greenhouse gases emitted (Mirabal, 2003). Global warming is widely debated as an ongoing occurance, but if it were found to be so, carbon dioxide emissions would be the main cause. Another by-product of fossil fuel combustion in air is the formation of nitrogen oxides (NOx) that contribute to ozone depletion as well as smog formation. Complex fossil fuels, such as petroleum and coal may also contain sulfur, which form sulfides that can cause acid rain. These environmental factors and others mentioned contaminate water supplies, damage ecosystem s, and are related to the occurrence of many respiratory illnesses in humans. In 1985, the US imported 27.3% of the oil it co nsumed. Over the past 18 years, as shown in Figure 1.3, the U.S. dependence on fo reign oil has steadily increased to 56.1% and is projected to be 69.6% of th at consumed by 2025 (Energy Information Administration, 2003). 0 10 20 30 40 50 60 70 80 1965197519851995200520152025 YearPercent of Oil Imported HistoricalPro j ected Figure 1.3. Foreign oil imported as a percentage of the total oil consumed in the U.S.

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4 With greater dependence on foreign oil, the U.S. will be reliant on a stable Middle East, Russia, and South America. International crises such as those recently in Iraq and Venezuela will have a more significant im pact on oil prices as they do today. It is important that alternative energy s ources are developed today to deal with the issues of tomorrow. Current research initiatives around the world are focused on hydrogen as the fuel of the future. With the development of a hydrogen economy based on renewable resources, greenhouse gas emissi ons will be reduced, the economy will be more independent of oil prices, and foreign po licy will be less influenced by oil reserves. Hydrogen as a Future Energy Carrier In 2001, 20.4% of global energy consumpti on supported transportation; of which 96% was supplied by crude oil (Energy In formation Administration, 2003). By developing an alternative fuel for transporta tion, world oil consumption could be reduced by as much as 19.6%. Reducing oil consump tion likewise reduces greenhouse emissions and ozone depletion. Hydrogen holds promise as the fuel to achieve these goals because it can be produced from water using renewa ble energy sources and it burns clean; with water and heat as the only combustion pr oducts (NOx emissions are possible when burned in air). One of the barriers to the widespread use of renewable resources is the geographical limitation. For example, hydropower can only be utilized in areas where damns can be built and solar power is depende nt on incident sunlight which varies from region to region. Renewable en ergy technologies can be utili zed more efficiently and on a broader scale by constr ucting large capacity plants in re gions with prominent sources of energy. The energy can subsequently be c onverted to chemical energy by producing hydrogen, enabling delivery to a larger market.

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5 Governments around the world realize the pot ential of hydrogen as an alternative fuel. Many countries have adopted research initiatives in the production, storage, and utilization of hydrogen. The U.S. Department of Energy has recently announced plans to advance toward a hydrogen-based energy sy stem making fuel-cell-powered vehicles available by 2010. Industry is following suit as most major automobile manufactures have significant programs in place to develop fuel cell powered vehicles (Ramsay, 2003). Hydrogen is a safe and clean fuel that when produced using renewable energy is virtually pollution free. Hydr ogen also provides a means to convert from a fixed source of energy to one compatible with the needs of transportation. With further development of production and storage tec hnology, hydrogen can become th e primary source of fuel for the transportation sector and can help usher in the renewable energy era.

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6 CHAPTER 2 BACKGROUND AND THEORY This chapter introduces hydrogen as a potenti al fuel and presents a brief overview of hydrogen storage and production systems. An emphasis is placed on the transportation sector and renewable technol ogies to develop the impor tance of electrolysis and liquefaction in a hydrogen economy. Following the theory of electrolysis and hydrogen liquefaction, the ammonia-wate r combined cycle is introduced as a means of converting low-temperature energy sources into usable electricity to power both systems; and refrigeration to pre-cool hydrogen prior to liquefaction. The scroll compressor is introduced as a potential high-efficiency expa nder for use with the combined cycle as motivation for the current study. Hydrogen as an Energy Carrier Hydrogen is the simplest, most abundant element in the universe comprising 75% of all visible matter by mass (Flynn, 1997). Currently, the majority of the hydrogen produced in the U.S. is used as a chemical in a variety of commercial applications including ammonia production, hydrogenation of fats and oils, and methanol production (National Hydrogen Association, 2004). With the continuing depletion and increasing cost of fossil fuels, however, greater consid eration is being given to hydrogen as an alternative fuel. Characteristics Hydrogen has several characteristics that make it a desirable a lternative fuel for transportation:

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7 Highest energy content per unit mass of any known fuel (51,574 Btu/lbm) – hydrogen produces 2.7 times more energy per unit mass than gasoline when burned. Table 2.1. Heating values of hydroge n and other common fuels at STP Btu/lbmkJ/gBtu/lbmkJ/g Hydrogen60954141.7851574119.96 Methane2386155.52150050.01 Propane2165150.361977245.99 Gasoline2046447.61900344.2 Diesel2024947.11883143.8 Methanol974622.67856419.92 Fuel Higher Heating ValueLower Heating Value (Gater, 2001) Clean – combustion of hydrogen produces no carbon dioxide or sulfur emissions. When burned with oxygen, the only byproducts are water and heat. If burned in air, nitrogen oxides may be produced. Renewable – hydrogen can be produced by a variety of methods using renewable energy sources for a virtually limitle ss and pollution free fuel supply. Technologically compatible – in the 1920s, German engineer Rudolf Erren successfully converted IC engines to hydrogen burning engines (National Hydrogen Association, 2004). Hydrogen can also be reacted with oxygen in a fuel cell to produce electricity to drive a motor. Efficient utilization – hydrogen IC engines are about 25% efficient, fuel cells are 45-60% efficient; typical gasoline IC engines are 18-20% efficient (National Hydrogen Association, 2004). Hydrogen fuel cell powered vehicles can be up to three times more efficient than today’s gasoline engines. Production Technologies The U.S. currently produces 9 million tons or 3.2 trillion cubic feet (90 billion Nm3) of hydrogen per year. Of this am ount, 95% is produced by steam/methane reformation (SMR) (National Hydrogen A ssociation, 2004). SMR operates by reacting a natural gas feedstock with steam at high temperatures (700 – 925 C) to produce carbon monoxide and hydrogen. The carbon monoxide is then consumed in a water/gas shift reaction to create CO2 and additional hydrogen. Other hydr ogen production methods are

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8 outlined in Figure 2.1. Detailed descriptions of each fossil fuel based production technology are given by Mirabal (2003). Re newable energy systems are outlined by the U.S. Department of Energy (2003). Figure 2.1. Hydrogen production te chnologies by energy source SMR is currently the most cost e ffective method of producing hydrogen; however, because of increasing fossil fuel co st due to diminishing supplies and reduced capital cost of renewable energy due to t echnological improvement s, wind and ammoniawater combined power/refrigera tion cycle solar power based el ectrolysis are projected to become the most cost competitive by 2020 (Mirabal, 2003). Table 2.2. Projected hydrogen cost s of various production methods1 Year2003201020302050 Steam Methane Reformation0.660.902.759.88 Partial Oxidation0.800.901.442.89 Coal Gasification1.121.201.652.83 Electrolysis Grid Power (fossil fuel based)1.531.632.424.12 Electrolysis PV / Antenna Power3.472.400.910.65 Electrolysis Wind Power1.331.140.780.60 Electrolysis Ammonia Water Combined Cycle2.501.370.890.63 Hydrogen Production Costs ($/lb) 1 Original data converted from $/GJ using the HHV of hydrogen (Mirabal, 2003)

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9 Although there are other methods availabl e to produce hydrogen from renewable resources, electrolysis is the most versatile and technologically devel oped. Electrolyzers do not require high temperature for operati on as do thermal decomposition, dissociation, or chemical processes nor are they dependent exclusively on sunlight. For these reasons, electrolysis is expected to be the predom inate method of hydrogen production in a future hydrogen economy. Storage Technologies One of the barriers preventing the wide use of hydrogen as a fuel is its storage. This issue centers on hydrogen’s low density and correspondingly low energy density. Table 2.3 displays these characteristics for hydrogen under several conditions as well as for other common fuels. Table 2.3. Mass and energy dens ity of select fuels lb/ft 3 kg/m 3 Btu/ft 3 MJ/m 3 Hydrogen gas (STP)0.0053090.085044323.6012.06 gas (3,000 psig, 60 F)0.963115.42858,7052,187 gas (10,000 psig, 60 F)2.48439.797151,4345,643 liquid4.419770.798269,39810,038 Methane gas (STP)0.0423580.67851010.7037.66 gas (3,000 psig, 60 F)10.778172.650257,1749,583 liquid26.367422.367629,14323,442 Propane gas (STP)0.11831.8952561.7595.45 liquid36.298581.450785,88829,283 Gasoline (liquid)45.884735.010938,97634,987 Diesel (liquid)53.064850.0121,074,48340,036 Methanol (liquid)49.380791.012481,26017,932 DensityEnergy density Fuel (National Institute of Standa rds and Technology 2003, Chevron 1998) Because of its low density, hydrogen requires a large volume for an equivalent amount of stored energy as compared to other common fuels. To illustrate this fact, the energy

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10 equivalent of 10 gallons (37.85 liters) of gasoline would requi re a tank size of 175 gallons (662.4 liters) for gaseous hydrogen at 3000 psig and 37.6 gallons (142.3 liters) for liquid hydrogen. Another issue with hydrogen storage in regards to its use as a motor fuel is the combined weight of the container, safe ty equipment and any required insulation. Container weights (including fuel) for severa l hydrogen storage methods are given for an energy equivalent of 7.93 gallons (3 0 liters) of gasoline in Figure 2.2. 0500100015002000Gasoline Methanol Hydrogen: Metal hydride* Hydrogen: Gas at 3000 psig Hydrogen: Gas at 5000 psig Hydrogen: LiquidTotal Weight [lb] Fuel Container 18 lb fuel, 161 lb total 18 lb fuel, 450 lb 18 lb fuel, 630 lb total 18 lb fuel, 1700 lb total 108 lb fuel, 125 lb total 49 lb fuel, 60 lb total Figure 2.2. Fuel and total weight of several hydrogen storage systems. *Storage capacity by weight approximately 1.1%. There are several methods of hydrogen st orage currently available or being researched. They are summarized as follows: Metal hydrides. Metal hydrides are specific a lloys consisting primarily of granular magnesium, nickel, iron, and/or titanium. These all oys are capable of adsorbing hydrogen (1% 8% by weight) at high pressure and moderate temperature and releasing it under low pressure and elevated temperatur e. Metal hydrides are characterized by deadsorption temperature. Low-temperature (< 200 F) hydrides operate at higher pressures to prevent hydrogen release at ambient temperat ures. These hydrides typically adsorb 1

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11 2 percent of their weight in hydr ogen. Higher temperature (> 250 F) hydrides hold 5 – 10 percent hydrogen by weight, but require significant amounts of heat to attain the temperatures required to release the stored hydrogen. (Sunatech Inc., 2001). Metal hydrides provide the safest mean s of storing hydrogen. Because the hydrogen is stored in a solid-state media, it cannot be ignited until released. In addition, the hydrogen is released at low pressures and moderate temperatures; therefore, no specialized storage tank is requi red to deal with high pressures or cryogenic temperatures. Despite these advantages, metal hydrides ar e undesirable for use in transportation. Large, heavy, and costly stor age units are required to hold equivalent amounts of energy as current gasoline tanks, as shown in Fi gure 2.2. Common hydrogen impurities such as oxygen and water reduce the ab ility of the tank to st ore hydrogen as they bond permanently to the metal. Additionally, vibr ations due to typical driving conditions can result in particle attrition that al so reduces the tank’s useful life. Compressed hydrogen. Compressed hydrogen is the simplest and one of the most common methods of hydrogen storage and tran sportation. Even at 10,000 psig, however, compressed hydrogen contains nearly 8 times less energy per unit volume than gasoline (not including the energy expended in compre ssing the hydrogen). Cylinders tend to be heavy because of the robust construction n ecessary to withstand the high pressures and impacts. These factors make compressed hydr ogen storage suitable for only short ranged applications or as a reserve fuel for liquid hydrogen powered vehicles. Liquefied hydrogen. Liquid hydrogen is formed by cooling hydrogen gas to -423 F (-253 C) at atmospheric pressure. Storage of such low temperature fluids is achieved using a dual-walled cylinder with an evacu ated space between the cylinder walls

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12 (Dewer’s flask). Due to the relatively high surface to volume ratio typical of the small tanks used in transportation applications, additional multi-layered radiation insulation sheets are also employed (Flynn, 1997). There are several technological challenge s that must be overcome in order for liquefied hydrogen storage to come into widesp read use. First is safe tank design to reduce weight and hydrogen boil off due to heat infiltration. The imperfect insulation of the inner tank supports, among other factors, cau ses a typical boil off rate of 3% per day (Clean Energy Research Center, 2003). Furthermore, improved methods of hydrogen liquefaction must be developed to reduce LH2 cost. Today, about 30% of the energy contained in LH2 is consumed by the liquefaction process (Fuel Cell Store, 2003). Lastly, re-filling stations must be developed such that the public can operate them safely. Liquefied hydrogen (LH2) is currently the optimum hydrogen storage method for vehicles in terms of tank si ze/weight and energy density. LH2 has the highest volumetric energy capacity of any commercially availabl e storage system being only four times less than gasoline; and because hydrogen burns more efficiently than gasoline, LH2 tanks are not necessarily four times the size of typica l gasoline tanks for a given vehicle range. This allows automobile manufactures to c ontinue using current vehicle designs, easing the transition into a hydrogen economy. Carbon nanotubes and glass microspheres. Carbon nanotubes store hydrogen in microscopic surface pores and within the tube structures via adso rption. The mechanism by which they store and rele ase hydrogen is similar to me tal hydrides, however carbon nanotubes are lighter, cheaper, and are capable of storing 4.2 to 65% hydrogen by weight (Fuel Cell Store, 2003). Carbon nanotubes ar e still under research and development and

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13 currently store between one and ten percen t hydrogen by mass (Clean Energy Research Center, 2003). Glass microspheres are currently being rese arched as a potential hydrogen storage method. Hydrogen is stored by first warming the tiny glass to increase their surface permeability and then immersing them in high-pressure hydrogen gas. The spheres are then cooled, locking the hydrogen inside of th e glass balls. Increasi ng the temperature of the spheres reverses this process. Experi ments to increase hydrogen release rates by crushing the spheres are also being performed. The key advantage of glass microspheres is storage at ambient temperature. The technology exists today for the in troduction of hydrogen-powered vehicles; however, the size, weight, and/ or cost limitations imposed on storage systems by the low energy density of hydrogen must first be overc ome. Liquid hydrogen holds the greatest promise for hydrogen-powered vehicles. These storage systems have the lowest weight and volume of those commercially availa ble, and with improved tank design and hydrogen liquefaction methods, the relative ly high costs will lessen over time. Electrolysis of Water English scientists William Nicholson and Sir Anthony Carlisle first discovered that the application of an electric current to water produces hydrogen and oxygen in 1800. The principle of electrolysis was later form ulated by Michael Faraday in 1820. Since then electrolysis has played only a minor role in worldwide hydrogen production; recently contributing to only 4% of to tal global production (National Hydrogen Association, 2004). Current electrolytic hydrogen production is limited to low-cost electricity sources such as hydroelectric or small-scale ons ite generation in which purity is essential.

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14 The importance of electrolysis in a future hydrogen economy is two fold: First, as discussed previously, electr olysis powered by wind or the ammonia water combined power/cooling cycle is projected to be the most cost efficient hydrogen production method by 2020. Second, it provides a prac tical link between hydrogen and renewable resources through electricity generation. In this manner, electrolysis can indirectly utilize any energy source that can be used to produ ce electricity. Furthermore, when powered by electricity generated from renewable sour ces of energy, electroly sis does not require fossil fuels and has zero polluting emissions. Process Description Electrolysis is defined by McMurray and Fa y as the use of an electric current to drive a non-spontaneous chemical reaction (1998). Electrolys is of water consists of a pair of oxidation/reduction reactions driven by a DC voltage applied across two electrodes as described by equations 2.1a – 2.1c. Cathode: OH H e O H 2 2 22 2 (2.1a) Anode: e O H O OH 2 2 1 22 2 (2.1b) Overall: 2 2 22 1 O H O H (2.1c) Water is reduced at the cathode to form hydrogen gas and hydroxide ions (OH). The OH ions migrate toward the anode where they are oxidized to form oxygen, water, and two free electrons. The free electrons are then attracted to the positively charged cathode, thus completing the circuit. A schematic of a simple electrolyzer and the overall electrolysis process is given in Figure 2.3. Each electrode is isolated from the othe r with an ion-conducting diaphragm to keep the product gases separate; and an electrolyte is used to make the solution conductive.

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15 The electrolyte is chosen such that its reduc tion and oxidation potenti als are less than that of water. In this manner, the electrolyte is conserved because it acts only as an ionconducting substance. Figure 2.3. Process diagram of a simple alka line electrolyzer (ada pted from Mirabal) Energy and Efficiency The voltage required for reversible or isentropic electrolysis is proportional to Gibb’s free energy of reaction as defined by Faraday’s Law: nFE G (2.2) where G is Gibb’s free energy of reaction n is the number of electrons transferred in the reaction F is Faraday’s constant, mol Coulombs410 648531 9 E is the cell voltage A negative sign is included on the righ t hand side of Equation 2.2 because by convention voltage input is considered negative (McMurray and Fay, 1998). The spontaneity of a given reacti on is determined by the sign of the Gibbs free energy of reaction (from hereon referred to as GFR) GFR is positive for non-spontaneous reactions and negative for spontaneous ones. For water at standa rd temperature and pressure, (25 C and 1 atm), the GFR is 50,941 Btu/lbmH2 (14.93 kW-h/lbmH2) and the corresponding reversible voltage is 1.23 V. Th e electrical energy re quired to drive the

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16 electrolysis reaction is equa l to the GFR (Casper, 1978). Th e enthalpy of reaction (higher heating value) of hydrogen, however, is 61,451 Btu/lbm (18.01 kW-h/lbmH2). Conservation of energy dict ates that the remaining 10,510 Btu/lbm (48.89 kJ/mol) must be supplied as heat. For a reversible pro cess, this heat would be obtained from the surroundings, and the electrolyzer w ould double as a refrigeration unit. The second law of thermodynamics states th at entropy always incr eases for any real process. Entropy production in electrolysi s increases the requir ed cell voltage as described by equation 2.3. PT E nFT nFE S T G H (2.3) where p PT E nF T G S from Faraday’s Law The entropy produced is liberated as heat, which supplies the additional 10,510 Btu/lbm necessary to form hydrogen. The voltage require d for isothermal electrolysis (defined as the thermoneutral voltage) is 1.47V. Th is result is obtained by replacing G in Equation 2.2 by the HHV of hydrogen. In reality, the ther moneutral voltage is the lowest that can possibly be achieved. Real electrolyzers require greater than th e thermoneutral voltage due to additional overvoltages independent of the entropy ge neration. Overvoltage is defined as the difference between the applied voltage and the reversible 1.2 3V and is proportional to the amount of current passed through the cell (Cas per, 1978). These overvoltages include: ohmic resistance of the electrolyte, con centration polarization (changes in the concentration of H+ or O2+ or water near the electrodes) voltage gradients at the electrode/electrolyte interf ace due to the slowness of reaction (proportional to cell

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17 operating temperature), and wire and component resistance (typically about 2% of total loss) (Casper, 1978). The primary source of electrolyte resistance is the formation of vapor bubbles on the electrodes (Wendt, 1990). Additional energy losses occur (typically 5% of total energy consumption) within each subsystem including AC to DC rectification, cooling water system, feed water pumps, and electrolyzer pumps (if necessary) (Casper, 1978). The majority of electrolyzer manufacture rs have taken steps to reduce these overvoltages. Concentration polarization can be avoided by adequate mixing of the electrolyte through circulation or by natural gas lift. One method developed to reduce electrolyte resistance is zero gap cell geomet ry in which porous electrodes are pressed on either side of the diaphragm, forcing the product gases to leave fr om the rear (Wendt, 1990). Another technique is to increase the cell operating temperature and pressure in order to speed up reaction kinetics and reduce electrolyte resistance. However, this also enhances corrosion of th e electrodes and shortens operating lifetime. The figures of merit measuring elec trolyzer performance are current, electrochemical, and thermal efficiencies. Curr ent efficiency measures deviation from the hydrogen yield predicted by Faraday’s law at 1.47 V and 1000 A-h due to extraneous electrode reactions (Casper, 1978). For most electrolyzers, this number approaches 100%. Electrochemical efficien cy is defined as the reversible voltage divided by the operating voltage. The maximum electrochemi cal efficiency under isothermal conditions is 83.7%. Thermal (1st law) efficiency is the ratio of the isothermal voltage to the operating voltage or the HHV of hydrogen di vided by electricity input as given by Equation 2.4.

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18 Elec H act th ElecE HHV V V2 (2.4) Using this definition of effi ciency, ideal electrolysis operates at an apparent 120% efficiency. Thermal efficiency is the most widely used figure of merit by electrolyzer manufactures, therefore any given efficiency will be thermal effi ciency. Commercial electrolyzers currently operate at efficien cies (excluding subsystems) of up to 85% (Stuart Energy, 2004) Electrolyzer Designs Electrolyzers are typically classified by their electrolyte; the most common of which is alkaline/water (Casper, 1978). Othe rs include solid polymer (SPE), seawater, and solid oxide; descriptions of which are given by Casper (1978). Alkaline/water electrolyzers typically operate with a 30% potassium hydroxide (KOH) solution at relatively low temperatures of 158 – 212 F (70 – 100 C). There are two varieties of alkaline/wate r electrolyzer: monopolar (tan k-type) and bipolar (filterpress). A summary of each type highlight ing the unique advantages and drawbacks of each is given below: Monopolar or tank-type cells are constructed as an alte rnating set of anodes and cathodes connected electri cally in parallel and hung verti cally from gas collectors into a tank of electrolyte. Mixing of the electrolyt e is achieved through simple gas lift. The cathodes are normally surrounded by a diaphragm to prevent the mixing of gases. This arrangement results in indivi dual tanks operating at low voltages (typically 1.9 – 2.5 V) and high currents (Casper, 1978). Bi-polar or filter-press electrolyzers are characterized by the stacked design of the cells. In this configuration, one side of an electrode serves as the cathode and the other as

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19 the anode of an adjoining cell. Electrodes ar e connected in series such that a desired operating voltage is achieved by increasing the total number of cells. The geometries of these cells are relatively thin; therefore, a pump is required to circulate the electrolyte through the cells. Bi-polar cell s typically operate at lower cu rrent levels due to higher operating voltage. Table 2.4 lists the pros a nd cons of each alkaline/water electrolyzer design. Table 2.4. Advantages and disadvantages of monopolar and bipolar electrolyzers Advantages Disadvantages Monopolar Require relatively few, inexpensive parts Easily maintained – individual cells can be isolated for repair with minimum plant downtime No pumps required for electrolysis circulation Unable to operate at high temperatures because of heat loss from large surface areas Bulky design requires greater space per unit hydrogen produced Tanks are difficult to design for pressurized electrolysis Relatively high voltage losses and non-uniform current density distribution result from long current paths (Wendt) Bi-polar Compact design Capable of operating at high pressures and temperatures Lower ohmic resistance and energy losses Requires precise fabrication tolerances and additional gaskets due to sealing problems Maintenance is more difficult – if one cell fails, the entire cell must be shut-down and dismantled One of the largest electrolytic hydrogen production plants in North America was built by Cominco, Ltd. in British Columbia, Cana da. Before being shut down due to high electricity costs, the plant produced 41 t ons/day of hydrogen with 3,229 individual tank-

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20 type cells operating at 2.1 V (70% efficien t) (Casper, 1978). Since that time, most manufactures have adopted the more effi cient bi-polar design (Wendt, 1990). Hydrogen Liquefaction Liquid hydrogen was first produced by Jame s Dewar in 1898; however, up until the mid 1940s to mid 1950s it remained nothing more than a laboratory curiosity (Flynn, 1997). In the late 1950s, the US Air For ce began producing substantial amounts of LH2 for its top secret “Bear” Program. Under c ontract to the Air Force, Air Products and Chemicals, Inc., constructed three producti on plants code named “Baby Bear,” “Mama Bear,” and “Papa Bear” to support Air For ce aerospace programs. The largest of these was Papa Bear, which produced 30 tons/day in 1959 (Flynn, 1997). Today, total annual production of LH2 in North America is nearly 300 tons/day (Drnevich, 2003). Demand for large-scale liquid hydrogen production was initially sparked by the Apollo space program. Liquid hydrogen de mand has increased and simultaneously shifted since the 1960s from aerospace to re search and industry. Flynn reports that aerospace accounted for only 20% of total liquid hydrogen demand in 1990 (1997). This trend is expected to continue with th e onset of a hydrogen economy including the advancement of fuel cell powered vehicles and the development of improved storage systems. Hydrogen production companies are already taking advantage of the higher energy density of LH2 vs. gaseous hydrogen to effectively reduce distribution costs. Where a full tube-trailer of gaseous hydrogen cont ains approximately 300 kg of deliverable gaseous hydrogen, a comparably sized liquid hydrogen trailer carr ies 4000 kg (Drnevich, 2003). Another benefit of liquefying hydrogen is the ultra high purity that results from the majority of trace impurities condensing out. High energy density and purity make

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21 liquid hydrogen a well-suited fu el for hydrogen fuel cell powered vehicles in an emerging hydrogen economy; giving equivalent performan ce and driving range as today’s gasoline and diesel automobiles. Process Description Hydrogen, like all gases, is liquefied by cooling it to its boiling point, -423 F (-252.8 C). There are several liquefi er designs; all of which ar e derived from the simple Linde cycle shown in Figure 2.4 and follow the same general process. The incoming gas is compressed isothermally from 1 to 2 on the diagram to a relatively high pressure. Heat is rejected to a cold return stream and the cooled gas is expanded from 3 to 4 on the diagram to atmospheric pressure and cryogenic temperatures. The two-phase flow that re sults is separated in a flash tank where the liquid yield is drawn off and collected and the remaining gas absorbs heat from the warmer high-pressure stream before it’s recy cled back to the compressor. The expansion can be accomplished using either a Joule-Thom pson (expansion or throttling) valve or a work-extracting device. Isenthalpic vs. isentropic expansion Joule-Thompson expansion is modeled as isenthalpic by neglec ting potential and kinetic energy changes as well as heat transfer (insulated valve). The effect that a change in pressure has on the temperature for an is enthalpic process is described by the JouleThompson coefficient given by Equation 2.5. A negative value indicates a temperature increase with expansion; a positive value indicates a temperature decrease. T p h JTp h h T p T (2.5)

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22 By substituting the definition of specific heat at constant pressure, p pT h c p TT v T v p h and the volumetric coefficient of thermal expansion, pT v v 1, the Joule-Thompson coefficient is given in its more useful form: v T cp JT1 1 (2.6) Equation 2.6 demonstrates that the sign of the Joule-Thompson coefficient depends only on the product of T. At a given pressure, the vol umetric coefficient of thermal expansion and the specific volume are functi ons of temperature onl y. Consequently, a temperature can be identified at which 0 JT This point is known as the inversion temperature; and represents the maximum temp erature at which a gas can be cooled by isenthalpic expansion. Most practical liquefaction systems us e an expansion valve to produce low temperatures (Barron, 1985). In the case of hydrogen, however, the maximum inversion temperature at STP is well below ambient (-90.7 F (-68 C)). Additional energy is required to pre-cool the hydrogen below its inversion temperature for isenthalpic expansion to be effective. Expansion in a work-extracting or work -producing device is commonly modeled as adiabatic and reversible (i.e. isentropic). This process is represented by an isentropic expansion coefficient, Equation 2.7. s sp T (2.7)

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23 By substituting the Maxwell relation p ss v p T applying the chain rule, and using the definitions defined previ ously, the isentropic expansion coefficient is given in the same terms as the Joule-Thompson coefficient: vT cp s 1 (2.8) Equation 2.8 shows that the isentropic expansion coefficient is always positive (the temperature always decreases with pr essure) because the coefficient of thermal expansion ( ) for gases is always positive (Hands 1986). This conclusion can also be arrived at intuitively by considering conservati on of energy. If work is extracted from a fluid adiabatically, the internal energy and hence temperature must decrease. Thermodynamically, isentropic expansion is more desirable than isenthalpic expansion. The T-S diagram in Figure 2.5 sh ows that an isentropic expansion will always result in a lower final temp erature than isenthalpic expansion. s T 1 2s 2h p1 p2 T2h T2s h = const Figure 2.5. T-S diagram comparing is enthalpic and isentropic expansion Practically, however, expansion devices cannot tolerate an ap preciable amount of liquid. For this reason, expansion valves are nece ssary in all liquefaction systems (Barron, 1985). The Claude cycle discussed later seek s to combine the bene fit of isentropic

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24 expansion with the necessity of isenthalpic expansion as an efficient means of liquefying hydrogen. Ortho/para conversion Another challenge to simple liquefaction sy stems is the unique sub-atomic structure of hydrogen. Hydrogen exists in two diffe rent molecular forms: ortho-hydrogen and para-hydrogen. Each form is distinguishe d by the relative spins of its protons. The protons of ortho-hydrogen spin in the same direction whereas the proton spins of parahydrogen oppose one another. Hydrogen at ST P (i.e. normal hydrogen) is composed of 74.928% ortho and 25.072% para hydrogen. At th e normal boiling point of hydrogen (-423 F (-293.4 C) at 1 atm) the equilibrium ortho/para composition is .21%/99.79% (Flynn, 1997). Converting from ortho to para hydrogen is an exothermic process, releasing 302.4 Btu/lbm (703.3 kJ/kg) of heat at STP (B arron, 1985). The conversion process is relatively slow and the resident time of the hydr ogen within the liquefier is short, so the liquid hydrogen essentially re tains its room temperature ortho/para composition. Conversion gradually takes plac e in the storage tank resultin g in boil-off losses because the heat of conversion exceeds the latent heat of vaporization (190.5 Btu/lbm or 443 kJ/kg) (Barron, 1985). The heat liberated dur ing the conversion process is sufficient to evaporate nearly 70% of the original am ount of hydrogen liquefied (Flynn, 1997). Storage time is a major issue with regard to liquid hydrogen as a motor fuel so it is important that the boil-off losses due to ortho/para conversion are minimized. Catalysts are used to speed up the conve rsion reaction allowing the heat to be absorbed by the liquefier. This alleviates bo il-off in storage, but at a penalty to the

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25 overall efficiency of the liquefier. The most efficient method of conversion is to have the process take place simultaneously as the hydr ogen is cooled. This is not possible in practice but can be simulated by cooling the hydrogen to liquid n itrogen temperatures (-320.4 F or –195.6 C) and passing it through an adiabatic converter then repeating this procedure in a step-wise manner (Flynn, 1997). Common materials proven effective as catalysts are ferric hydroxide ge l, chromic oxide on alumina part icles, and nickel silicate; all of which provide nearly 100% conversi on to para hydrogen within a few minutes (Hands, 1986). Claude cycle The Claude cycle is the most commonl y used system for large-scale hydrogen liquefaction (Hands, 1986). The performance of the cycle is enhan ced by pre-cooling the compressed hydrogen gas to liquid nitrogen (LN2) temperatures. Adding catalysts in the LN2 and LH2 baths provides a convenient and effec tive means of absorbing the heat of conversion. Figure 2.6 shows a schematic of th is variation of the Claude cycle with labeled state points and flow paths. Hydrogen gas typically enters the cycle at 1 atm and 80.6 F (27 C) (state 1). It is compressed isothermally (isothermal compression is achieved through multistage compression with inner-cooling and after-coo ling) to state 2, typically 20 to 40 atm (Barron, 1985). The pressurized gas then ex changes heat with the return hydrogen and nitrogen streams (state 2a) before entering the LN2 bath where it is cooled to -320.4 F (-195.6 C) and where the first step of ortho/para conversion occu rs (state 2b). At this temperature, the equilibrium concen tration of para hydrogen (assuming 100% conversion) is 50%.

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26 Compressor Wc 2 1 L Liquid hydrogen bath 8 g e 7 Catalyst bed 2 5 4 2a 2b Catalyst bed 1We 10 Expander 10a 3Liquid nitrogen bath 9a 9b Figure 2.6. Claude cycle with liquid nitrogen pre-cooling and ort ho/para catalyzation It is desirable to perform the maximum amount of conversion at this stage because liquid LN2 is less expensive to produce then liquid oxygen. The hydrogen is further cooled in the first heat exchanger to state 3. At this point, a portion of the flow (typically 60 to 80%) is diverted and expande d isentropically thro ugh a work-extracting device and used to pre-cool the compressed hydrogen. The expander work is used to offset the compressor work requirement, increasing the overall cycle efficiency. The remaining flow continues through the next two heat exch angers and into the liquid receiver. Here the flow streams are halved and throttled th rough expansion valves. The liquid yield from the first stream (state 9a) is collected in the receiver and used sacrificially to absorb the heat of conversion from the second catalytic be d. The second stream (state 9b) is passed

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27 through the catalytic bed where it is ideally converted to 99.789% para hydrogen and extracted. Ammonia-Water Combined Power/Cooling Cycle The ammonia-water combined power/cool ing cycle proposed by Goswami (1995) utilizes a binary ammonia/water working fl uid to produce both power and refrigeration. The cycle is a combination of an ammoniawater refrigeration system and an ammoniabased Rankine cycle. An ammonia-water mixture is used b ecause of its desirable thermodynamic properties. Binary mixtures have varyi ng boiling points depending on the concentration of the more volatile sp ecies. This characteristic gi ves a good thermal match with a sensible heat source, thereby reducing the i rreversibility associated with heat transfer (Hasan, Goswami, 2003). Additionally, the lo w boiling point of ammonia allows the utilization of low temperature heat sources such as low-grad e waste heat from industrial processes, solar water heaters, and geotherm al sources. In a theoretical investigation performed by Tamm et al., the cycle is shown to operate with heat source temperatures as low as 116.6 F (47 C) albeit with low first law effici ency (~ 5%). When operating with a heat source temperature of 224.6 F (107 C) and idealized parameters, however, second law efficiencies greater than 65% are possible (2003). The unique ability of this cycle to produ ce both power and refrig eration gives rise to two advantages for use in a hydrogen econom y. First, the cycle can utilize low-grade renewable heat sources such as that availabl e from inexpensive flat plate solar collectors to produce the power needed to drive an elec trolyzer and liquefier Second, the cooling produced by the cycle can be used to precool hydrogen prior to liquefaction, thereby

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28 reducing the power requirement of the comp ressor. In this manner renewable energy source utilization is improved compared to technologies such as wind or P.V. electrolysis. Process Description Figure 2.7 gives a schematic of the cycle showing state points and flow paths. CHWS CHWR HHWS HHWR Cooler Vapor Generator Absorber Expander Rectifier Column Superheater Recovery Heat Exchanger HHWS HHWR Solution Pump CWS CWR CWR CWS Figure 2.7. Combined cycle flow diagram The fluid leaves the absorber at state 1 as a saturated solution at the cycle low pressure with a relatively high ammonia concentration. It is pumped to the system high pressure (state 2) before traveling th rough the recovery heat exchange r where it absorbs heat from the weak solution returning to the absorber. The solution is then partially boiled in the vapor generator by the heat source producing saturated a mmonia vapor and relatively weak concentration ammonia-water saturated liquid. The weak solution leaves the vapor generator at state 4 and rejects heat to the hi gh concentration stream before it is throttled

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29 to the system low pressure and sprayed in to the absorber. The rectifier cools the saturated ammonia vapor to condense out any remaining water. The vapor is then superheated to state 7 and expanded to pr oduce work. The sub-ambient exhaust vapor (state 8) provides refrigeration before return ing to the absorber where it is re-absorbed into the weak solution. The heat of condensat ion is rejected to the low-temperature source and the cycle repeats. The power output and cooling capacity of the cycle under given operating parameters is highly dependent on the expander e fficiency. Irreversibilties due to friction and leakage decrease the amount of work extr acted from the fluid. Because less work is extracted, the expander exhaust temperature is higher and the cooling ca pacity is reduced. Losses in the expander have the greatest impact on the overall cycle efficiency (Tamm et al., 2003), so it is important to select an optimal design. The main criteria for expander selection are operating pressures and temperatures, flow rate of ammonia vapor and material co mpatibility with ammonia. Ammonia is a corrosive substance that reacts with metals such as copper, brass, and bronze, all of which are commonly used as bearing or bushing mate rial. The expander selected for use in the combined cycle must be sized correctly for the flow rate and for the operating pressure ratio for maximum power production and refr igeration capacity. It must also be constructed out of steel, aluminum, or any other material compatible in an ammonia environment. Expander Design An expansion device extracts mechanical energy from a fluid by expanding it from a high to a low pressure and converting it in to shaft work. Various expander designs using unique expansion methods exist th roughout industry. These designs can be

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30 organized into two categories, positive-disp lacement and turbo-machinery, based on the method of fluid displacement. Positive-displacement expanders Positive-displacement machines such as reciprocating and rotary piston, rotary vane, and screw operate by expanding a fixed volume of fluid per os cillation. Torque pulsation is a common phenomenon due to the in herent discontinuity associated with the finite number of pistons or lobes and fixed displacement. Reliability is an issue with positive-displacement machines because of a greater number of moving parts (i.e. piston linkages, sliding vanes); and in the case of pi stons, a lubrication sy stem to reduce leakage encountered in the gap between the moving seals and volute. Turbo-machinery Turbo-machinery, comprised of axial and radi al flow turbines, utilizes the pressure differential across a series of ra dial blades to provide a “lif t” force to turn the rotor, thereby producing shaft work. In this manne r, a continuous power output is provided. Reliability is improved over positive-displ acement expanders because the rotor is the only moving part. Turbines are designed with a clearance between the blade tips and the volute to allow free rotation; however, leakage at the tip s (windage loss) is the primary cause of irreversibility in the expansion process. Blade tip clearances remain approximately constant for varying turbine size. As turbin e size is decreased, the loss due to windage as a percentage of the work output becomes increasingly significant. For this reason, positive-displacement expanders are more suited for small-scale operations. The amount that the blade tip clearan ces can be reduced is limited by the centrifugal force and/or thermal expansion of the blade material. Typical turbine

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31 operating speeds range from a few thousand up to tens of thousands RPM. Centrifugal force is dependent on blade tip speed, which is function of the RPM and the rotor diameter. As a result, larger turbines suffer greater radial blade de formation and are less suited for blade tip clearance reduction. Scroll compressor/expander The scroll compressor was first invent ed by Lon Creux in 1905 (Gravesen and Henriksen, 2001). Commercial interest in the technology wasn’t strong until the introduction of computer numerically cont rolled (CNC) machines in the 1970s. CNC machines provided the basis for machining the precise elements needed for a scroll compressor to operate efficiently and quietly (Copeland corp., 2001). A scroll compressor consists of two identical spiral elements assembled with a 180 phase difference. During operation, one scro ll remains stationary and the other is attached eccentrically to a motor shaft. This c onfiguration allows the sc roll to rotate in an orbiting motion within the fixed scroll. The phase difference between the two scrolls is maintained using an anti-rotation device, typically an Oldham coupling (Copeland corp., 2001). The fluid flow path within a scroll comp ressor or expander is described by Figure 2.8. As the rotating scroll (green ) orbits about the fixed scro ll (red), the outer periphery forms a line of contact with the fixed scroll capturing a crescent shaped volume of gas (step 1). The gas is forced toward the center discharge port in steps 2 thru 5 and compressed due to the decreasing volume of the crescents. This is indicated by the brilliance of the yellow color representing th e gas pocket. Because several of these gas pockets are being compressed simultaneously, as depicted in step 6, torque pulsation

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32 common with other positive-disp lacement machines is low. Scrolls compressors have been widely adopted by the HVAC industry be cause of the advantages they offer, including: simplistic design (i.e. fewer movi ng parts), low friction, low torque pulsation, and compliance. Figure 2.8. Flow path of a single fluid pocke t through a scroll compressor (Adapted from Gravesen and Henriksen, 2001) Because of their unique geometry, scrolls do not require valves or valve actuators; furthermore, there are no linkages or sliding vanes. The relative rolling motion of the contact points offers less resistance than sliding friction. Additionally, the rolling contacts provide a seal such that large volumes of oil used as a s ealant are not required and leakage is reduced (C opeland corp., 2001). Continual compression process of the scroll results in a smoother power output and consequently less noise and vibration than piston-type devices. Compliance mechanis ms balance the dynamic pressure and centrifugal forces in order to maintain prope r sealing. These loading mechanisms correct tolerances as the scroll surfaces wear and allow the scroll elements to separate slightly in the axial or radial directions in response to a sudden pressure spike (axial compliance) or 6. 3. 4. 5. 2. 1.

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33 the presence of small amounts of debris or liquid (radial complia nce). Taken together, these attributes contribute to the fact that scroll compressors typically have 10% higher mechanical efficiencies than comparably sized piston compressors (Wells, 2000) and less leakage than other compressors in its class (Schein and Radermacher, 2001). Literature suggests the potential use of a scroll compressor as a high efficiency expander (Wells, 2000). Copeland compressors have been used successfully as expanders with R-134A and R-245FA refrigerants as the working flui d. Efficiencies over 70% were demonstrated when operated with pressure ratios betw een three and five (Warner, Wayne – Copeland Corporation, Pe rsonal Conversation, 10 May 2004). Scroll expanders have also been utilized in an or ganic Rankine micro combined heat and power system patented by Yates et al. in 2002 (US Patent and Trademark Office, 2002). 5 kW Prototype The applicability of the ammonia-water combined cycle for small scale power generation utilizing low temper ature heat sources is curr ently being studied at the University of Florida’s Energy Research Par k. A prototype produci ng 5 kW of electrical power has been designed a nd is under construction. Heat source and sink. The low-temperature heat source is simulated using a liquid-propane-fired boiler to heat water to 180 F. The heat sink for the cycle is cooling water, which is continually circulat ed through a 500,000 btu/h cooling tower. Temperature control is accomplished using a combination of 3-way automatic control valves and several shell and tube heat exchangers. Absorber and solution pump. The absorber is a falling-film type. This design offers a combination of sufficiently high heat transfer rates and large surface areas for

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34 absorption. The fluid leaving the absorber is saturated, therefore no net positive suction head (NPSH) is available for the pump, leadi ng to cavitation. For this reason, a rollertype positive-displacement pump is used. Vapor generator and rectifier. The vapor generator and rec tifier are integrated as a single unit such that no separator is require d. The vapor generator is a shell and tube heat exchanger with hot water on the tube side; the rectifier is a packed column. As the ammonia bubbles out of solution, it travels through the rectifie r and the remaining effluent drips back down into the va por generator where it is re-boiled. Electricity production and cooling capacity. The maximum power output of the expander is 5.6kW. This work is used to run an electric genera tor that produces 200 Vrms single phase AC at 400 Hz. A frequency conve rter switches the frequency from 400 to 60 Hz required by the electrolyzer. The maxi mum equivalent cooling capacity of the system is 1.25 kW; this is demonstrated by cooling a fixed volume of water.

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35 CHAPTER 3 ANALYSIS METHODOLOGIES This chapter outlines the analytical procedure developed to find the expected energy requirements for electrolysis and hydr ogen liquefaction, as well as the heat and work interactions of the combined cycle at steady state. An analysis on impact of the combined cycle expander efficiency on the cooling capacity and the liquid hydrogen yield is discussed as motivat ion for an experimental study. Hydrogen Energy Requirements Electrolysis of Water The electrolyzer model used in this study is based on the Stuart Energy Vandenborre IMET Electrolyzer. The IMET is selected for two reasons: its relatively simple design due to pump-less electrolyzer circulation, and its high thermal efficiency (operating at a cell voltage of approximately 1.7V) (Stuart Energy, 2004). It utilizes an alkaline electrolyte in a filt er-press arrangement and can de liver hydrogen at pressures of up to 363 psi (25 atm), which reduces the co mpressor power required for liquefaction. The analysis determines the total electrolyzer power cons umption per unit mass hydrogen produced including the power re quired to operate the sub-sy stems of the electrolyzer, namely the cooling water system, feed water / deionization system, and AC/DC rectifier. Equation 3.1 defines the thermal efficien cy of the electrolyzer, assuming 100% current efficiency (Casper 1978). Elec H actual tn thE HHV V VElec 2 (3.1)

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36 The losses that occur in the electrolysis pr ocess are dissipated as heat. A cooling water system is employed to remove this heat and keep the electrolyte temperature relatively low. At temperatures above 302 F (150 C), the corrosiveness of the alkaline electrolyte causes significant electrode co rrosion (Wendt, 1990). The cooling load is determined using the definition of thermal e fficiency and the higher heating value (HHV) of hydrogen as shown in Equation 3.2. elec electh H coolHHV Q 12 (3.2) Using a typical COP value of three for many re frigeration systems, the work required to produce the cooling water is estimated by: COP Q Weleccool CW (3.3) The cooling water volumetric flow rate, given by Equation 3.4, is found by applying conservation of energy and specifying a 10 F (5.56 C) temperature drop across the electrolyzer. T c Qcw p cw cool cwelec (3.4) Pump work is calculated us ing Equation 3.5, assuming a pressure drop of 10ft of water and a pump efficiency of 70%. p cw cw Pp W * (3.5) The feed water required for electrolysis is obtained by assumi ng the reaction takes place in stoichiometric proportion. From the overall chemical reaction of Equation (2.1c), one mole of water is required for ev ery mole of hydrogen or 9 lbm of water for every lbm of hydrogen. On a volumetric basis, this equates to 1.0825 gal/lbm H2. The

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37 maximum energy required for deionization of water is assumed to be 10% of the energy required for electrolysis as suggest ed in the literat ure (Casper, 1978). Electh H FWHHV E21 0 (3.6) Casper reports the typical efficiency of an AC/DC rectification system to be 95% (1978). The total energy consumed per uni t mass of hydrogen by the electrolyzer and sub-systems is given by Equation 3.7. FW CW rect th H elecE W HHV Eelec 2 (3.7) Hydrogen Liquefaction The Claude cycle is analy zed to determine the total liquefaction energy per unit mass hydrogen liquefied. The inlet pressure and temperature, as well as the expander mass flow ratio are varied independently to de velop a family of performance curves used to gauge each parameter’s effect on liquid yiel d and the total specific liquefaction energy. Each configuration is then evaluated based on its figure of merit (FOM). The figure of merit (FOM) is used to measure the performance of liquefaction systems. It is defined as the ratio of the work required by an ideal liquefier to the work of an actual liquefier. W W FOMideal (3.8) Ideal liquefaction. Ideal liquefaction is described by the first two processes of a reverse Carnot cycle: isothermal comp ression followed by an isentropic expansion (Barron, 1985). Additionally, all gas that enters the cycle is liquefied. Figure 3.1 shows the T-S diagram of the process.

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38 p 1 p 2 1 2 f T s Figure 3.1. T-S diagram of ideal liquefaction process Applying the First Law to the entire cy cle (neglecting changes in potential and kinetic energy) yields: 1h h m Q Wf C net (3.9) For a reversible isothermal compression proces s, the heat rejected is given by the Second Law as: f Cs s T m s s T m Q 1 1 2 1 1 (3.10) Substituting this result into Equation 3.9 give s the ideal work requirement per unit mass gas compressed. 1 1 2s s T h h m W m Wf f f net netideal ideal (3.11) Claude cycle. The assumptions for the Claude cycle analysis are listed below: Heat transfer from the environment is negligible Heat exchangers and liquid baths are 100% effective Negligible pressure drop through pipe, fittings, and heat exchangers Negligible loss in power transmission from expander to compressor T10 = T1, T10a = T2b T7 = T8 = Te to minimize irreversibility upon mixing (Hands, 1986) T3 = -350 F Compressor efficiency, c = .75 Expander efficiency, e = .85

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39 Electrolyzer produces 100% pure normal hydrogen (74.928% ortho, 25.072% para) Ortho/para conversion proceeds to equilibrium within the liquid nitrogen (LN2) bath In this model, ortho-para conversion takes place in two isothermal stages. First, the gas is cooled to LN2 temperatures (-320.4 F, -195.6 C) and passed over a catalyst bed. Equilibrium concentration of para hydroge n at this temperature is 50.5%. This corresponds to an approximate 25.43% c onversion from normal hydrogen, releasing 75.28 btu/lbm (175.1 kJ/kg) of heat (heat of conversion at –320.4 F is 296.07 btu/lbmH2 (688.62 kJ/kg)). The second stage takes plac e in the liquid hydrogen-receiving tank at liquid hydrogen (LH2) temperatures (-423 F, -252.8 C). The heat of conversion from normal to para hydrogen at -423 F is 302.38 btu/lbm (703.3 kJ/ kg). The heat released in proceeding from 50.5% to 99.789% para hydr ogen is 134.56 btu/lbm (312.97 kJ/kg). The liquid yield of the cycle per uni t mass hydrogen compressed is found by applying the First Law to a control volume including the three heat exchangers, JouleThompson valve, and liquid hydrogen-receiving tank (subscripts refer to Figure 2.6). f f e e a f e e b C fh m h m h m m h m h m H m 10 2 2 (3.12) where 2 CH is the heat of convers ion in the second stage m is the inlet mass fl ow rate of hydrogen Dividing by m, introducing the liquid yield pe r unit mass hydrogen compressed, m m yf and the expander mass flow ratio, m m xe e and solving for y gives: 2 10 3 2 10 2 10 C f a e e e C f a b aH h h h h x H h h h h ys ad (3.13) The amount of liquid nitrogen required to pre-cool the co mpressed hydrogen and absorb

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40 a portion of the heat of conversion is determ ined by applying the First Law to a control volume encompassing the three-stream heat ex changer, liquid nitrogen receiver and the control volume from the previous analysis. f f e e C N f A N e e C f Ch m h m h m h m m h m h m h m H m H m 2 210 2 2 1 Dividing by m, defining the mass ratio of liquid ni trogen to compressed hydrogen as m m zN 2, and solving for z yields Equation 3.14 A C C f A C e e e A C A C Ch h H h h y h h h h x h h h h h h H zs ad 2 10 3 10 2 1 (3.14) where 1CH is the heat of conversion in the first stage Dividing Equation 3.14 by the liqui d yield, y, gives the hydrogen requirement in terms of unit mass hydrogen liquefied. Based on the lite rature, the specific energy required to produce liquid nitrogen is assumed 766.82 btu/lbm-N2 or 0.225 kW-h/lbm-N2 (Gross et al., 1994). An energy balance on the compressor, including work contributed from the expander, gives the specific power required per unit mass hydrogen to drive the cycle. se e e c Ch h x s s T h h m W 3 1 2 1 1 2 (3.15) Dividing this result by the liquid yield ra tio gives the compressor work per unit mass hydrogen liquefied. Total liquefaction energy is the summation of compressor work and the liquid nitrogen power requirement. The expander mass flow ratio,ex, is varied from 0 to 0.9 with four other independent parameters: expander and compress or isentropic efficiency, and compressor inlet pressure and temperature in individual cases to determine their influence on the

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41 cycle performance. In cases one and two, the expander and compressor isentropic efficiencies are decreased from 1.0 to 0.4 in 0.2 increments to gauge their effect on the cycle performance. Case three looks at a ra nge of compressor inlet pressures (1 to 25 atmospheres in increments of five) at a fixed inlet temperature of 80 F (26.7 C) to simulate the operating pressure range of th e IMET electrolyzer. In case four, the compressor inlet temperature is varied from 0 to 80 F (-17.8 to 26.7 C) in twentydegree increments; representing the pre-cooling effect of the combined cycle. Plots are created displaying the temperat ure, pressure, and component efficiency dependence of the key liquefaction parameters: total specific work, liquid yield, liquid nitrogen required, and figure of merit. The critical state points required to calc ulate the performance parameters given by Equations 3.13 thru 3.15 are defined based on th e inlet temperature and pressure (state 1) as well as the zero pressure drop assumption and the isentropic efficiencies of the compressor and expander. A computer progr am has been developed to assist in calculating the state properties and performa nce parameters for each iteration as well as for plotting the data. A detailed description of the program including a portion of the code follows in Appendix A. Ammonia-Water Combined Power/Cooling Cycle The ammonia water combined power/cooling cycle of this study is based on the experimental system under construction at th e University of Florida’s Energy Research Park. This particular system is designed to provide 5kW of electrical power from a heat source temperature of 180 F in order to simulate temperatures attainable from inexpensive flat-plate solar collectors. Additionally, the maximum pressure is

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42 constrained such that high-pressure fittings are not required, thereby reducing the capital cost. Other assumptions and/or specifica tions made in the design are listed below: Fluid exiting the absorber and vapor generator is satu rated liquid/vapor Absorber operating temperature is 100 F Vapor generator operates at 170 F Cycle high and low pressures are 110 psia (7.58 bar) and 40 psia (2.76 bar), respectively Rectification is 100% ef ficient (100% pure ammoni a vapor at state 7) Recovery heat exchanger has a 85% effectiveness, Weak and strong solution streams have equal specific heats 75% electric generator efficiency, g 5 F approach temperature in the cooler Negligible pressure drop through pipes, fittings, heat ex changers, and other components Binary mixtures differ from pure subs tances in that knowledge of three thermodynamic properties is needed to comple tely define a state (two under saturated conditions). As such, by specifying the ope rating temperature and pressure of the absorber, and assuming satura ted conditions exist at the exit, the mass fraction of ammonia in the strong solution stream is fi xed. The mass fraction of ammonia in the weak solution stream leaving the vapor genera tor at state 4 is determined in a similar matter. The next step in the analysis is to fi nd the mass flow rate of ammonia vapor through the expander. Equation 3.16 is obtaine d from an energy balance on the expander including the electric generator efficiency. 8 73h h W mg e NH (3.16) The strong and weak solution mass flow rates follow from species and mass balances on the vapor generator as described by Equations 3.16 and 3.17.

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43 weak NH strong NH strong NH NH NH weak NHX X X X m m, , ,3 3 3 3 3 3 (3.17) where X is the mass fraction of ammonia 3 3 3, NH weak NH strong NHm m m (3.18) The temperatures of the cold (state 3) a nd hot (state 5) exit stream are found from the definition of heat exchanger effectiven ess. Because the specific heats of the two streams are approximated as equal, the equati ons become a ratio of only temperatures and mass flow rates. 2 2 4 33 3T m T T m Tstrong NH weak NH (3.19) 2 4 4 5T T T T (3.21) where is the heat exchanger effectiveness Heat and work interactions of the absorb er, pump, and cooler are calculated from energy balances on all inlet a nd outlet streams. The four equations summarizing this process are given below: 1 9 6 ,3 3 3h m h m h m Qstrong NH NH weak NH ab (3.21) 1 2 ,3h h m Wstrong NH p (3.22) Cooling capacity is dependent on the temperature of the cooled fluid. It is assumed that hydrogen at 90 F is being cooled; therefore, T9 is 85 F (assuming a 10 approach temperature). 9 83h h m QNH c (3.23) The total heat input to the cycle is determ ined by “black boxing” the vapor generator, rectifier, and superheater and cons idering state points 3, 4, and 7.

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44 3 7 5 ,3 3 3h m h m h m Qstrong NH NH weak NH vg (3.24) Lastly, the cycle thermal efficiency is comput ed from the work and heat interactions as shown in Equation 3.25. The cooling affect is accounted for by scaling it with the same coefficient of performance used in the electrolyzer analysis. vg c p e cycle thQ COP Q W W (3.25) Properties at each state point are estimated using the Gibbs energy method combined with pure fluid correlations as described by Tamm (2003). This procedure is repeated for a fi xed power output and varied expander efficiencies. These data are pl otted to study the eff ect on the cycle coo ling capacity, heat input, and pump work and to relate thes e quantities to the liquid hydrogen yield. Additionally, the effect of trace quantities of water in the expander inlet stream on cycle efficiency and cooling capacity is analyzed. A MatLAB program is developed to calcu late all state point s of the combined cycle, equations 3.16 thru 3.24, and the optimum liquid hydrogen yield for each value of expander efficiency. A detailed description of the program and portions of its code are presented in Appendix A.

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45 CHAPTER 4 EXPERIMENTAL SETUP AND DESIGN The potential application of a scroll comp ressor as a high-efficiency expander for small-scale power generation (i.e the 5kW combined cycle) is discussed in this chapter as background for the experimental study. A detailed description of the compressor and testing apparatus is given followed by an outline of the experimental methods. Scroll Machines as Expanders Scroll compressors have been proven as viable expansion devices. Copeland has performed limited research on scroll expanders using their refrigeration scroll compressor with R-134A and R-245FA as the working fluid. Results show that efficiencies of greater than 70% are attainable (Warner, Wayne – Copeland Corporation, Personal Conversation, 10 May 2004). Other publications have investigated the use of scroll expanders in small-scale solar driven Rankine cycles (Wells, 2000). To date, however, no known research has been conducted w ith an ammonia working fluid. Ammonia offers particular challenges to the design or selection of any expander. One of which is corrosiveness. Ammonia is corrosive to coppe r and copper-containing alloys present in the bearings and motor stators of hermetically sealed compressors like those manufactured by Copeland. Additionally ammonia is a small molecule and thus has relatively low density compared to R134-A (0.0433 lbm/ft3 vs. 0.2622 lbm/ft3), so leakage losses become more prevalent. Small-scale, high-efficiency expanders are desired for the 5kW ammonia-water combined power/cooling cycle because its ov erall performance and cooling capacity is

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46 highly dependent on the expander efficiency as discussed in later sections. For a designed power output, increasing the expande r efficiency reduces the required mass flow through the system and hence reduces th e total energy consumption. Individual component and pipe size is reduced as well. At the 5kW size, the scroll design offers several advantages over turbines as explained in the background and theory. Ammonia turbines in the 5kW range are inherently inefficient due primarily to leak age loss at the tips. Tom Revak of Revak Industries reports that the efficiency of a 5kW is likely to be approximately 40% whereas Sam Ni of Scroll Labs predicts an isentropi c efficiency of 67% for a comparably sized scroll expander. Custom-design is cost prohibitive however; with the design and fabrication cost of the aforementione d scroll expander being $280,000. The objective of the experiment is to test an “off-the-shelf” unit with air and predict its performance with ammoni a from the data obtained. From these observations, an indication of whether the scroll expander is f easible in the combined cycle is determined and recommendations for design improvements are made. This experime nt also lays the foundation for further research of scroll expanders for use in the ammonia-water combined cycle and other small-scale power generation systems. Testing Apparatus and Instrumentation The Sanden TRS-90 automotive scroll comp ressor (shown in Figure 4.1) was selected as the test compressor for three reasons: it operates in the 5kW range, the scroll elements and the housing is constructed of alum inum and the bearings and clutch of steel (ammonia compatible), and it has a pulley and clutch assembly convenient for testing. The only modification necessary to run the co mpressor in reverse is the removal of a reed-type check valve located beneath the stationary scroll element within the housing.

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47 The compressor is designed to operate at a pressure ratio of appr oximately six with R134A refrigerant. Displacement of the compressor is 85.7 cc/rev. Figure 4.1. Sanden TRS-90 automotive scroll compressor and test stand The expander is connected to compressed air source at the suction port (1) using ” I.D. plastic tubing. The discha rge port (2) is ” I.D. and is vented to the atmosphere. Also shown in Figure 4.1 is the pulley and clutch assembly (4). The clutch is on/off modulated by applying 12 volts DC at point 3. Figure 4.2 shows the 5-Hp compressor and tank used as the compressed air source. The compressor has a maximum pressure of 125 psig and a pumping capacity of 15.7 scfm at 90 psig. A 110-psig regulator is used to adjust the expander inlet pressure. Temperatures measurements are taken from thermocouples inserted into the inlet and exit flows at points 1 and 2 as shown in Figure 4.3. The signal from each thermocouple is calibrated and conditioned to 1mV/F using two thermocouple-to-analog converters (3) and recorded from a pair of multimeters. 1 3 4 2

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48 Figure 4.2. Piston compressor with integrated tank and regulator Figure 4.3. Thermocouple locations and flow meter The volumetric flow rate of compressed air is measured in standard cfm (standard conditions are 1 atm and 70 F) using an in-line acrylic gas rotameter (number 4). The 1 3 4 2

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49 reading is adjusted to actual cfm using th e ideal gas relation with the observed inlet temperature and pressure as described by Equation 4.1. Figure 4.4 shows the pony brake used to meas ure the torque output of the expander and the back pressure gauge (1). The pony brake frame is constructed of wood with ordinary go-cart brake pad material employed as the friction material. An enlarged view of the pulley showing the brake material is s een in Figure 4.5. This material has the added advantage in that it act s as an insulator, protecti ng the wood from the excessive heat. The frictional force applied to the pulle y is varied by adjusti ng a pair of wing nuts (2). The force exerted by the expander torq ue is measured 14.125” from the centerline of the expander shaft (3) using a Pelouze 5-pound sc ale. Rotational speed is measured in RPM from the center of the pulley with a handheld tachometer (not shown). Figure 4.4. Pony brake and back pressure gauge and valve 1 3 2

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50 Figure 4.5. View of expander pulley showing the brake pads used as frictional surfaces A detailed component list of the experi mental apparatus including the range and resolution of each instrument (if applicable) follows in Appendix C. Experimental Methodology Procedure Startup: 1. Activate the voltage supply, multimeters, and thermocouple-to-analog converters. 2. Close the compressor valve. 3. Start the compressor and allow it to charge to 125 psig. Test: 1. Cap the expander exit port. 2. Crack the compressor valve and allow system to charge. 3. Select the desired source pressure by adjusting the tank regulator. 4. Once pressure is selected, close the compressor valve and open the backpressure valve to discharge the system. 5. Close the backpressure valve and remove the expa nder exit port cap. Brake pads Tachometer p lacement

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51 6. Loosen the wing nuts on the pony brake to ensure that testing begins with minimum brake force. 7. Initiate the test by fully opening the compressor valve. 8. Record rotational speed (RPM), inle t and exit temperature, flow rate, backpressure and arm force. 9. Tighten the pony brake wing nut s about 1/8 of a turn and repeat step 8 for each trial. 10. Continue until the expander is stalled. 11. Terminate the test by closing the compressor valve. 12. Allow 15-20 minutes between each te st for the compressor motor and expander clutch assembly to cool. Data Analysis Experimental data is collected in an Ex cel spreadsheet programmed to perform the conversions and calculations necessary to complete the analysis. Each calculation performed in the spreadsheet and the formulas used for them are explained below. The corrected volumetric flow rate for the given inlet pressure and temperature is related to the indicated value by treating th e air as an ideal gas (Equation 4.1). T p pT V Vo o indicated corrected (4.1) where o oT p,are at standard co nditions (1 atm and 70 F) The mass flow at standard conditions is found by multiplying the fluid density by the corrected volumetric flow rate as described by Equation 4.2. corrected o oV RT p m (4.2) where RT p is substituted for the density Mass flow is corrected to th e actual inlet cond itions using Equation 4.3 (Holman, 2001). 2 1 o o correctedpT T p m m (4.3)

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52 Shaft power output is defined by Equation 4.4, the product of the force measurement and the expander rotational speed. Force Wshaft (4.4) The volumetric efficiency quantifies the amount of tip leakage encountered during operation. It is defined as the ratio of flow usefully expanded to the total flow through the expander (Equation 4.5). V dv (4.5) where is the rotational speed (RPM) d is the expander displacement per revolution Inlet and exit enthalpies are computed from the measured temperatures and pressures and are used in Equation 4.6 to calcu late the isentropic efficiency. sout in out in eh h h h (4.6)

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53 CHAPTER 5 RESULTS AND DISCUSSION The electrolyzer and its sub-systems are analyzed to find the specific energy consumption, thermal efficiency, and cell voltage. Following the electrolyzer investigation, simulations of the Claude cy cle are made to determine the effects of component efficiencies and compressor inlet co nditions on specific energy consumption. Results of each test are presented in tabul ar form with several graphs displaying the important trends. The analysis concludes w ith the selection of the optimum operating parameters. The ammonia-water combined cycle simulation examines the dependency of the boiler heat input, pump work, and cooling capac ity on the expander efficiency for a fixed output and establishes the motivation for th e scroll expander performance study. The influence of trace amounts of water in the vapor stream on cycle performance is also investigated. The analytical portion of the re sults concludes with the calculation of the maximum rate of hydrogen production. Results of the scroll expander performa nce study are examined to predict the expander’s behavior with ammonia and to determine its feasibility for use in the combined cycle. Several trends are develope d to describe the performance of the scroll expander. The data is compared to a perf ormance chart of the same unit operated as a compressor in order to determine if such information can reliably predict expander performance.

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54 Hydrogen Production and Liquefaction Electrolysis of Water Specific energy requirements for the electrol ysis of water are displayed in Table 5.1. The majority of the el ectrical energy is re quired by the electrolyze r itself with the subsystems representing only 16.2% of the to tal. Cooling water pump work is found to be negligible compared to the energy consumed by the cooler (0.005 kW-h/lbm-H2 compared to 0.884 kW-h/lbm-H2). Including all subsystems, the total specific energy required to electrolyze water is 24.839 kW/lbm-H2 (54.76 kW-h/kg-H2). Contrasting with the energy requirement of therm oneutral electrolysi s (17.865 kW-h/lbm-H2 (39.385 kW-h/kg-H2)), the electrolyzer has a thermal effici ency of 85.8%; however, the efficiency drops to 71.9% when all subsystems are cons idered. At 85.8% electrolyzer efficiency, the cell voltage required to drive the process is 1.713 V. Table 5.1. Specific energy requirements of the IMET electrolyzer kW-h/lbm-H2kW-h/kg-H2Electrolyzer20.81445.886 AC/DC Rectifier1.0952.415 Cooling Water0.8441.860 Feed Water2.0814.589 Pump0.0050.010 Total 24.83954.760 Energy Requirements The amount of cooling water and feed water corresponding to their energy consumption are 1.726 gpm/lbm-H2 (6.534 Lpm/lbm-H2) and 1.085 gal/lbm-H2 (4.107 Lpm/lbm-H2), respectively. Hydrogen Liquefaction Initial inspection of equations 3.13 and 3. 15 indicate that the li quid yield and work per unit mass hydrogen compressed are proportio nal to the expander mass flow ratio. This is evidenced more clearly by defining the work per unit mass LH2 (Equation 5.1).

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55 2, 2 10 3 2 10 3 1 2 1 1 2N f C f a e e e b a e e e c f C fw y z H h h h h x h h h h x s s T h h y m W ws ad s (5.1) (State points referenced from Figure 2.6). Equation 5.1 shows that increasi ng the expander mass flow ratio, ex, always reduces the specific work for a given set of operating conditions; however, the amount of liquid yield is physically constrai ned as described by Equation 5.2. 1 y xe (5.2) The liquid yield continues to increase as de fined by Equation 3.13 until the constraint is met at which time it becomes a mono tonically decreasing function of ex and T5. This implies that an optimum value of the expa nder mass flow ratio exists at which the liquefaction energy is minimized. The exact form of the constraint is f ound from an analysis of the third heat exchanger and the expansion valve. Heat exchanger cold side inlet and outlet temperatures Tg = -423 F (-252.8 C) and T7 = -402.32 F (-241.29 C) are known from the saturation temperature of hydrogen at atmospheric pressure and by assuming T7 = Te, respectively. The “hot” side inlet temperature T4 = -402.32 F (-241.29 C) is equal to T7 because the flow passes through the 100% effective second heat exchanger as the minimum capacity stream. The percent of th e mass flow through the J-T valve that is liquefied, k, is initially guessed as 80. T5 is then calculated from Equation 5.3 and used to find the quality of the expanded stream. Th e value of k is iterated until convergence is achieved. gT T k T T 7 4 51 (5.3)

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56 Convergence is achieved in only thr ee iterations with k = .725 and T5 = -408.05 F (-244.47 C) because the temperature change of the supply stream is restricted by the lower volume of the return stream. This exactly defines the constraint as: ex y 1 725 (5.4) The optimum value of xe occurs when y exactly equals the constraint; an example of which is seen in Figure 5.1. Figure 5.1. Sample output showing the optimum expander mass flow ratio, xe Prior to analyzing the effect of co mpressor inlet temperature and pressure variations on the performance parameters of the Claude cycle, the expander and compressor isentropic efficiencies are studied independently with regard to motivation for further research and deve lopment of these components. Expander efficiency. The effect of the expander isentropic efficiency on the Claude cycle performance is summarized in Table 5.2. As an approximation, the liquid yield constraint is held c onstant. In reality, however the liquid yield is further optimum

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57 constrained with decreasing expander efficiency. At e = 0.4, the percent of the source stream liquefied is approximately 48% compared to 72.5% for e = 0.85. The simulation was run with the compressor efficiency fi xed as 100% and an inlet temperature and pressure of 80 F and 25 atm, respectively. Table 5.2. Claude cycle simulation results for expander isentropic efficiency variation wf,minwidealkW-h/lbm-LH2kW-h/lbm-LH210.58900.297515.3863.6841.2680.3444 0.80.63500.264615.4853.7431.2680.3389 0.60.68880.225615.6383.8351.2680.3308 0.40.75220.179615.9043.9951.2680.3175 Ymax(z/y)optFOMmaxeXe,opt Table 5.2 shows that an incr ease in the expander efficiency from 40% to 100% reduces the optimum expander mass flow ratio by 21.7% from 0.7522 to 0.589. This shift in xe increases the maximum liquid yield by 65.4% since it is re lated by the constraint of Equation 5.4. The trend between xe,opt and ymax for different expande r efficiencies is observed in Figure 5.2. Table 5.2 also shows a 3.26% reduction of the cycle liquid nitrogen requirement. The relationship betw een these two parameters is depicted in Figure 5.3. 0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.40.50.60.70.80.91 Expander Isentropic Efficiency Specific Liquid Yield Expander Mass Flow Ratio Figure 5.2. Specific liquid yield and expander mass flow ratio as functions of the expander efficiency

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58 15.3 15.4 15.5 15.6 15.7 15.8 15.9 16 0.40.50.60.70.80.91 Expander Isentropic EfficiencyLiquid Nitrogen Requirement (lbm/lbm-LH2) Figure 5.3. Required liquid ni trogen vs. expander efficiency The combination of these effects results in a 7.78% reduction in specific work as described by equation 5.1. This is shown in Figure 5.4, as well as the shift in xe that accompanies the decrease in specific work. The ideal work requirement depends only on the inlet and liquid conditions and hence is unc hanged; therefore, the FOM scales directly with specific work. Figure 5.4. Specific work vs. expa nder mass flow ratio for varied e

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59 Compressor efficiency. By inspection of Equations 3.13, 3.14, and 5.1, it is clear that the expander mass flow ratio, liquid yield, and liqui d nitrogen requirement are independent of compressor isentropic effi ciency for fixed inlet and outlet conditions; provided that the compressor can supply the necessary pressure. The simulation was run with an expander efficiency of 100% and an inlet temperature and pressure of 80 F and 25 atm, respectively. Results are displayed in Table 5.3. Table 5.3. Claude cycle simulation results fo r compressor isentropic efficiency variation wf,minwidealkW-h/lbm-LH2kW-h/lbm-LH210.58900.297515.3863.6841.2680.3444 0.80.58900.297515.3863.7461.2680.3386 0.60.58900.297515.3863.8511.2680.3294 0.40.58900.297515.3864.0591.2680.3125 cXe,optYmax(z/y)optFOMmax The only effect that increasing the compressor e fficiency has is to lower the specific work from 4.059 to 3.684 or 9.2%, as illustrated by Figure 5.5. The ideal work is again independent of component effici ency and thus scales directly with the specific work. Figure 5.5. Specific work vs. expa nder mass flow ratio for varied c

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60 Comparing the figures of merit for the tw o cases in Figure 5.6 indicates a more profound impact of the compressor efficiency on the cycle performance. A 60% decrease in efficiency from ideal results in a 9.2% reduction in the FOM for the compressor case compared to 7.8% for the expander. A great er emphasis should therefore be placed on the development of high efficiency hydr ogen compressors to minimize liquefaction energy. 0.31 0.315 0.32 0.325 0.33 0.335 0.34 0.345 0.35 0.40.50.60.70.80.91Isentropic EfficiencyFigure of Merit (FOM) Expander Efficiency Effect Compressor Efficiency Effect Figure 5.6. Impact of compressor and expa nder efficiencies on Claude cycle FOM Compressor inlet pressure. Compressor inlet pressure is varied to determine the advantage of using pressurized electrolysis. A simulation was run fo r an outlet pressure of 40 atm and compressor and expander effi ciencies of 75 and 85%, respectively. Table 5.4. Claude cycle simulation results for compressor inlet pressure variation P1wf,minwidealatm kW-h/lbm-LH2kW-h/lbm-LH210.62300.273315.4576.2751.7720.2824 50.62300.273315.4575.0501.5210.3012 100.62300.273315.4574.5211.4130.3125 150.62300.273315.4574.2101.3490.3204 200.62300.273315.4573.9891.3040.3268 250.62300.273315.4573.8171.2680.3357 FOMmaxXe,optYmax(z/y)opt

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61 Table 5.4 shows that the op timum expander mass flow ratio, maximum liquid yield, and minimum liquid nitrogen requireme nt are independent of the compressor inlet pressure. This is true because the inlet temperature and exit pressure are held constant and the compression process is modeled as isothe rmal. Furthermore, the liquid yield and expander mass flow ratio are decoupled from the compressor inlet and exit temperatures by the 100% effective cooling ba th assumption. The functi onal relationship between the specific work and inlet pressure is gr aphically described in Figure 5.7. 3 3.5 4 4.5 5 5.5 6 6.5 051015202530 Compressor Inlet Pressure (atm)Work per Unit Mass LH2 Figure 5.7. Effect of compressor in let pressure on the specific work Increasing the pressure from 1 to 25 atmos pheres reduces the specific work requirement by 39.2% while also reducing the theoretical work requirement by 28.4%; the net result is an 18.9% increase in the figure of merit. Compressor inlet temperature. Compressor inlet temperatures ranging from 0 to 80 F are analyzed to gauge the merits fo r hydrogen pre-cooling using the combined cycle. A simulation was run for an inlet pressu re of 25 atm, exit pressure of 40 atm, and compressor and expander efficiencies of 75 and 85%, respectively.

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62 Results of the simulation are summarized in Table 5.5. An interesting result of this simulation is the reduction in the figur e of merit with lower inlet temperatures. Table 5.5. Claude cycle simulation results for compressor inlet temperature variation T1wf,minwideal(F) kW-h/lbm-LH2kW-h/lbm-LH200.62300.273315.6213.8021.0040.2639 200.62300.273315.5773.8061.0690.2808 400.62300.273315.5353.8091.1340.2978 600.62300.273315.4953.8131.2010.3150 800.62300.273315.4573.8171.2680.3323 Ymax(z/y)optFOMmaxXe,opt This phenomenon is explained by the increase in the liquid nitrogen requirement due to the assumption that Tc = T2. Reducing the inlet temperature of the compressor effectively decreases Tc and lowers the enthalpy at state C. From Equation 3.14, the denominator hc – ha is reduced from 187.2 Btu/lbm (435.5 kJ/kg) to 167.3 (Btu/lbm) 389.2 kJ/kg, thereby increasing the liquid nitrogen requirement from 15.457 to 15.621 lbm-LN2/lbm-LH2. This effect partially offsets the benefit of pre-cooling the hydrogen gas resulting in a 0.393% decrease in specific work. The ideal specific work simultaneously decreases by 20.8%. As a result, the FOM actually decreases by 20.6% from 80 to 0 F. The liquid nitrogen requirement and specific work in relation to the compressor inlet temperature is presented in Figures 5.8 and 5.9. 15.44 15.46 15.48 15.5 15.52 15.54 15.56 15.58 15.6 15.62 15.64 0 10 20 30 40 50 60 70 80 Compressor Inlet Temperature (F)Liquid Nitrogen Requirement (lbm/lbm-LH2) Figure 5.8. Liquid nitrogen requireme nt vs. compressor inlet temperature

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63 Figure 5.9 concludes that the reduction in specific liquefaction energy due to decreased inlet temperature can be neglecte d within the specified range. A qualitative comparison of the figures of merit made between cases three and four show that increasing inlet pressure is significantly more effective in reducing the specific liquefaction energy (Figure 5.10). 3.8 3.802 3.804 3.806 3.808 3.81 3.812 3.814 3.816 3.818 3.82 0 10 20 30 40 50 60 70 80 Compressor Inlet Temperature (F)Work per Unit Mass LH2 (Btu/lbm) Figure 5.9. Specific work requireme nt vs. compressor inlet temperature The analysis concludes that the optimum operating point for the Claude cycle under the assumed component efficiencies (e = 0.85, c = 0.75) is P1 = 25 atm. The cycle parameter values are displayed in Table 5. 6 at the optimum design point and at normal operating conditions. The specific work of liquefaction is reduced by 46.7% and the figure of merit is increased 18.9%. The total energy required to produce and liquefy hydrogen is 28.656 kW-h/lbm-H2 (63.175 kW-h/kg-H2). Over 86% of this energy is c onsumed in electrolysis. Properties at each state point are given for the optimum design condition in Appendix B. Given the conclusion that compressor inlet temperature has little effect on the specific work, the

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64 maximum hydrogen production rate of the combined cycle is 0.1775 lbm/hr or 7.21 gallons/day. 0.25 0.26 0.27 0.28 0.29 0.3 0.31 0.32 0.33 0.34Figure of Merit (FOM) Increasing Inlet Pressure Decreasing Inlet Temperature Figure 5.10. Comparison of inlet pressure and temperature affect on the cycle FOM Table 5.6. Claude cycle performance para meters for normal and optimum configuration NormalOptimum T1= 80F P1= 1 atmT1= 80F P1= 25 atm Xe,opt0.62300.6230 Ymax0.27330.2733 (z/y)opt15.45715.457 wf,minBtu/lbm-LH221411.7013025.04 kW-h/lbm-LH26.2753.817 widealBtu/lbm-LH26046.884327.92 kW-h/lbm-LH21.7721.268 FOMmax0.28240.3357 Ammonia-water Combined Cycle The ammonia-water combined cycle was analyzed for a fixed power output to observe the impact of expander efficiency on the heat and work requirements as well as the cooling capacity. The simulation was run under the assumptions listed in chapter 3 while varying the expander isentropic effici ency from 10 to 100% in 10% increments.

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65 Results of the simulation are given in Figures 5.11 thru 5.15 and a sample output of the simulation is given in Appendix B. The ammonia vapor mass flow rate requi red to drive the e xpander and produce 5kW of electricity is a functi on of only the exhaust enthalpy at state 8; since the power output and specified temperatures and pressures are held constant. From the definition of isentropic efficiency, the vapor mass flow scales with 1/e. The weak and strong solution flow rates follow a similar trend as shown in Figure 5.11 as they are related to the vapor flow by a constant ratio of the ammonia mass fractions as described by Equations 3.17 and 3.18. The maximum mass flows for the vapor, weak, and strong solutions occur at the lowest efficiency (10%) and are are: 3586.34 lbm/hr (0.4519 kg/s), 72037.7 lbm/hr (9.0767 kg/s), and 75624.1 lbm/hr (9.5286 kg/s), respectively. Figure 5.11. Mass flow rate de pendence on expander efficiency

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66 Likewise the minimum mass flows are (i n the same order): 358.63 lbm/hr (0.0452 kg/s), 7203.77 (0.9077 kg/s), and 7562.41 (0.9529 kg/ s). The mass fraction of the strong, weak, and vapor streams are xS = 0.3988, xW = 0.3689, and xV = 1.0 (assumed). A decrease in mass flow through the system manifests itself in the reduction of work and heat interactions of the cycle for a given output as seen in Figures 5.12 and 13. Figure 5.12. Pump work varia tion with expander efficiency Minimum pump work, boiler heat input and abso rber heat rejection are: 0.724 Hp (0.540 kW), 331,566 Btu/hr (97.18 kW), and 321,117 Btu/hr (94.11 kW). The ideal cooling capacity under these conditions is 9130.26 Btu/hr (2.68 kW). Figure 5.14 concludes that at least 60% efficient expansi on is required to obtain any cooling capacity. Below this point, the exhaust temperat ure of the expander exceeds the assumed temperature of the substance to be cooled (85 F). This effect is also evidenced in the plot of the expander isentropi c efficiency versus the cycle thermal efficiency (Figure 5.15).

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67 Figure 5.13. Boiler heat input and absorber heat rejection vs. expander efficiency Figure 5.14. Cycle cooling capacity as a function of expander efficiency Thermal efficiency increases linearly w ith expander efficiency; however, a sudden increase in slope occurs at approximately e = 0.6 at which point the cooling effect

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68 begins to enhance the thermal efficiency of the cycle. The highest achievable thermal efficiency for the given operating conditions is 7.22%. Figure 5.15. Cycle thermal effici ency vs expander efficiency The mass fraction of ammonia entering the e xpander was analyzed more closely to judge the assumption of pure vapor leaving the rectif ier and to determine the impact that trace quantities of water have on the cycle performance. The analysis was carried out for an ideal expander. Figure 5.16 shows the profound negative effect on cooling capacity. The cooling capacity diminishes to zero rapi dly as trace amounts of water are introduced into the expander stream up to only 2.5% by mass. Boiler heat input is reduced from 331,566 Btu/hr (97.18 kW) to 320,934 Btu/hr (94.06); however, thermal efficiency is reduced 9.26% from 7.22% to 6.55% because of the lost cooling benefit. Another concern is that the expander exhaust temper ature drops below the mixture dew point as shown in Figure 5.17. At a 2.5% water vapor concentration by mass, the mixture quality is 0.967. This most likely is not an issue fo r compliant devices such as scrolls in which a

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69 small quantity of liquid can be tolerated, or with high-speed devices such as turbines, in which the residence time of the fluid is shorte r than the time require d for condensation to occur (metastable condition). 0 1000 2000 3000 4000 5000 6000 7000 8000 9000 10000 0.975 0.98 0.985 0.99 0.995 1 Expander Ammonia Mass Fraction, x7Cooling Capacity (btu/hr) Figure 5.16. Effect of trace amounts of wate r within in the expander inlet stream on cycle cooling capacity 250 260 270 280 290 300 310 320 330 0.975 0.98 0.985 0.99 0.995 1 Expander Ammonia Mass Fraction, x7Temperature (K) Isentropic Exhaust Dew Point at 40 psia Figure 5.17. Expander exhaust and dew point temperature at several water concentrations These results show that rectifier design is a crucial element for the success of a smallscale combined cycle in the hydrogen production field in which high efficiency translates into greater liquid yield per unit energy input.

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70 Scroll Expander Performance Study Scroll expander performance was measured for inlet pressures of 60, 70, and 80 psig; a range suitable for the 5kW combined cycle. Two tests were performed at each pressure to verify repeatability of the result s. Tests at pressures over 80 psig were not feasible due to the relativel y small tank and the inability of the compressor to supply compressed air at high flow rates (> 60 scfm). Furthermore, the compressor motor is equipped with a high-temperature cut-o ff switch that disconnects power after approximately 10 minutes of continuous operati on. A fan was used to aid in cooling the motor; however periods of lockout continued to occur, limiting the maximum duration of each test. Figure 5.18 shows the results of the repeat ability analysis applied to shaft power measurements at 65 psig. The second set of data indicated by the square points agrees well with the trend line of the initial data. 0 0.05 0.1 0.15 0.2 0.25 0.3 0.35 050010001500200025003000 Rotational Speed (RPM)Shaft Power (Hp) Figure 5.18. Repeatability analysis a pplied to shaft power output at 65 psig Results of the study are summarized in Fi gures 5.19 thru 5.22. Shaft power output is plotted with respected to expander rota tional speed in Figure 5.19. Rather than

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71 beginning at a maximum value and decreasing monotonically with RPM as expected, the power output reaches a maximum at approxi mately 1500 RPM before decreasing toward zero in all three cases. This is thought to occur due to c hoked conditions at the expander exit. Flow becomes choked when the port to fit ting area ratio is smaller than the critical area ratio given by the temperature and pressu re of the exiting air. The area of the expander exit port and fitting is 0.375” and 0.25”, respectively. Further evidence of choked flow is given by the fact that the ma ximum attainable rota tional speed is only 3000 RPM at source pressures up to 110 psig, whereas the TRS-90 scroll compressor can normally achieve speeds of up to 9000 RPM (S anden engineer, personal conversation). 0 0.05 0.1 0.15 0.2 0.25 0.3 0.35 0.4 050010001500200025003000 Rotational Speed (RPM)Shaft Power (Hp) 60 psi 70 psi 80 psi Figure 5.19. Shaft power vs. rotational speed at 60, 70, and 80 psig inlet pressure A similar trend is witnessed with isentropic efficiency, e (Figure 5.20). Low values of e are attributed to the poor volumetric efficiency, v, of the expander at low RPM and relatively high torsional load. Increa sed torsional resistance raises the pressure within each pocket of the scroll, enha ncing tip leakage and reducing volumetric efficiency. Figure 5.21 illustrates the rela tionship between volumetric efficiency and

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72 rotational speed. At each pressure, v increases asymptotically toward a final value between 0.8 and 0.9. 0 0.02 0.04 0.06 0.08 0.1 0.12 0.14 0.16 0.18 0.2 050010001500200025003000 Rotation Speed (RPM)Isentropic Efficiency 60 psi 70 psi 80 psi Figure 5.20. Scroll expande r isentropic efficiency The volumetric efficiency indicates the percentage of air that passes through without doing any useful work. This proce ss can be modeled as isenthalpic, with the approximation of constant temperature (ideal gas). The warmer air mixes with the cold air, from which work was extracted, with in the scroll housing e ffectively raising its temperature prior to the measurement location. Furthermore, heat is exchanged from the surroundings to the fluid through the exit port fi ttings. This temper ature rise causes an erroneous calculation of the exit enthalpy and thus the isentropic efficiency. However, trends may still be observed to determine wher e the point of maximum efficiency occurs. The exit temperature variation with rotati onal speed is shown in Figure 5.22. The points of minimum exit temperature coincide with those of maximum power output as expected from the First Law of Thermodynamics.

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73 0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 0500100015002000250030003500 Rotational Speed (RPM)Volumetric Efficiency 60 psi 70 psi 80 psi Figure 5.21. Volumetric efficiency va riation with expande r rotational speed 30 32 34 36 38 40 42 44 46 48 50 050010001500200025003000 Rotational Speed (RPM)Exit temperature (F) 60 psi 70 psi 80psi Figure 5.22. Expander exit temperatur e and rotational speed relationship The maximum power output of 0.368 Hp (0.274 kW) occurred at 1460 RPM for the 80-psig inlet pressure case. The most e fficient operating point is 18.2%. Rotational speed, inlet pressure, and power output at this point is 2000, 80 psig, and 0.282 Hp. The temperature of the working fluid (excluding leakage) is found at any point using the volumetric efficiency and flow rate in Equation 5.5.

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74 v inlet v exit wf inlet leakage exit wfT T cfm T cfm T T 1 (5.5) Therefore, with a volumetric efficiency of 0.6092 and temperatures of 71.6 F and 31.5 F at the inlet and exit at this point, th e temperature of the working fluid is 4.77 F. The low value of isentropic efficiency is due primarily to leakage caused by the density mismatch. The TRS-90 is designed for R-134A with a density of 0.262 lbm/ft3 at STP whereas the density of air at STP is .07298 lbm/ft3; nearly 3.6 times lower than R134A, and the density of ammonia is 0.04333 lbm/ft3; 1.6 times lower than air. The performance of the expander with ammonia is expected to be worse than with air because higher pressures are required for a unit volum e of ammonia to store an equal amount of energy as a unit volume of air at a given temper ature. This relations hip is arrived at by considering the ideal gas law as a fi rst approximation (Equation 5.6). 8 6 1 air ammonia ammonia air air ammoniam m p p mT p mT p (5.6) Higher pressures lead to increased leakage with in the scroll and a loss of performance. Additionally, ammonia is a smaller molecule than air and much smaller than R-134A, further facilitating tip leakage and reducing efficiency. Fundamental design changes are required fo r the scroll concept to be utilized as an expander. The geometry of each scroll el ement should be altered such that the total number of chambers is increased as shown in Figure 5.23. This design reduces pressure differences between chambers and hence leakage (Hans-Joachim and Radermacher, 2003).

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75 Figure 5.23. Comparison of optimum geomet ries of a scroll compressor (left) and expander (right) (Adapted from Hans-Joachim and Radermacher, 2003)

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76 CHAPTER 6 RECOMMENDATIONS Analytical Study The analytical study of the electroly zer, Claude cycle, and ammonia-based combined power/cooling cycle examined a limited range of operating parameters. By modeling the overall process w ith a program such as ASPEN, a greater number of operating configurations could be analyzed. ASPEN is a chemical processing software package that allows the user design a cycle and specify a set or range of operat ing and boundary conditions. Using algorithms included in the code for most devices, ASPEN performs a complete thermodynamic analysis and outputs user specified da ta in an interactive manner. Additionally, an optimization of the combined cycle for maximum hydrogen production would indicate the operating cond itions, power output, and overall system size required to minimize energy cost. The economic viability of a large-scale implementation of this system should be ex amined through a life-cycle cost analysis. Scroll Expander Performance Test The scroll expander used in the perf ormance test was an automotive airconditioning compressor modified to run in re verse. Recommendations for future scroll expander experimentation are: 13. Test the expander in a closed loop system with ammonia vapor 14. Pre-heat the inlet vapor to simulated the combined cycle operating conditions 15. Re-design the compressor housing to allow higher flow rates and eliminate choking 16. Design an oil injection and separatio n system to reduce leakage losses 17. Use a dynometer or motor to improve control on the applied torque

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77 Future work should also include improveme nts to the scroll design. Manufacturing the scroll involute using the optimum expa nder geometry shown in Figure 5.23 would improve its performance as an expander. Fu rthermore, the use of low-friction materials such as those under development at the Univer sity of Florida would eliminate the need for an oiling system, making the scroll an attractive design for the ammonia based combined power/cooling cycle.

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78 CHAPTER 7 CONCLUSIONS Global energy consumption is projected to increase 54% over the next 25 years. With proven oil reserves being called in to question beyond 2030 it is important to develop renewable technologies to sustain the future global energy demand. By introducing an alternative fuel for transporta tion only, oil consumption can be reduced by as much as 20%. Hydrogen has many characteristics that make it a desirable fuel. It has the highest energy content per unit mass of any known fuel – nearly 3 times higher than gasoline, it burns cleanly and efficiently, and it can be produced from water via electrolysis powered by renewable energy. Two major obstacles to the emergence of a hydrogen economy are the limited means available to efficien cy produce mass quantities of hydrogen from renewable energy sources and the storage issu es related to the low energy density of hydrogen. Liquefying hydrogen provides a so lution to its low density; however, the process requires additional energy. This thesis explored the possibility of using a 5-kW ammonia-based combined power/cooling cycle to produce hydrogen from re newable resources and pre-cool it prior to liquefaction in an effort to reduce the overall energy consumption. The advantage of this cycle is its ability to utilize low temperature heat s ources available from solar and geothermal resources. Simulations of the Claude liquefacti on process and the 5-kW ammonia-based combined power/cooling cycle were devel oped to model the effects of component

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79 efficiencies and operating parameters on the maximum hydrogen production rate and system energy requirement. Additionally, a performance test of a scroll compressor was performed to gauge its effectiveness as an expander for the combined cycle. Conclusions resulting from tests an d analyses are summarized below: 1. Pre-cooling hydrogen has littl e effect on the specific liquefaction energy and is detrimental to the liquefier efficiency. 2. Pressurized electrolysis is the most effective method of reducing the energy consumed in liquefaction. 3. The total energy required to produce and liquefy hydrogen is 28.656 kW-h/lbm-H2 (63.175 kW-h/kg-H2); 86% of which is consumed during electrolysis. A maximum of 7.21 gallons (27.3 liters) per day of liqui d hydrogen can be produced from a 5-kW combined cycle. 4. The mass flows as well as the heat and work interactions of the 5-kW combined cycle scale with inverse of expander efficiency (1/e). Sixty percent e xpansion efficiency is required to extract cooling from the cycle. 5. Cooling capacity of the cycle is extremely sensitive to the vapor mass fraction of the expander inlet stream. At 2.5% water by ma ss and for perfect expansion, the cooling capacity completely diminishes. 6. Results of the performance test indicate that scroll compressors operate poorly as expanders. Low isentropic efficiencies re sult from leakage around the scroll tips. Improvements in the scroll design such as in creasing the wrap of each scroll element and using low-friction material for oilless operation would make the scroll an efficient expansion device suitable for the combined cycle.

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80 APPENDIX A COMPUTER PROGRAM FOR CYCLE SIMULATIONS Two computer programs were written to a ssist in the evaluation of thermodynamic properties and to perform cycle analyses of the Claude liquefaction cycle and the ammonia-water combined power/cooling cycle. A description of each program is given below, including portions of the source code. Claude Cycle Simulation The program was written to assist in the parametric analysis of the specific work and efficiency of the Claude cycle. A subroutine was included to evaluate the thermodynamic properties at each state point coinciding with Figure 2.6. The code has the flexibility of single point calculations or variable inputs for a parametric analysis. Thermodynamic Property Evaluation The property code incorporates portions of RGAS and PSAT, two programs written by Dr. Roger Gater (2001). Property evaluati on is carried out as a subroutine of the overall cycle simulation. The properties defi ned by user input and the listed assumptions are passed into either routine depending on the fluid condition. For saturated conditions, the pressure is defined; for superheated va por, pressure-temperature pressure-enthalpy, or pressure-entropy is input. Properties ar e then evaluated using the Redlich-Kwong gas model and returned to the main program. Cr itical properties and coefficients required by the Redlich-Kwong model are listed in Table A.1.

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81 Program Description The Claude cycle simulation program is writt en in MatLAB. It consists of three sub-routines and a data file : saturation2.m, gas_properties.m, gas_properties_base.m, gas.dat, all of which must be present for the program to operate. The program begins by reading data from the “gas.dat” file. It th en asks for user input of compressor inlet temperature and pressure; giving the option of English or SI units. From the user input and given assumptions, the thermodynamic prope rties at each state poi nt are evaluated by the “gas_properties.m” subroutine. If saturated conditions are known to exist, “saturation2.m is invoked. The key performan ce parameters of the Claude cycle are then calculated using the equations of Chapter 3. Results are output to the screen in figure form. Additional aspects of the program are described by the imbedded comments. Main Program Claude.m [gas_num gas_name R Tc Pc cpoR a b c Zc A w] = text read('GAS.dat','%f %s %f %f %f %f %f %f %f %f %f %f', 'headerlines',1); units = input('Select Units: 0 = Metric, 1 = English: '); while (units < 0) | (units > 1) units = input('\nError, Try again: '); end if units == 0 T1 = input('\nEnter compre ssor inlet temperature (K): '); P1 = input('\nEnter compressor inlet pressure (atm): '); P1 = P1 1.0132; else T1 = input('\nEnter compre ssor inlet temperature (F): '); T1 = (T1 32)*5/9 + 273.15; P1 = input('\nEnter compressor inlet pressure (atm): '); P1 = P1 1.0132; end P2 = input('\nEnter compressor discharge pressure (atm): '); P2 = P2 1.0132; eta_e = input('\nEnter expander adiabatic efficiency: '); %eta_e = .85; eta_c = input('\nEnter compressor efficiency: '); %eta_c = .75; % properties in J/g or kJ/kg P_stp = 1.0132; %bar gas = 10; % selects hydrogen gas from GAS.DAT data file Pe = P_stp; Pg = P_stp; Pf = P_stp; P7 = P_stp; P8 = P_stp;

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82 P9 = P_stp; P10 = P_stp; P10a = P_stp; PA = P_stp; PC = P_stp; P2b = P2; P3 = P2; P4 = P2; P5 = P2; HC1 = 175.1; HC2 = 312.97; %Heats of conversion kJ/kg WN2 = 1783.623; %kJ/kg-N2 Energy of LN2 liquefaction T2 = T1; T10 = T1; TC = T1; %State g and f %call saturation program routine = 1; Ps = Pg; P = Pg; [Ts,Zf,Zg,vf,vg,hfg,ufg,sfg] = saturation2(Ps,R(gas),Tc(gas),Pc(gas),Zc(gas),A(gas),w(gas)); Tg = Ts; Tf = Ts; T = Ts; gas_properties; hg = h; sg = s; hf = hg hfg; sf = sg sfg; %State 1 routine = 1; %pressure and temp specified T = T1; P = P1; gas_properties; h1 = h; s1 = s; %State 2 routine = 1; %isothermal compression T = T2; P = P2; gas_properties; h2 = h; s2 = s; %State 2b Ps = P_stp; gas = 13; %sets nitrogen properties Ts = saturation2(Ps,R(gas),Tc(gas),P c(gas),Zc(gas),A(gas),w( gas)); %calculates temperature of nitrogen gas = 10; %returns to hydrogen T2b = Ts; T = T2b; P = P2b; routine = 1; gas_properties; h2b = h; s2b = s; %State 3 T3 = T2b; test = 0; while T3 >= 70 % above critical temperat ure of hydrogen (asymptotic problems) routine = 1; T = T3; P = P3; gas_properties; h3 = h; s3 = s; %state e_s routine = 2; %test to see if saturated conditions exist P = Pe; Ps = Pe; se = s3; ss = se; gas_properties; he_s = h; Te_s = T; if Te_s <= Tf xe = (se sf)/(sfg);

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83 he_s = hf + xe*(hfg); Te_s = Tf; %isentropic temperature end if (h3 he_s) > test T3opt = T3; Te_s_opt = Te_s; delta_h_opt = h3 he_s; h3opt = h3; s3opt = s3; he_sopt = he_s; end test = h3 he_s; T3 = T3 .1; end %state e routine = 3; % pressure and enthalpy specified he = h3opt eta_e*(delta_h_opt); hh = he; P = Pe; gas_properties; Te = T; se = s; %state 4 routine = 1; T = Te; P = P1; T4 = Te; gas_properties; h4 = h; s4 = s; %state 7 and 8 routine = 1; T = Te; P = Pe; T7 = Te; T8 = Te; gas_properties; h7 = h; h8 = h; s7 = s; s8 = s; %state 10 routine = 1; T = T10; P = P10; gas_properties; h10 = h; s10 = s; %state 10a routine = 1; T = T2b; P = P10; gas_properties; h10a = h; s10a = s; %state A (saturated liquid) routine = 1; Ps = PA; P = PA; gas = 13; [Ts,Zf,Zg,vf,vg,hfg,ufg,sfg] = saturation2(Ps,R(gas),Tc(gas),Pc(gas),Zc(gas),A(gas),w(gas)); TA = Ts; gas_properties; hgA = h; sgA = s; hA = hgA hfg; sA = sgA sfg;

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84 %state C routine = 1; T = TC; P = PC; gas_properties; hC = h; sC = s; gas = 10; %return to hydrogen %Specific work, liquid yield, liquid nitrogen requirement, and figure of merit calculation X(1) = 0; step = .001 for i = 1:1/step X(i+1) = X(i) + step; y(i) = ((h10a-h2b) + eta_e*X(i)*(delta_h_opt))/(h10a hf + HC2); if y(i) >= .725*(1 X(i)) %.725 found from iterative procedure on HX 3 y(i) = .725*(1-X(i)); end z(i) = (HC1 + (h2 h10) + eta_e*X(i)*(delta_h_opt) + y(i)*(h10 hf + HC2))/(hC hA); if units == 0 W(i) = (((h2 h1) T1*(s2 s1))/eta_c eta_e*X(i)*(delta_h_opt))/3600; %work per unit mass compressed Wf(i) = W(i)/y(i) + z(i)/y(i)* WN2/3600; %work per unit mass liquefied kJ/kg W_ideal = ((hf h1) T1*(sf s1))/3600; else W(i) = (((h2 h1) T1*(s2 s1))/eta_c eta_e*X(i)*(delta_h_opt))/(2.326*3412); Wf(i) = W(i)/y(i) + z(i)/y(i)*WN2/(2.236*3412); W_ideal = ((hf h1) T1*(sf s1))/(2.326*3412); end FOM(i) = W_ideal/Wf(i); if X(i+1) >= .9 break end end z(i+1) = z(i); Wf(i+1) = Wf(i); W(i+1) = W(i); FOM(i+1) = FOM(i); y(i+1) = y(i); Xopt = (1 max(y)/.725) Ymax = max(y) Wfopt = min(Wf) FOMopt = max(FOM) zopt = min(z./y) W_ideal figure(1) plot(X,Wf) title('Work Per Unit Mass LH2 vs. Expander Mass Flow Ratio, X'); xlabel('Expander Mass Flow Ratio, X') ylabel('Wf [kW-h/lbm-LH2]')

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85 Saturation Property Evaluation Saturation2.m function [Ts,Zf,Zg,vf,vg,hfg,ufg,sfg] = saturation2(Ps,R,Tc,Pc,Zc,A,w) Psr = Ps/Pc; Tsr = (A + w*log(Psr))/(A log(Psr)); Ts = Tsr*Tc; Zfg = .824*(log(1.3/(Psr + .3)))^.467; Zf = Zc*Psr*( 1 Tsr^1.72*(1/Tsr 1)^.295); Zg = Zfg + Zf; vf = Zf*R*Ts/Ps; vg = Zg*R*Ts/Ps; hfg = R*Tc*Zfg*(A*Tsr^2*(1 + w) / (Tsr + w)^2); ufg = hfg Ps*(vg vf); sfg = hfg/Ts; %need to add warning about pressure and temperature above critical point Gas Thermodynamic Property E valuation – Gas_properties.m %P & T given cpo = R(gas)*cpoR(gas); T0 = 300; P0 = 1; if routine == 1 [v,u,h,s] = gas_properties_base(T,P,R(gas),cpo,Tc(gas),Pc(gas),a(gas),b(gas),c(gas)); end if routine == 2 T = 1.2 T0*exp(.8*ss/cpo + (R(gas)/cpo)*log(P/P0)); errorS = 1; while errorS > 1E-6 [v,u,h,s] = gas_properties_base(T,P,R(gas),cpo,Tc(gas),Pc(gas),a(gas),b(gas),c(gas)); errorS = abs(s ss)/(abs(s+ss)+1); T = T*(.8 + .2*exp((ss s)/cpo)); Tr = T/Tc; end end if routine == 3 T = 1.2*T0 + hh/cpo; errorH = 1; while errorH > 1E-6 [v,u,h,s] = gas_properties_base(T,P,R(gas),cpo,Tc(gas),Pc(gas),a(gas),b(gas),c(gas)); errorH = abs(h hh)/(abs(h + hh)+100); T = T + .5*(hh-h)/cpo; Tr = T/Tc; end end Gas properties base.m function [v,u,h,s] = gas_properties_base(T,P,R,cpo,Tc,Pc,a,b,c) alpha = .42748; beta=.08664; T0 = 300; P0 = 1;

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86 Zp = 1; errorZ = 1; Tr = T/Tc; Pr = P/Pc; tau = (T/T0) 1; while errorZ > 1E-6 vr = Zp*Tr/Pr; Z = 1 + beta/(vr beta) alpha/((vr + beta)*Tr^1.5); errorZ = abs(Z Zp)/(Z + Zp); Zp = .8*Zp + .2*Z; end v = Z*R*T/P; vr = Z*Tr/Pr; phi = 1.25*vr^.1; ud = R*Tc*phi*1.5*(alpha/beta)/Tr^.5*log((vr + beta)/vr); hd = ud + (1 Z)*R*T; sd = R*(ud/(R*Tc*Tr) + log(vr/(vr beta)) (alp ha/beta)/Tr^1.5*log((vr + beta)/vr) log(Z)); hig = cpo*T0*(tau + a*tau^2/2 + b*tau^3/3 + c*tau^4/4); uig = hig R*T; sig0 = cpo ((1 a + b c)*log(tau + 1) + (a b + c)*tau + (b c)*tau^2/2 + c*tau^3/3); sig = sig0 R*log(P/P0); u = uig ud; h = hig hd; s = sig sd; Table A.1. Critical properties and coefficients contained within the “gas.dat” file HydrogenNitrogen R4.1240.2968 Tc33.3126.2 Pc1333.9 Cpo/R3.473.503 a2.49E-022.98E-03 b-6.75E-033.50E-02 c2.00E-03-5.86E-03 Zc0.360.291 A4.9255.64 w0.14150.0039 Ammonia-Based Combined Pow er/Cooling Cycle Simulation The program was written to assist in the pa rametric analysis of the energy transfer and cooling capacity dependence of the combin ed cycle on expander isentropic efficiency and ammonia vapor mass fraction.

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87 Thermodynamic Property Evaluation Thermodynamic properties of the ammoniawater mixture and pure ammonia vapor are evaluated using subroutines adapted fr om a program developed by Tamm (2003). The evaluation method is based on a Gibbs free en ergy approach incorporating experimental correlations. A detailed desc ription of the evaluation met hod and the coefficients used for the calculations are outlined by Tamm (2003). Program Description The combined cycle simulation program consists of a main program written in MatLAB, “combined cycle.m” and fi ve property evalua tion subroutines: ammonia_water.m, PTX.m, bubble_dew.m, cr itical_properties.m, and hsv_properties.m. The main program accepts user input values of cycle high and low pressure as well as the absorber and boiler temperatures. Each state point corresponding to Figure 2.7 is defined from these inputs and by the assumptions listed in Chapter 3. The program evaluates the thermodynamic properties at each point and uses these values to calculate the mass flows, mass fractions, energy transfers, and efficien cies given by Equations 3.16 thru 3.25. Results are output to text file named result s.txt and displayed in several graphs. A sample output of the program is given in A ppendix B. Additional as pects of the program are described by the imbedded comments. Combined Cycle Main Program – Combined_cycle.m %This program calculates the state points, work and heat exchanges, and flow rates %of the combined cycle for varied turbine isentropic efficiencies using the user %inputs of high pressure,low pressure, boile r temperature, and absorber temperature. %The turbine isentropic efficiency is varied from 10 to 100% to study its effect on %the overall cycle efficiency. % %Results are output to results.txt and graphs % %This program can easily be adapted for different operating conditions by %adjusting values in the assumptions section. %

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88 %Sub-programs necessary for op eration: ammonia_water.m, PTx.m, hsv_properties.m, %critical_props.m, bubble_dew.m, bubble_dew_base.m clear all; clc; cycle = 1; %signifies whether inputs are for cycle or individual states fprintf('\n***********************************************\n'); fprintf('* Combined Cy cle Analysis *\n'); fprintf('* Analysis code written by Robert Reed *\n'); fprintf('* Property code written in C++ by Gunner Tamm *\n'); fprintf('* Adapted to MatL AB by Robert Reed *\n'); fprintf('* Septembe r 12, 2004 *\n'); fprintf('***********************************************\n'); fprintf('IMPORTANT: This program consists of six sub-routines which must be present:\n'); fprintf('\tammonia_water.m, bubble_dew_base.m, bubble_dew.m, critical_props.m,\n\thsv_properties.m, and PTx.m\n\n'); P_low = input('Enter the cycle low pressure (psia): '); P_low = P_low/14.504; P_high = input('Enter the cycle high pressure (psia): '); P_high = P_high/14.504; while P_high <= P_low fprintf('High pressure must be greater than low pressure!!!\n'); P_high = input('Try again: '); P_high = P_high/14.504; end Tabs = input('Enter absorber temperature (F): '); Tabs = (Tabs 32)/1.8 + 273.15; Tboil = input('Enter boiler temperature (F): '); Tboil = (Tboil 32)/1.8 + 273.15; while Tboil <= Tabs fprintf('Boiler temperature must be greater than absorber temperature!!!\n'); Tboil = input('Try again: '); Tboil = (Tboil 32)/1.8 + 273.15; end %Future additions will include cooling and heating hot water flow rates. % Tcws = input('Enter cooling wate r source temperature (F): '); % Tcwr = input('Enter cooling wa ter return temperature (F): '); % Thws = input('Enter heating hot water source temperature (F): '); % Tcwr = input('Enter heating hot water source temperature (F): '); %assumptions x7 = 1; %pure ammonia vapor exiting the rectifier %x7 = input('\nEnter mass fraction ammonia entering turbine: '); Elec = 5; %Electrical output of the generator [kW]. etaG = .75; %Generator efficiency eps = .85; %recovery HE effectiveness WT = Elec/etaG; %Define all states by calling ammonia_water.m %state 1 P = P_low; T = Tabs; option = 8; %option sets the sub-routine used in ammonia_water.m ammonia_water; P1 = P; T1 = T; h1 = hL; s1 = sL; v1 = vL; x1 = xb; %return results %state 2 P = P_high; T2 = T1; x = x1; option = 1; ammonia_water; P2 = P; T2 = T; h2 = hm; s2 = sm; v2 = vm; x2 = x;

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89 %state 4 P = P_high; T = Tboil; option = 8; ammonia_water; P4 = P; T4 = T; h4 = hL; s4 = sL; v4 = vL; x4 = xb; %could also have the actual vapor exit state %state 5 P = P_high; x = x4; option = 1; T = T4 eps*(T4 T2); %assuming equal specific heats ammonia_water; T5 = T; P5 = P; h5 = hm; s5 = sm; v5 = vm; x5 = x; %state 6 P = P_low; h = h5; x = x4; option = 3; ammonia_water; T6 = T; P6 = P; h6 = h; s6 = sm; v6 = vm; x6 = x; %state 9 T = 302.594444444; %assumed to be 85F for air/hydrogen cooling P = P_low; x = x7; option = 1; ammonia_water; T9 = T; P9 = P; h9 = hm; s9 = sm; v9 = vm; x9 = x; %state 7 P = P_high; T = Tboil; x = x7; option = 1; ammonia_water; P7 = P; T7 = T; h7 = hm; s7 = sm; v7 = vm; %state 8s Imaginary state obtained from isentropic expansion P = P_low; s = s7; x = x7; option = 4; ammonia_water; P8s = P; T8s = T; h8s = hm; s8s = s; v2s = v2; x8s = x; %state 8 etaT(1) = .1; %initial value of the turbine efficiency step = .01; %step change in lo op of the turbine efficiency for n = 1:1/step P = P_low; x = x7; h = h7 etaT(n)*(h7 h8s); option = 3; ammonia_water; P8 = P8s; T8(n) = T; h8(n) = h; s8(n) = sm; v8(n) = vm; x8 = x; %Calculate solution mass flow rates mT(n) = WT / (h7 h8(n)); %Turbine mass flow mS(n) = mT(n)*((x7 x4)/(x1 x4)); % Strong solution mass flow rate mW(n) = mS(n) mT(n); %Weak solution mass flow rate %state 3 P = P_high; x = x1; option = 1; T = eps*mW(n)*(T4 T2)/mS(n) + T2; %assuming equal specific heats ammonia_water; T3 = T; P3 = P; h3 = hm; s3 = sm; v3 = vm; x3 = x; %Heat and work flows

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90 Qin(n) = mW(n)*h4 + mT(n)*h7 mS(n)*h3; %heat input to vapor generator Qout(n) = mW(n)*h6 + mT(n)*h9 mS(n)*h1; %heat rejected from absorber Qc(n) = mT(n)*(h9 h8(n)); %Cooling capacity WP(n) = mS(n)*(h2-h1); %Pump work if Qc(n) < 0 % This step disallows negative cooling capacity Qc(n) = 0; end COP = 3; %typical value eta_cycle(n) = (WT WP(n) + Qc (n)/COP) / Qin(n); %cycle efficiency etaT(n+1) = etaT(n) + step ; %set new value for next loop end %Set vector lengths equal, last value is ignored Qin(n+1) = Qin(n); Qout(n+1) = Qout(n); WP(n+1) = WP(n); Qc(n+1) = Qc(n); eta_cycle(n+1) = eta_ cycle(n); mS(n+1) = mS(n); mW(n+1) = mW(n); mT(n+1) = mT(n); results = fopen('r esults.txt','w'); fprintf(results,'***************************\n'); fprintf(results,'* Cycle Analysis Resu lts *\t\t Created: %s \n',datestr(now)); fprintf(results,'***************************\n\n\n'); fprintf(results,'Assumptions:\t Saturated conditions at states 1 and 4.\n'); fprintf(results,' \t Component pressure losses are negligible.\n'); fprintf(results,' \t Equal weak and strong solution specific heats.\n'); fprintf(results,' \t Superheat er temperature equal to boiler temperature.\n'); fprintf(results,' \t Mass fraction of ammonia in the rectifier exit stream, x7 = %g\n',x7); fprintf(results,' \t Evaporator exit temp erature = %g F (%.2f C)\n',((T9-273.15)*1.8+32),T9-273.15); fprintf(results,' \t Elect ric generator efficiency = %g%%\n',etaG*100); fprintf(results,' \t Recove ry heat exchanger effectiveness = %g\n',eps); fprintf(results,' \t Elect ric generator output = %g kW\n\n\n',Elec); fprintf(results,'User Inputs:\t Absorber temperature = %g F (%g C)\n',(Tabs-273. 15)*1.8+32,Tabs-273.15); fprintf(results,' \t Boiler temperat ure = %g F (%g C)\n',(Tboil-273.15)*1.8+32,Tboil-273.15); fprintf(results,' \t System low pr essure = %g psia (%g bar)\n',P_low*14.504,P_low); fprintf(results,' \t System high pressure = %g psia (%g bar)\n',P_high*14.504,P_high); fprintf(results,'__________________________________________________________________________ ____________________________________\n\n'); fprintf(results,'\t State 1 \t\t\t\t\t\t State 2 \t\t\t\t\t\t State3\n\n'); fprintf(results,' T = %g F (%.2f C) \t\t\t\t T = %g F (%.2f C) \t\t\t\t T = %g F (%.2f C)\n',(T1273.15)*1.8+32,T1-273.15,(T2-273.15)*1.8+32, T2-273.15,(T3-273.15)*1.8+32,T3-273.15); fprintf(results,' P = %g psia (%.2f bar) \t\t\t P = %g psia (%.2f bar) \t\t\t P = %g psia (%.2f bar)\n',P1*14.504,P1,P2*14 .504,P2,P2*14.504,P2); fprintf(results,' h = %g BTU/lbm (%.2f kJ/kg) \t h = %g BTU/lbm (%.2f kJ/kg) \t h = %g BTU/lbm (%.2f kJ/kg)\n',h1/2.326,h1,h2/2.326,h2,h3/2.326,h3); fprintf(results,' s = %g BTU/lbm-R (%.4f kJ/kg-K) \t s = %g BTU/lbm-R (%.4f kJ/kg-K) \t s = %g BTU/lbm-R (%.4f kJ/kg-K)\n\n',s1/4.1 868,s1,s2/4.1868,s2,s3/4.1868,s3); fprintf(results,'\t State 4 \t\t\t\t\t\t State 5 \t\t\t\t\t\t State6\n\n'); fprintf(results,' T = %g F (%.2f C) \t\t\t\t T = %g F (%.2f C) \t\t\t\t T = %g F (%.2f C)\n',(T4273.15)*1.8+32,T4-273.15,(T5-273.15)*1.8+32, T5-273.15,(T6-273.15)*1.8+32,T6-273.15); fprintf(results,' P = %g psia (%.2f bar) \t\t\t P = %g psia (%.2f bar) \t\t\t P = %g psia (%.2f bar)\n',P4*14.504,P4,P5*14 .504,P5,P6*14.504,P6); fprintf(results,' h = %g BTU/lbm (%.2f kJ/kg) \t h = %g BTU/lbm (%.2f kJ/kg) \t h = %g BTU/lbm (%.2f kJ/kg)\n',h4/2.326,h4,h5/2.326,h5,h6/2.326,h6); fprintf(results,' s = %g BTU/lbm-R (%.4f kJ/kg-K) \t s = %g BTU/lbm-R (%.4f kJ/kg-K) \t s = %g BTU/lbm-R (%.4f kJ/kg-K)\n\n',s4/4.1 868,s4,s5/4.1868,s5,s6/4.1868,s6); fprintf(results,'\t State 7 \t\t\t\t\t\t State 8s \t\t\t\t\t\t State9\n\n');

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91 fprintf(results,' T = %g F (%.2f C) \t\t\t\t T = %g F (%.2f C) \t\t\t T = %g F (%.2f C)\n',(T7273.15)*1.8+32,T7-273.15,(T8s-273.15)*1.8+32,T8s-273.15,(T9-273.15)*1.8+32,T9-273.15); fprintf(results,' P = %g psia (%.2f bar) \t\t\t P = %g psia (%.2f bar) \t\t\t P = %g psia (%.2f bar)\n',P7*14.504,P7,P8*14 .504,P8,P9*14.504,P9); fprintf(results,' h = %g BTU/lbm (%.2f kJ/kg) \t h = %g BTU/lbm (%.2f kJ/kg) \t h = %g BTU/lbm (%.2f kJ/kg)\n',h7/2.326,h7,h8s/2.326,h8s,h9/2.326,h9); fprintf(results,' s = %g BTU/lbm-R (%.4f kJ/kg-K) \t s = %g BTU/lbm-R (%.4f kJ/kg-K) \t s = %g BTU/lbm-R (%.4f kJ/kg-K)\n',s7/4.1868,s7,s8s/4.1868,s8s,s9/4.1868,s9); fprintf(results,'__________________________________________________________________________ ____________________________________\n\n'); fprintf(results,'Weak solution mass fraction, xW = %g \t Strong solution mass fraction, xS = %g\n',x4,x1); fprintf(results,'__________________________________________________________________________ ____________________________________\n\n'); fprintf(results,'Turbine shaft work output, WT = %g kW\n\n',WT); fprintf(results,'Turbine adiabatic efficien cy varied from %g%% to %g%% in %g%% increments\n\n',etaT(1)*100,etaT(n)*100,step*100); fprintf(results,' \t\t\t State 8\n\n'); fprintf(results,' Turbine efficiency \t\t T8 \t\t\t\t\t h8 \t\t\t\t\t\t s8\n'); if (1/step) <= 20 for n = 1:1/step fprintf(results,' \t %g \t\t\t %g F (%.2f C) \t %g BTU/lbm (%.2f kJ/kg) \t %g BTU/lbm-R (%.4fkJ/kgK)\n',etaT(n),(T8(n)-273.15)*1.8+32,T8(n)-273.15,h8(n)/2.326,h8(n),s8(n)/4.1868,s8(n)); end fprintf(results,'\n\n \t\t\t Energy transfers and mass flow rates\n\n'); fprintf(results,' Turbine efficiency \t eta_cycle \t\t\t\t WP \t\t\t\t Qin \t\t\t\t\t Qout \t\t\t\t Qc\n'); for n = 1:1/step fprintf(results,' \t %g \t\t\t %.5f \t\t %g BT U/h (%.2f kW) \t %g BTU/h (% .2f kW) \t %g BTU/h (%.2f kW)\n',etaT(n),eta_cycle(n),WP(n)*3412,WP(n),Qin(n)*3412,Qin(n),Qout(n)*3412,Qout(n)); end fprintf(results,'\n\n Turbine efficiency \t\t\t Qc \t\t\t mS \t\t\t\t\t mW \t\t\t\t\t mT\n'); for n = 1:1/step fprintf(results,' \t %g \t\t\t %.4f BTU/h (% .2f kW) \t %g lbm/h (%.4f kg/s) \t %g lbm/h (%.4f kg/s) \t %g lbm/h (%.4f kg/s)\n',etaT(n),Qc(n)*3412,Qc(n),mS(n)*7936.56,mS(n),mW(n)*7936.56,mW(n),mT(n)*7936.56,mT(n)); end else fprintf(results,'Too much data to display...see output plots.'); end fclose(results); fprintf('\n\nResults of analysis were written to results.txt'); figure(1) plot(etaT,WP) title('Pump work') xlabel('Expander Isentropic Efficiency') ylabel('Pump Work [kW]') figure(2) plot(etaT,Qin.*3412,etaT,Qout.*3412) title('Qin') xlabel('Expander Isentropic Efficiency') ylabel('Heat [Btu/hr]') figure(3)

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92 plot(etaT,eta_cycle) title('cycle efficiency') xlabel('Expander Isentropic Efficiency') ylabel('Overall Cycle Thermal Efficiency') figure(4) plot(etaT,Qc.*3412) title('Cooling Capacity') xlabel('Expander Isentropic Efficiency') ylabel('Cooling Capacity [Btu/hr]') figure(5) plot(etaT,mW.*7936.56,etaT,mS.*7936.56,etaT,mT.*7936.56) title('Mass flows') xlabel('Expander Isentropic Efficiency') ylabel('Mass Flow Rate [lbm/hr]') Ammonia Water Code – Ammonia_water.m format long global a b Ai Aij Ci Cij global TB PB R Aa Aw Ba Bw Ca Cw Da Dw E hroaL hrowL hroaG hrowG sroaL srowL sroaG srowG Troa Trow Proa Prow %Input constants from data files [Aa Aw Ba Bw Ca Cw Da Dw E] = textread('gibbs_coefficients.dat','% f%f%f%f%f%f%f%f%f','headerlines',1); [hroaL hrowL hroaG hrowG sroaL srowL sroaG srowG Troa Trow Proa Prow] = textread('reduced_props.dat','%f%f%f %f%f%f%f%f%f%f%f%f','headerlines',1); [a Ai Aij(:,1) Aij(:,2) Aij(:, 3) Aij(:,4) b Ci Cij(:,1) Cij(:,2) Cij(:,3) Cij(:,4) Cij(:,5) Cij(:, 6) Cij(:,7) Cij(:,8) Cij(:,9) Cij(:,10)] = textread('Bdc_coefficients.dat ','%f%f%f%f%f%f%f%f%f%f%f%f%f%f%f%f%f%f','headerlines',1); TB = 100; PB = 10; R = 8.314; global istate %returns mixture condition (superheated vapor, etc.) if option == 1 if cycle == 0; P = input('Pressure (bar): '); T = input('Temperature (K): '); x = input('Mass fraction: '); end [hm,sm,vm] = PTx(P,T,x); %calls PTx.m elseif option == 2 x = .5; incr2=.01; limit=.000001; if cycle == 0 P = input('Pressure (bar): '); T = input('Temperature (K): '); v = input('Specific volume (m^3/kmol): '); end [hm,sm,vm]= PTx(P,T,x); while abs(v vm)>limit n = 1; while (vm > v) if n >= 11

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93 incr2 = incr2*10; n=1; end x = x incr2; if x < 0 fprintf('Mixture not possible!!'); break; end [hm,sm,vm] = PTx(P,T,x); n = n + 1; end if x < 0 break; end incr2 = incr2/10; n=1; while (vm < v) if n >= 11 incr2 = incr2*10; n = 1; end x = x + incr2; if x > 1 fprintf('Mixture not possible!!'); break; end [hm,sm,vm] = PTx(P,T,x); n = n + 1; end if x > 1 break; end incr2 = incr2/10; end elseif option == 3 T = 400; incr2 = 10; if cycle == 0; P = input('Pressure (bar): '); h = input('Enthalpy (kJ/kg): '); x = input('Mass fraction: '); end [hm,sm,vm] = PTx(P,T,x); while (abs(h hm)) > .01 while (hm > h) T = T incr2; [hm,sm,vm] = PTx(P,T,x); end incr2 = incr2/10; while (hm < h) T = T + incr2; [hm,sm,vm] = PTx(P,T,x); end incr2 = incr2/10; end elseif option == 4 T = 400; incr2 = 10; if cycle == 0 P = input('Pressure (bar): '); s = input('Entropy (kJ/kg-K): '); x = input('Mass fraction: '); end [hm,sm,vm] = PTx(P,T,x); while (abs(s sm)) > .001

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94 while (sm > s) T = T incr2; [hm,sm,vm] = PTx(P,T,x); end incr2 = incr2/10; while (sm < s) T = T + incr2; [hm,sm,vm] = PTx(P,T,x); end incr2 = incr2/10; end elseif option == 5 T = 300; incr2 = 1; limit = .000001; if cycle == 0; P = input('Pressure (bar): '); v = input('Specific volume (m^3/kmol): '); x = input('Mass fraction: '); end [hm,sm,vm] = PTx(P,T,x); while (abs(v vm)) > limit n = 1; while (vm > v) if n >= 11 incr2 = incr2*10; n = 1; end T = T incr2; [hm,sm,vm] = PTx(P,T,x); n = n + 1; end incr2 = incr2/10; n = 1; while (vm < v) if n >= 11 incr2 = incr2*10; n = 1; end T = T + incr2; [hm,sm,vm] = PTx(P,T,x); n = n + 1; end incr2 = incr2/10; end elseif option == 6 P = 1; incr2 = .01; limit = .000001; if cycle == 0 T = input('Temperature (K): '); v = input('Specific volume (m^3/kmol): '); x = input('Mass fraction: '); end [hm,sm,vm] = PTx(P,T,x); while (abs(v vm)) > limit while (vm < v) P = P incr2; [hm,sm,vm] = PTx(P,T,x); end incr2 = incr2/10; n = 1; while (vm > v) if n >= 11 incr2 = incr2*10; n = 1; end P = P + incr2; [hm,sm,vm] = PTx(P,T,x); n = n + 1; end incr2 = incr2/10; end elseif option == 7

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95 if cycle == 0 P = input('Pressure (bar): '); x = input('Mass fraction: '); end [Tb,Td] = bubble_dew(P,x); Pr = P/PB; M = 18.015*17.031/((1-x)*17.031+x*18.015); y = x*18.015/(x*18.015 + (1-x)*17.031); [hLm, crap, sLm, crap, vLm, crap] = hsv_properties(Tb/TB,Pr,y); [crap, hgm, crap, sgm, crap vgm] = hsv_properties(Td/TB,Pr,y); hL = hLm/M; sL = sLm/M; vL = vLm/M; hg = hgm/M; sg = sgm/M; vg = vgm/M; elseif option == 8 if cycle == 0; P = input('Pressure (bar): '); T = input('Temperature (K): '); end Tr = T/TB; Pr = P/PB; choice = 3; bubble_dew_base; if xb < 0 xb = 0; hg = 0; hL = 0; sg = 0; sL = 0; vg = 0; vL = 0; fprintf('Not a saturated condition!!'); break; end x = xb; M = 18.015*17.031/((1-x)*17.031+x*18.015); y = x*18.015/(x*18.015 + (1-x)*17.031); if xd > 1 xd = 1; hg = 0; hL = 0; sg = 0; sL = 0; vg = 0; vL = 0; fprintf('Not a saturated condition!!'); end yd = xd*18.015/(xd*18.015 + (1-xd)*17.031); Md = 18.015*17.031/((1-xd)*17.031+xd*18.015); [hLm, crap, sLm, crap, vLm, crap] = hsv_properties(Tr,Pr,y); [crap, hgm, crap, sgm, crap, vgm] = hsv_properties(Tr,Pr,yd); hL = hLm/M; sL = sLm/M; vL = vLm/M; hg = hgm/Md; sg = sgm/Md; vg = vgm/Md; elseif option == 9 if cycle == 0 T = input('Temperature (K): '); x = input('Mass fraction: '); end Tr = T/TB; M = 18.015*17.031/((1-x)*17.031+x*18.015); y = x*18.015/(x*18.015 + (1-x)*17.031); choice = 2; bubble_dew_base; [hLm, crap, sLm, crap, vLm, crap] = hsv_properties(Tr,Pb/PB,y); [crap, hgm, crap, sgm, crap vgm] = hsv_properties(Tr,Pd/PB,y); hL = hLm/M; sL = sLm/M; vL = vLm/M; hg = hgm/M; sg = sgm/M; vg = vgm/M; end %Enter code to calculate all bubble and dew point properties for display if istate <= 3 choice = 1; bubble_dew_base; Tbb = Tb; Tdd = Td; choice = 2; bubble_dew_base; choice = 3; bubble_dew_base; end

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96 Pressure, Temperature, and Mass Fraction Evaluation – PTX.m %input: Pressure (bar), temperature (K), and mass fraction %output: enthalpy (kJ/kg), entropy (kJ/kg-K), and specific volume (m^3/kg) function [hm,sm,vm] = PTx(P,T,x) global qm istate TB PB Tr Pr M y Tr = T/TB; Pr = P/PB; M = 18.015*17.031/((1-x)*17.031+x*18.015); %molecular weight y = x*18.015/(x*18.015 + (1-x)*17.031); %mole fraction [Tb,Td] = bubble_dew(P,x); [hLm, hgm, sLm, sgm, vLm, vgm] = hsv_properties(Tr,Pr,y); if T < Tb %compressed liquid istate = 1; %?????? hm = hLm/M; sm = sLm/M; vm = vLm/M; xNH3v = 0; xH2Ov = 0; xNH3L = x; xH2OL = 1-x; elseif T > Td %superheated vapor istate = 3;%????? hm = hgm/M; sm = sgm/M; vm = vgm/M; xNH3v = x; xH2Ov = 1-x; xNH3L = 0; xH2OL = 0; else %liquid-vapor mixture istate = 2; %????? choice = 3; bubble_dew_base; qm = (x xb)/(xd xb); % quality of mixture yb = xb*18.015/(xb*18.015 + (1-xb)*17.031); yd = xd*18.015/(xd*18.015 + (1-xd)*17.031); Mb = 18.015*17.031/((1-xb)*17.031+xb*18.015); Md = 18.015*17.031/((1-xd)*17.031+xd*18.015); [hLmb, hgmb, sLmb, sgmb, vLmb, vgmb] = hsv_properties(Tr,Pr,yb); [hLmd, hgmd, sLmd, sgmd, vLmd, vgmd] = hsv_properties(Tr,Pr,yd); hm = (1-qm)/Mb*hLmb + qm/Md*hgmd; sm = (1-qm)/Mb*sLmb + qm/Md*sgmd; vm = (1-qm)/Mb*vLmb + qm/Md*vgmd; xNH3v = (x xb)/(xd xb)*xd; xH20v = (x xb)/(xd xb)*(1-xd); xNH3L = (1-(x xb)/(xd xb))*xb; xH20v = (1-(x xb)/(xd xb))*(1-xb); end Bubble and Dew Point Property Evaluation – Bubble_dew.m %input: pressure (bar) and mass fraction %output: Bubble and dew po int temperatures (K) function [Tb,Td] = bubble_dew(P,x) global Ai Aij Ci Cij %empirical constants from ammonia_water.m [Tc,Pc] = critical_props(x); %calculate critical temperature and pressure sum2 = 0; for i = 1:7 sum1 = 0;

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97 for j = 1:10 sum1 = sum1 + Cij(i,j)*x^j; end sum2 = sum2 + (Ci(i) + sum1)*(log(Pc/P))^i; end Tb = Tc sum2/1.8; %unit conversion -bubble point temperature (K) sum2 = 0; for i = 1:6 sum1 = 0; for j = 1:4 sum1 = sum1 + Aij(i,j)*(log(1.0001-x))^j; end sum2 = sum2 + (Ai(i) + sum1)*(log(Pc/P))^i; end Td = Tc sum2/1.8; %unit conversion -dew point temperature (K) Critical Property Evaluation – Critical_properties.m %input: mass fraction of ammonia in mixture %output: critical temperature (K) and pressure (bar) function [Tc,Pc] = critical_props(x); Tcw = 1165.14 ; Pcw = 3206.79; %critical properties of water (R and psia) global a b Tc Pc sum1 = 0; i = 1; while (i <= 4) sum1 = sum1 + (a(i)*x^i); i = i + 1; end Tc = (Tcw sum1)/1.8; %convert from R to K sum1 = 0; i = 1; while (i <= 8) sum1 = sum1 + b(i)*x^i; i = i + 1; end Pc = Pcw*exp(sum1)/14.504; %convert from psia to bar Enthalpy, Entropy, and Specific Volume Evaluation – HSV.m %input: Reduced temperature, reduced pressure, and mole fraction %output: enthalpy, entropy, and specifi c volume for liquid and gas mixures function [hLm, hgm, sLm, sgm, vLm, vgm] = hsv_properties(Tr,Pr,y) %input empirical constants global TB PB R Aa Aw Ba Bw Ca Cw Da Dw E hroaL hrowL hroaG hrowG sroaL srowL sroaG srowG Troa Trow Proa Prow hLw = -R*TB*(-hrowL + Bw(1)*(Trow Tr) + Bw(2)/2*(Trow^2 Tr^2) + Bw(3)/3*(Trow^3 Tr^3) (Aw(1) + Aw(4)*Tr^2)*(Pr-Prow) Aw(2)/2*(Pr^2 Prow^2)); hLa = -R*TB*(-hroaL + Ba(1)*(Troa Tr) + Ba(2)/2*(Troa^2 Tr^2) + Ba(3)/3*(Troa^3 Tr^3) (Aa(1) + Aa(4)*Tr^2)*(Pr-Proa) Aa(2)/2*(Pr^2 Proa^2)); hgw = -R*TB*(-hrowG + Dw(1)*(Trow Tr) + Dw(2)/2*(Trow^2 Tr^2) + Dw(3)/3*(Trow^3 Tr^3) – Cw(1)*(Pr Prow) 4*Cw(2)*(Pr*Tr^-3 Prow*Trow^-3) 12*Cw(3)*(Pr*Tr^-11 Prow*Trow^-11) –

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98 4*Cw(4)*(Pr^3*Tr^-11 Prow^3*Trow^-11)); hga = -R*TB*(-hroaG + Da(1)*(Troa Tr) + Da(2)/2*(Troa^2 Tr^2) + Da(3)/3*(Troa^3 Tr^3) Ca(1)*(Pr Proa) 4*Ca(2)*(Pr*Tr^-3 Proa*Troa^-3) 12*Ca(3)*(Pr*Tr^-11 Proa*Troa^-11) 4*Ca(4)*(Pr^3*Tr^-11 Proa^3*Troa^-11)); hE = -R*TB*y*(1-y)*(-E(1) E(2)*Pr 2*E(5)/Tr 3*E(6)*Tr^-2 + (2*y 1)*(-E(7) E(8)*Pr 2*E(11)/Tr 3*E(12)*Tr^-2) + (2*y 1)^2*(-E(13) E(14)*Pr 2*E(15)/Tr 3*E(16)*Tr^-2)); hLm = y*hLa + (1-y)*hLw + hE; hgm = y*hga + (1-y)*hgw; sLw = -R*(-srowL Bw(1)*log(Tr/Trow) + Bw(2)*(Trow Tr) + Bw(3)/2*(Trow^2 Tr^2) + (Pr Prow)*(Aw(3) + 2*Aw(4)*Tr)); sLa = -R*(-sroaL Ba(1)*log(Tr/Troa) + Ba(2)*(Troa Tr) + Ba(3)/2*(Troa^2 Tr^2) + (Pr Proa)*(Aa(3) + 2*Aa(4)*Tr)); if Pr == 0 sgw = 0; sga = 0; else sgw = -R*(-srowG Dw(1)*log(Tr/Trow) + Dw(2)*(Trow Tr) + Dw(3)/2*(Trow^2 Tr^2) + log(Pr/Prow) 3*Cw(2)*(Pr*Tr^-4 Prow*Trow^-4) 11*Cw(3)*(Pr*Tr^-12 Prow*Trow^-12) 11/3*Cw(4)*(Pr^3*Tr^-12 Prow^3*Trow^-12)); sga = -R*(-sroaG Da(1)*log(Tr/Troa) + Da(2)*(Troa Tr) + Da(3)/2*(Troa^2 Tr^2) + log(Pr/Proa) 3*Ca(2)*(Pr*Tr^-4 Proa*Troa^-4) 11*Ca(3)*(Pr*Tr^-12 Proa*Troa^-12) 11/3*Ca(4)*(Pr^3*Tr^-12 Proa^3*Troa^-12)); end sE = -R*(1 y)*y*(E(3) + E(4)*Pr E(5)*Tr^-2 2*E(6)*Tr^-3 + (2*y 1)*(E(9) + E(10)*Pr E(11)*Tr^2 2*E(12)*Tr^-3) + (2*y 1)^2*(-E(15)*Tr^-2 2*E(16)*Tr^-3)); if y==0 | y==1 smix = 0; else smix = -R*(y*log(y) + (1-y)*log(1-y)); end sLm = y*sLa + (1-y)*sLw + sE + smix; sgm = y*sga + (1-y)*sgw + smix; vLw = R*TB/PB*(Aw(1) + Aw(3)*Tr + Aw(4)*Tr^2 + Aw(2)*Pr); vLw = vLw/100; vLa = R*TB/PB*(Aa(1) + Aa(3)*Tr + Aa(4)*Tr^2 + Aa(2)*Pr); vLa = vLa/100; if Pr == 0 vgw = 0; vga = 0; else vgw = R*TB/PB*(Tr/Pr + Cw(1) + Cw(2)*Tr^-3 + Cw(3)*Tr^-11 + Cw(4)*Pr^2*Tr^-11); vgw = vgw/100; vga = R*TB/PB*(Tr/Pr + Ca(1) + Ca(2)*Tr^-3 + Ca(3)*Tr^-11 + Ca(4)*Pr^2*Tr^-11); vga = vga/100; end vE = R*TB/PB*y*(1-y)*(E(2) + E(4)*Tr + (2*y 1)*(E(8) + E(10)*Tr) + (2*y 1)^2*E(14)); vE = vE/100; vLm = y*vLa + (1-y)*vLw + vE; vgm = y*vga + (1-y)*vgw;

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99 APPENDIX B CYCLE SIMULATION OUTPUT Claude Cycle Simulation Results TemperaturePressureEnthalpy*Entropy* KbarkJ/kgkJ/kg-K 1299.8025.337.500069-13.367272299.8040.52814.621-15.320292b77.6040.528-3242.375-35.44997370.0040.528-3368.811-37.18702e31.861.0132-3802.249-31.92948431.8640.528-4341.877-57.91807731.861.0132-3802.246-31.92956831.861.0132-3802.246-31.9295610a77.601.0132-3152.423-19.1928410299.801.0132-2.375333-0.064529g20.331.0132-3980.47-39.02225L20.331.0132-4443.23-61.78849 A 77.601.0132-436.1071-4.032287C299.801.0132-0.603545-0.005793State *Raw results not adjusted to proper significant figures

PAGE 116

100 Combined Cycle Simulation Results

PAGE 117

101

PAGE 118

102

PAGE 119

103 APPENDIX C EXPERIMENTAL COMPONENT LIST The major components used in the scroll e xpander performance test are listed with technical specifications where applicable. Basic tubing and fittings are not included. Scroll compressor Description: modified expansion device Manufacturer: Sanden International (USA) Inc., Wylie, TX Specifications: model TRS-90; displace ment 85.7 cc/rev; max speed 9000 RPM Compressor Description: compressed air source Manufacturer: Puma Air Compressors Specifications: 5 hp; 230 Vac input; ta nk 60 gallons; capacity 15.7 cfm Power supply Description: DC power supply for expander clutch Manufacturer: BK Precision, Yorba Linda, CA Specifications: 120 Vac input; 0 – 30 Vdc output Pony brake Description: shaft power measurement Specifications: moment arm length 14.125” Thermocouple Description: temperature measurement Manufacturer: Omega Engineering, Stamford, CT Quantity: 2 Specifications: T-type copper-constantan; accuracy 2F; grounded junction Thermocouple analog converter Description: signal conditioner Manufacturer: Omega Engineering, Stamford, CT Quantity: 2 Specifications: model TA C80B-T; range -4 – 572 F

PAGE 120

104 Multimeter Description: meter used to read thermocouple voltage Manufacturer: Fluke Corporation, Everett, WA Quantity: 2 Specifications: 0.01 mV resolution Pressure gauge Description: pressure measurement Manufacturer: Campbell Hausfeld, Harrison, OH Quantity: 2 Specifications: model IFA112; range 0 – 160 psig; resolution 1 psig Scale Description: moment arm force measurement Manufacturer: Pelouze, Bridgeview, IL Specifications: capacity 5lb; resolution 0.5 ounce; calibrated w/ standard masses Tachometer Description: shaft rotational speed measurement

PAGE 121

105 LIST OF REFERENCES Barron, R. F., 1985, Cryogenic Systems, Oxford University Press, New York. Casper, M. S., 1978, Hydrogen Production by Electrolysis, Thermal Decomposition and Unusual Techniques, Noyer Data Corporation, Park Ridge, NJ Chevron U.S.A. Inc., 1998, “Diesel Fuel Re fining and Chemistry,” San Ramon, CA. http://www.chevron.com/prodserv/fuel s/bulletin/diesel/L2_4_6_rf.htm, last accessed: 11/2004. Clean Energy Research Center, 2003, “Journey to Sustainable Energy: The H2 Solution,” University of South Florida, Tampa, FL. http://www.cerc.eng.usf.edu, last accessed: 11/2004. Copeland Corp., 2001, “Scroll Compressor Technology and Air Conditioning, Heat Pump, and Refrigerat ion Applications,” Vi Ibero-American Congress of Air Conditioning and Refrigeration, Available at http://www.copelandcorp.com/americas/news/news004.htm/scroll-english.pdf Drnevich, R., 2003, “Hydrogen Delive ry: Liquefaction and Compression,” Hydrogen Delivery Workshop Proceedings, U.S. Department of Energy – Energy Efficiency and Renewable Energy, Washington, DC. Energy Information Administration, 2003, “Annual Energy Review 2003,” DOE/EIA – 0384(2003), Washington, DC. http://www.eia. doe.gov/emeu/aer/contents.html, last accessed: 11/2004. Energy Information Administration, 2004, “International Energy Outlook 2004,” Washington, DC. Available at http://www. eia.doe.gov/oiaf/ieo/world.html, last accessed: 11/2004. Flynn, T. M., 1997, Cryogenic Engineering, Marcel Dekker, New York. Fuel Cell Store, 2003, “Hydrogen Storage,” Boulder, CO. http://www.fuelcellstore.com/information/ hydrogen_storage.html, last accessed: 11/2004. Gater, R., 2001, Engineering Thermodynamics: Advanced Topics, Course notes, University of Florida, Gainesville, FL.

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106 Goswami, D. Y., 1995, “Solar Thermal Power: Status of Technologi es and Opportunities for Research,” Heat and Mass Transfer 95, Proceedings of the 2nd ASME-ISHMT Heat and Mass Transfer Conference, Tata-McGraw Hill Publishers, New Delhi, India, pp. 57 – 60. Gravesen, J., Henriksen, C., 2001, “The Geometry of the Scroll Compressor,” SIAM Review, Vol. 43, No. 1, pp. 113-126. Gross, R., Otto, W., Patzelt, A., Wanner, M., 1994, “Liquid Hydrogen for Europe – the Linde Plant at Ingolstadt,” Reports on Science and Technology, Linde, Vol. 54, pp. 37 – 43. Hands, B. A., ed., 1986, Cryogenic Engineering, Academic Press, London. Hans-Joachim, H., Radermacher, R., 2003, CO2 Compressor-Expander Analysis: Final Report, Air-Conditioning and Refrigeration Technology Institute, Arlington, VA, ARTI-21CR/611-10060-01. Hasan, A. A., Goswami, D. Y., 2003, “ Ex ergy Analysis of a Combined Power and Refrigeration Thermodynamic Cycle Dr iven by a Solar Heat Source,” ASME Journal of Solar Energy Engineering, Vol. 125, No. 1, pp. 55 – 60. Holman, J. P., Experimental Methods for Engineers, McGraw-Hill, New York. McMurry, J., Fay, R., 1998, Chemistry, Prentice Hall Inc., Upper Saddle River, NJ. Mirabal, S. T., 2003, “An Economic Anal ysis of Hydrogen Production Technologies Using Renewable Energy Resources,” Mast er’s thesis, University of Florida National Hydrogen Association, 2004, “Hydrogen FAQs,” Washington, DC. http://www.hydrogenus.com/h2-FAQ. asp, last accessed: 11/2004. National Institute of Standards and Technology, 2003, Program: Thermophysical Properties of Fluid Systems, http://w ebbook.nist.gov/chemistry/fluid/, last accessed: 11/2004. Ramsay, W. C., 2003, Public-Private Dialogue, International Partne rship for a Hydrogen Economy, International Energy Agency, Paris. Schein, C., Radermacher, R., 2001, “Scroll Compressor Simulation Model,” ASME Journal of Engineering for Gas Turbines and Power, Vol. 123, pp. 217 – 225. Stuart Energy Systems, 2004, “Stuart Energy Station Product Brochure,” Ontario, Canada. http://www.stuartenergy.com/ma in_our_products.html, last accessed: 11/2004.

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107 Sunatech Inc., 2001, IEA Agreement on the Production and Utilization of Hydrogen, Task 12: Metal Hydrides and Carbon for Hydrogen Storage, Executive Summary, edited by Sandrock, U.S. Department of Energy, Washington, DC. Tamm, G. O., 2003, “Experimental Investig ation of an Ammonia-Based Combined Power and Cooling Cycle,” Ph.D. disse rtation, University of Florida Tamm, G., Goswami, D. Y., Lu, S., Hasan, A., “A Novel Combined Power and Cooling Thermodynamic Cycle for Low Temperature Heat Sources – Part I: Theoretical Investigation,” ASME Journal of Solar Energy Engineering, Vol. 125, No. 2, np. Turns, S. R., 2000, An Introduction to Combustion: Concepts and Applications, McGraw-Hill, New York. Wells, D. N., 2000, “Scroll Expansion Machin es for Solar Power and Cooling Systems,” Proceedings of Solar 2000, American Society of Mechan ical Engineers, Madison, WI, np. Wendt, H., 1990, Electrochemical Hydrogen Technol ogies: Electrochemical Production and Combustion of Hydrogen, Elsevier Science Publishing, New York. U.S. Department of Energy – Energy Efficiency and Renewable Energy, 2003, “Hydrogen Production,” Washington, DC. http://www.eere.energy.gov/RE/hydrogen_produ ction.html, last accessed: 11/2004.

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108 BIOGRAPHICAL SKETCH Robert Joseph Reed was born in Baltimore, Maryland, in 1980. He graduated summa cum laude from the University of Florida in 2003 with a B.S. degree in mechanical engineering. He will graduate from the University of Florida in May 2005 with an M.S. degree. Robert is currently involved in the design and construction of a 5kW ammonia-based combined power and coo ling cycle at the University of Florida Energy Research Park.


Permanent Link: http://ufdc.ufl.edu/UFE0009201/00001

Material Information

Title: Study of the Feasibility and Energy Savings of Producing and Pre-Cooling Hydrogen with a 5-KW Ammonia Based Combined Power/Cooling Cycle
Physical Description: Mixed Material
Copyright Date: 2008

Record Information

Source Institution: University of Florida
Holding Location: University of Florida
Rights Management: All rights reserved by the source institution and holding location.
System ID: UFE0009201:00001

Permanent Link: http://ufdc.ufl.edu/UFE0009201/00001

Material Information

Title: Study of the Feasibility and Energy Savings of Producing and Pre-Cooling Hydrogen with a 5-KW Ammonia Based Combined Power/Cooling Cycle
Physical Description: Mixed Material
Copyright Date: 2008

Record Information

Source Institution: University of Florida
Holding Location: University of Florida
Rights Management: All rights reserved by the source institution and holding location.
System ID: UFE0009201:00001


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STUDY OF THE FEASIBILITY AND ENERGY SAVINGS OF PRODUCING AND
PRE-COOLING HYDROGEN WITH A 5-KW AMMONIA BASED COMBINED
POWER/COOLING CYCLE















By

ROBERT JOSEPH REED


A THESIS PRESENTED TO THE GRADUATE SCHOOL
OF THE UNIVERSITY OF FLORIDA IN PARTIAL FULFILLMENT
OF THE REQUIREMENTS FOR THE DEGREE OF
MASTER OF SCIENCE

UNIVERSITY OF FLORIDA


2004

































Copyright 2004

by

ROBERT JOSEPH REED














ACKNOWLEDGMENTS

First and foremost, I would like to thank my wife for her constant love and support

during the pursuit of my degree. Her patience and understanding while I completed this

thesis will be forever appreciated. I would also like to thank my fellow graduate students

for making our office an enjoyable work environment and a place I looked forward going

to everyday.

I would like to thank my advisor, Dr. Herbert (Skip) Ingley, for his guidance during

my research efforts. Always willing to help, he provided much needed advice and

knowledge; but he also allowed me to develop my own ideas and solutions, providing a

wonderful learning experience. I thank my committee members Sherif A. Sherif, D.

Yogi Goswami, and Herbert (Skip) Ingley for all of their support.

Finally, I would like to thank my parents for believing in me from the beginning

and always encouraging me that I could accomplish anything I put my mind to.















TABLE OF CONTENTS

page

A C K N O W L E D G M E N T S ......... ...................................................................................... iii

LIST OF TABLES ........ .................................... .................. .............. vii

LIST OF FIGU RE S ........................................ ............. .............. .. viii

N O M EN C L A TU R E ........................................................................ ........................... x

A B S T R A C T .........x.................................... ....................... ................. xv

CHAPTER

1 M O T IV A T IO N ........................................................................................................ 1

Current E energy Trends ................................................. ... ......... .... ............... 1
Hydrogen as a Future Energy Carrier .... ....................... ...............4

2 BACKGROUND AND THEORY ........................................... ........................... 6

Hydrogen as an Energy Carrier ............................................................................. 6
Characteristics ................................... ....... ............ .. .. ..6
Production Technologies ......................................................... ..................... 7
Storage Technologies ............................ ....... .................... .. ..
Electrolysis of W after ............. .......... ....................................... ................13
Process D description ..................................... ........ .... .... .. .......... .. 14
Energy and Efficiency ............................................................ ............... 15
Electrolyzer Designs .......... .......... ................ .................... ...... 18
H hydrogen L iquefaction ...................................................................... ...................20
P process D description .............................. ........................ .. ........ .... ............2 1
Isenthalpic vs. isentropic expansion....................... ........... ............ 21
O rtho/para conversion ....................................................... ............... 24
Claude cycle ........................................................ ... ... .. ......... 25
Ammonia-Water Combined Power/Cooling Cycle.................... ..................27
P process D description .............................. ........................ .. ........ .... ............2 8
E xpander D esign ............. ....................................................... ............. 29
Positive-displacement expanders ............ .........................................30
T urbo-m achinery .......................................................... .. .. ........... ... 30
Scroll com pressor/expander ............... ................................. ............... 31
5 kW P prototype ................................................................................. 33











3 ANALYSIS METHODOLOGIES................... ........ ........................... 35

H ydrogen Energy R equirem ents........................................... .......................... 35
Electrolysis of W ater ......................................... .. ................ .............. 35
Hydrogen Liquefaction....................... .. ...... .......................... 37
Ammonia-Water Combined Power/Cooling Cycle............................... ...............41

4 EXPERIMENTAL SETUP AND DESIGN...................................... ............... 45

Scroll M machines as Expanders .................................. ............. ................. 45
Testing Apparatus and Instrumentation....... ................. ...............46
Experim ental M ethodology ............................................... ............................. 50
Procedure ............... ......... .......................50
D ata A n a ly sis ................................................................................................. 5 1

5 RESULTS AND DISCUSSION ........... .. .................................... 53

Hydrogen Production and Liquefaction..................... .... ......................... 54
Electrolysis of W after .......... ........................ ......... ... ....................... 54
Hydrogen Liquefaction.................... ......... ............................. 54
Ammonia-water Combined Cycle.....................................................64
Scroll Expander Performance Study................................................. ................. 70

6 R E C O M M EN D A TIO N S ....................................................................... ..................76

A analytical Study .................................................... ................. 76
Scroll Expander Performance Test ....................................... ............... 76

7 C O N C L U SIO N S ....... .......................................................................... ....... ...... .. 78

APPENDIX

A COMPUTER PROGRAM FOR CYCLE SIMULATIONS .......................................80

C lau de C y cle Sim ulation ...................... .. .. ............. .............................................80
Thermodynamic Property Evaluation....................................... 80
Program D escription............................ ....... .......................... 81
Ammonia-Based Combined Power/Cooling Cycle Simulation .................................86
Thermodynamic Property Evaluation....................................... ............... 87
P rog ram D description ......................................... .............................................87

B CYCLE SIM ULATION OUTPUT ........................................ .......................... 99

Claude Cycle Sim ulation R results ........................................ .......................... 99
Combined Cycle Simulation Results ................... .................... ...............100




v









C EXPERIMENTAL COMPONENT LIST ...........................................................103

L IST O F R E FE R E N C E S ......................................................................... ................... 105

BIOGRAPHICAL SKETCH ............................................................. ..................108















LIST OF TABLES
Table page

2.1 Heating values of hydrogen and other common fuels at STP ..................................

2.2 Projected hydrogen costs of various production methods .......................................8

2.3 M ass and energy density of select fuels................................... ......................... 9

2.4 Advantages and disadvantages of monopolar and bipolar electrolyzers ..................19

5.1 Specific energy requirements of the IMET electrolyzer .........................................54

5.2 Claude cycle simulation results for expander isentropic efficiency variation ............57

5.3 Claude cycle simulation results for compressor isentropic efficiency variation ........59

5.4 Claude cycle simulation results for compressor inlet pressure variation....................60

5.5 Claude cycle simulation results for compressor inlet temperature variation.............62

5.6 Claude cycle performance parameters for normal and optimum configuration .........64

A. 1 Critical properties and coefficients contained within the "gas.dat" file ..................86















LIST OF FIGURES
Figure page

1.1 World energy consumption since 1970 with projections to 2025............................. 1

1.2 US energy consumption by sector in 2002 ....................................... ............... 2

1.3 Foreign oil imported as a percentage of the total oil consumed in the U.S ................ 3

2.1 Hydrogen production technologies by energy source................................................8

2.2 Fuel and total weight of several hydrogen storage systems................... ............10

2.3 Process diagram of a simple alkaline electrolyzer.................... .................15

2.5 T-S diagram comparing isenthalpic and isentropic expansion .................................23

2.6 Claude cycle with liquid nitrogen pre-cooling and ortho/para catalyzation...............26

2.7 Combined cycle flow diagram ....................................................... ............... 28

2.8 Flow path of a single fluid pocket through a scroll compressor.............................32

3.1 T-S diagram of ideal liquefaction process ...................................... ............... 38

4.1 Sanden TRS-90 automotive scroll compressor and test stand................ ........ 47

4.2 Piston compressor with integrated tank and regulator..............................................48

4.3 Therm ocouple locations and flow m eter ............. ................................. .............. 48

4.4 Pony brake and back pressure gauge and valve............................... ............... 49

4.5 View of expander pulley showing the brake pads used as frictional surfaces............50

5.1 Sample output showing the optimum expander mass flow ratio, xe.....................56

5.2 Specific liquid yield and expander mass flow ratio as functions of the expander
efficiency ............ ...... .. ................ .......... .......... ........... 57

5.3 Required liquid nitrogen vs. expander efficiency ....... ... ........................................ 58

5.4 Specific work vs. expander mass flow ratio for varied 1 ......................................58









5.5 Specific work vs. expander mass flow ratio for varied r, .......................................59

5.6 Impact of compressor and expander efficiencies on Claude cycle FOM ...................60

5.7 Effect of compressor inlet pressure on the specific work...................... ...............61

5.8 Liquid nitrogen requirement vs. compressor inlet temperature..............................62

5.9 Specific work requirement vs. compressor inlet temperature...............................63

5.10 Comparison of inlet pressure and temperature affect on the cycle FOM .................64

5.11 Mass flow rate dependence on expander efficiency ...........................................65

5.12 Pump work variation with expander efficiency .......................................................66

5.13 Boiler heat input and absorber heat rejection vs. expander efficiency ...................67

5.14 Cycle cooling capacity as a function of expander efficiency .................................67

5.15 Cycle thermal efficiency vs expander efficiency .............. ...................................68

5.16 Effect of trace amounts of water within in the expander inlet stream on cycle
co olin g cap city ............................. .................................................. ............... 6 9

5.17 Expander exhaust and dew point temperature at several water concentrations........69

5.18 Repeatability analysis applied to shaft power output at 65 psig..............................70

5.19 Shaft power vs. rotational speed at 60, 70, and 80 psig inlet pressure ...................71

5.20 Scroll expander isentropic efficiency......................................................... ......... 72

5.21 Volumetric efficiency variation with expander rotational speed.............................73

5.22 Expander exit temperature and rotational speed relationship............................. 73

5.23 Comparison of optimum geometries of a scroll compressor (left) and expander
(rig h t) ..............................................................................................7 5














NOMENCLATURE


A ampere [A]

AC alternating current

CHWS chilled water source

CHWR chilled water return

CWS cooling water source

CWR cooling water return

CO2 carbon dioxide

COP coefficient of performance

DC direct current

E voltage [V] or energy transfer rate[Btu/hr or kW]

F Faraday's constant

FOM figure of merit

G Gibbs energy [Btu/lbm]

GFR Gibbs free energy of reaction [Btu/lbm]

H enthalpy [Btu/lbm]

HHV higher heating value [Btu/lbm]

fHHWS heating hot water source

HHWR heating hot water return

HX heat exchanger

IC internal combustion









I.D. Inner diameter [in.]

KOH potassium hydroxide

L liquid

LH2 liquid hydrogen

LHV lower heating value [Btu/lbm]

LN2 liquid nitrogen

P pressure [psia]

PV photovoltaic

Q heat transfer rate [Btu/hr or kW]

R mass specific gas constant [Btu/lbm-R]

S entropy [Btu/lbm-R]

SMR steam/methane reformation

STP standard temperature and pressure

T temperature [R or F]

V volts [V] or volumetric flow rate [cfm]

V volumetric flow rate [ft3/min or cfm]

W work transfer rate [kW]

X ammonia mass fraction

Cp isobaric heat capacity [Btu/R]

d displacement [cm3/rev]

e- electron

g vapor

h enthalpy [Btu/lbm] or hour [hr]










m

n

v

w

x

y

z

Greek

P

;




I^JT

[ls

P




Subscripts

C

CW

Elec

FW

H2

N2

NH3


mass flow rate [Ibm/hr]

number of electrons

specific volume [ft3/lbm]

specific work [kW/lbm]

mass flow ratio

liquid yield ratio

nitrogen requirement ratio




coefficient of thermal expansion

heat exchanger effectiveness

efficiency

Joule-Thompson expansion coefficient

isentropic expansion coefficient

density [Ibm/ft3]

rotational speed [rad/s]




ortho/para conversion process

cooling water

electrolyzer

feed water

hydrogen

nitrogen

ammonia vapor









P isobaric or pump

T isothermal

ab absorber

act actual

ad adiabatic

c compressor

cool cooling load

e expander

f liquid

g electric generator

h isenthalpic

in expander gas inlet

max maximum

min minimum

o standard conditions

opt optimum

out expander gas outlet

rect rectifier

s isentropic

shaft expander pulley shaft

strong high ammonia concentration stream

th thermoneutral

v volumetric









vg vapor generator

weak low ammonia concentration stream

wf working fluid














Abstract of Thesis Presented to the Graduate School
of the University of Florida in Partial Fulfillment of the
Requirements for the Degree of Master of Science

STUDY OF THE FEASIBILITY AND ENERGY SAVINGS OF PRODUCING AND
PRE-COOLING HYDROGEN WITH A 5-KW AMMONIA BASED COMBINED
POWER/COOLING CYCLE

By

Robert Joseph Reed

May 2005

Chair: H. A. (Skip) Ingley
Major Department: Mechanical and Aerospace Engineering

This thesis presents the results of a study on hydrogen production and liquefaction

and the feasibility of the 5-kW ammonia based combined power/cooling cycle to energize

these processes. Analytical models of the electrolysis, Claude liquefaction, and

combined cycle processes are developed to study the effects of variable boundary

conditions and component efficiencies on the hydrogen production rate and to determine

the optimum operating conditions. Additionally, a performance study is implemented to

gauge the applicability of a scroll expander with the 5-kW combined cycle. This research

is motivated by the current energy crisis and recent research efforts in the development of

renewable energy-based hydrogen production methods.

Analytical models are adapted to computer simulations that calculate the

thermodynamic properties, heat and work interactions, and efficiencies of each system

for variable boundary conditions and component efficiencies. Data from these

simulations are used to deduce the optimum configuration that results in the maximum









hydrogen production rate. The scroll expander performance test was carried out with a

common automotive air-conditioning scroll compressor arranged in an open-cycle

configuration using air at variable inlet pressures. Predictions on its performance with

ammonia were made based on the observed trends and by contrasting the properties of

the two working fluids.

The minimum specific energy required for electrolysis and liquefaction is 24.839

kW-h/lbm-H2 (54.76 kW-h/kg-H2) and 3.817 kW-h/lbm-H2 (8.41 kW-h/kg-H2),

respectively, for a total of 28.656 kW-h/lbm-H2 (63.18 kW-h/kg-H2). With a 5-kW

output from the combined cycle, the maximum liquid hydrogen production rate is 7.21

gallons (27.3 liters) per day. Experimental measurements of the scroll expander's

performance show isentropic efficiencies of 15 to 20 percent with maximum power

output of 0.368 Hp (0.274 kW) at 1460 RPM with an inlet pressure of 80 psig (653 kPa).

Simulation results show pre-cooling the hydrogen prior to liquefaction does not

reduce the specific energy consumption and, in fact, is detrimental to the thermal

efficiency. Furthermore, pressurized electrolysis is found to be the most effective means

of reducing the specific energy of liquefaction. The heat and work interactions of the

combined cycle scale with the inverse of the expander efficiency. Additionally, isentropic

expander efficiencies above 60% are required to extract any cooling from the cycle. The

performance test proved that scroll tip leakage is the major cause of poor expander

performance. Improvements of the scroll design such as increasing the scroll wrap and

introducing low-friction materials would significantly increase its efficiency and make it

a suitable design for low-output applications.















CHAPTER 1
MOTIVATION

Current energy consumption and forecasted demand with regard to limited fossil

fuel reserves is presented in this chapter to demonstrate the necessity for the conversion

to a renewable resources-based global energy market. Economical, environmental, and

political factors are addressed as further motivation. The remainder of the discussion

introduces hydrogen as a potential energy carrier for a renewable energy market.

Current Energy Trends

Approximately 85.7% of the world's energy is currently supplied by fossil fuels,

with crude oil making up 38.8% of that total. Global energy consumption is projected to

increase 54% over the next 25 years (Energy Information Administration, 2004).


800
70 Historical Proiected
a 700 622.9
0 567.8
600 517.3
E 470.8
0 500 403
U) 403.c
~ 400 3484368.7
= 300 2067242.8
0 200 1
100



Year

Figure 1.1. World energy consumption since 1970 with projections to 2025 (Energy
Information Administration, 2004)










This increased demand is being fed primarily from countries with rapidly industrializing

and emerging economies such as India and China. Proven oil reserves are sufficient to

satisfy this demand over the next 20 years, after which there is debate as to whether oil

production will peak before 2030 or that continued technological progress and new oil

discoveries will satisfy the demand well into this century (Ramsay, 2003).

The economic effects of increasing energy demand on a limited supply are apparent

today with peak 2004 oil prices near $50/barrel and average gas prices in the US near

$2.00/gallon. As fossil fuel production peaks and inevitably begins to decline, and

without other viable energy sources, prices will continue to escalate.





S. Residential
Transportation Residential
27% 21%




ornmercial
18%
Industrial
34%



Figure 1.2. US energy consumption by sector in 2002 (Energy Information
Administration, 2003)

Figure 1.2 gives an overview of how energy is consumed in the US economy. Industry is

affected directly and indirectly by the cost of energy. The direct effect is to increase the

cost of processing raw materials and production. Fuel costs involved with transporting

finished goods is the indirect effect. The natural response of industry to increasing cost is

to slow production and/or reduce labor forces, thus slowing the entire economy.










A number of adverse environmental phenomena such as the greenhouse effect, air

pollution, acid rain, and oil spills are attributed to the use of fossil fuels. The burning of

all fossil fuels produces carbon dioxide, a greenhouse gas. The Energy Information

Administration reports that carbon dioxide contributes over 84% to the total of

greenhouse gases emitted (Mirabal, 2003). Global warming is widely debated as an on-

going occurance, but if it were found to be so, carbon dioxide emissions would be the

main cause. Another by-product of fossil fuel combustion in air is the formation of

nitrogen oxides (NOx) that contribute to ozone depletion as well as smog formation.

Complex fossil fuels, such as petroleum and coal may also contain sulfur, which form

sulfides that can cause acid rain. These environmental factors and others mentioned

contaminate water supplies, damage ecosystems, and are related to the occurrence of

many respiratory illnesses in humans.

In 1985, the US imported 27.3% of the oil it consumed. Over the past 18 years, as

shown in Figure 1.3, the U.S. dependence on foreign oil has steadily increased to 56.1%

and is projected to be 69.6% of that consumed by 2025 (Energy Information

Administration, 2003).


80
70
60
50
40
30
20
10
0
1


965


1975 1985 1995 I 2005 2015 2025
Year


Figure 1.3. Foreign oil imported as a percentage of the total oil consumed in the U.S.


Historical ---- 0 Proiected









With greater dependence on foreign oil, the U.S. will be reliant on a stable Middle East,

Russia, and South America. International crises such as those recently in Iraq and

Venezuela will have a more significant impact on oil prices as they do today.

It is important that alternative energy sources are developed today to deal with the

issues of tomorrow. Current research initiatives around the world are focused on

hydrogen as the fuel of the future. With the development of a hydrogen economy based

on renewable resources, greenhouse gas emissions will be reduced, the economy will be

more independent of oil prices, and foreign policy will be less influenced by oil reserves.

Hydrogen as a Future Energy Carrier

In 2001, 20.4% of global energy consumption supported transportation; of which

96% was supplied by crude oil (Energy Information Administration, 2003). By

developing an alternative fuel for transportation, world oil consumption could be reduced

by as much as 19.6%. Reducing oil consumption likewise reduces greenhouse emissions

and ozone depletion. Hydrogen holds promise as the fuel to achieve these goals because

it can be produced from water using renewable energy sources and it bums clean; with

water and heat as the only combustion products (NOx emissions are possible when

burned in air).

One of the barriers to the widespread use of renewable resources is the

geographical limitation. For example, hydropower can only be utilized in areas where

damns can be built and solar power is dependent on incident sunlight, which varies from

region to region. Renewable energy technologies can be utilized more efficiently and on

a broader scale by constructing large capacity plants in regions with prominent sources of

energy. The energy can subsequently be converted to chemical energy by producing

hydrogen, enabling delivery to a larger market.









Governments around the world realize the potential of hydrogen as an alternative

fuel. Many countries have adopted research initiatives in the production, storage, and

utilization of hydrogen. The U.S. Department of Energy has recently announced plans to

advance toward a hydrogen-based energy system making fuel-cell-powered vehicles

available by 2010. Industry is following suit as most major automobile manufactures

have significant programs in place to develop fuel cell powered vehicles (Ramsay, 2003).

Hydrogen is a safe and clean fuel that when produced using renewable energy is

virtually pollution free. Hydrogen also provides a means to convert from a fixed source

of energy to one compatible with the needs of transportation. With further development

of production and storage technology, hydrogen can become the primary source of fuel

for the transportation sector and can help usher in the renewable energy era.














CHAPTER 2
BACKGROUND AND THEORY

This chapter introduces hydrogen as a potential fuel and presents a brief overview

of hydrogen storage and production systems. An emphasis is placed on the transportation

sector and renewable technologies to develop the importance of electrolysis and

liquefaction in a hydrogen economy. Following the theory of electrolysis and hydrogen

liquefaction, the ammonia-water combined cycle is introduced as a means of converting

low-temperature energy sources into usable electricity to power both systems; and

refrigeration to pre-cool hydrogen prior to liquefaction. The scroll compressor is

introduced as a potential high-efficiency expander for use with the combined cycle as

motivation for the current study.

Hydrogen as an Energy Carrier

Hydrogen is the simplest, most abundant element in the universe comprising 75%

of all visible matter by mass (Flynn, 1997). Currently, the majority of the hydrogen

produced in the U.S. is used as a chemical in a variety of commercial applications

including ammonia production, hydrogenation of fats and oils, and methanol production

(National Hydrogen Association, 2004). With the continuing depletion and increasing

cost of fossil fuels, however, greater consideration is being given to hydrogen as an

alternative fuel.

Characteristics

Hydrogen has several characteristics that make it a desirable alternative fuel for

transportation:









S Highest energy content per unit mass of any known fuel (51,574 Btu/lbm) -
hydrogen produces 2.7 times more energy per unit mass than gasoline when
burned.

Table 2.1. Heating values of hydrogen and other common fuels at STP

Fuel Higher Heating Value Lower Heating Value
Btu/lbm kJ/g Btu/lbm kJ/g
Hydrogen 60954 141.78 51574 119.96
Methane 23861 55.5 21500 50.01
Propane 21651 50.36 19772 45.99
Gasoline 20464 47.6 19003 44.2
Diesel 20249 47.1 18831 43.8
Methanol 9746 22.67 8564 19.92
(Gater, 2001)

* Clean combustion of hydrogen produces no carbon dioxide or sulfur emissions.
When burned with oxygen, the only byproducts are water and heat. If burned in
air, nitrogen oxides may be produced.

* Renewable hydrogen can be produced by a variety of methods using renewable
energy sources for a virtually limitless and pollution free fuel supply.

* Technologically compatible in the 1920s, German engineer Rudolf Erren
successfully converted IC engines to hydrogen burning engines (National
Hydrogen Association, 2004). Hydrogen can also be reacted with oxygen in a fuel
cell to produce electricity to drive a motor.

* Efficient utilization hydrogen IC engines are about 25% efficient, fuel cells are
45-60% efficient; typical gasoline IC engines are 18-20% efficient (National
Hydrogen Association, 2004). Hydrogen fuel cell powered vehicles can be up to
three times more efficient than today's gasoline engines.

Production Technologies

The U.S. currently produces 9 million tons or 3.2 trillion cubic feet (90 billion

Nm3) of hydrogen per year. Of this amount, 95% is produced by steam/methane

reformation (SMR) (National Hydrogen Association, 2004). SMR operates by reacting a

natural gas feedstock with steam at high temperatures (700 925 C) to produce carbon

monoxide and hydrogen. The carbon monoxide is then consumed in a water/gas shift

reaction to create CO2 and additional hydrogen. Other hydrogen production methods are










outlined in Figure 2.1. Detailed descriptions of each fossil fuel based production

technology are given by Mirabal (2003). Renewable energy systems are outlined by the

U.S. Department of Energy (2003).


Hydrogen Production
Technologies




Renewable Energy
Fossil Fuel Based Based
Based




*Steam Methane Reformation -Electrolysis -Photochenical
(SMR) Grid -Photo Electrochegnical
*Partial Oxidation of Heavy Wind -Biological (algae)
Hydrocarbons (POX) Solar Photovoltaic -Photo Degradation
SCoal Gasification Solar Thernal
*Biomass Gasification
*Tlhermal Dissociation
Figure 2.1. Hydrogen production technologies by energy source

SMR is currently the most cost effective method of producing hydrogen;

however, because of increasing fossil fuel cost due to diminishing supplies and reduced

capital cost of renewable energy due to technological improvements, wind and ammonia-

water combined power/refrigeration cycle solar power based electrolysis are projected to

become the most cost competitive by 2020 (Mirabal, 2003).

Table 2.2. Projected hydrogen costs of various production methods1
Hydrogen Production Costs ($/Ib)
Year 2003 2010 2030 2050
Steam Methane Reformation 0.66 0.90 2.75 9.88
Partial Oxidation 0.80 0.90 1.44 2.89
Coal Gasification 1.12 1.20 1.65 2.83
Electrolysis Grid Power (fossil fuel based) 1.53 1.63 2.42 4.12
Electrolysis PV / Antenna Power 3.47 2.40 0.91 0.65
Electrolysis Wind Power 1.33 1.14 0.78 0.60
Electrolysis Ammonia Water Combined Cycle 2.50 1.37 0.89 0.63


1 Original data converted from $/GJ using the HHV of hydrogen (Mirabal, 2003)











Although there are other methods available to produce hydrogen from renewable

resources, electrolysis is the most versatile and technologically developed. Electrolyzers

do not require high temperature for operation as do thermal decomposition, dissociation,

or chemical processes nor are they dependent exclusively on sunlight. For these reasons,

electrolysis is expected to be the predominate method of hydrogen production in a future

hydrogen economy.

Storage Technologies

One of the barriers preventing the wide use of hydrogen as a fuel is its storage.

This issue centers on hydrogen's low density and correspondingly low energy density.

Table 2.3 displays these characteristics for hydrogen under several conditions as well as

for other common fuels.

Table 2.3. Mass and energy density of select fuels


Density Energy density
Fuel ______
Ib/ft kg/m Btu/ft MJ/mn
Hydrogen
gas (STP) 0.005309 0.085044 323.60 12.06
gas (3,000 psig, 60 F) 0.9631 15.428 58,705 2,187
gas (10,000 psig, 60 F) 2.484 39.797 151,434 5,643
liquid 4.4197 70.798 269,398 10,038
Methane
gas (STP) 0.042358 0.6785 1010.70 37.66
gas (3,000 psig, 60 F) 10.778 172.650 257,174 9,583
liquid 26.367 422.367 629,143 23,442
Propane
gas (STP) 0.1183 1.895 2561.75 95.45
liquid 36.298 581.450 785,888 29,283
Gasoline (liquid) 45.884 735.010 938,976 34,987
Diesel (liquid) 53.064 850.012 1,074,483 40,036
Methanol (liquid) 49.380 791.012 481,260 17,932


(National Institute of Standards and Technology 2003, Chevron 1998)

Because of its low density, hydrogen requires a large volume for an equivalent amount of

stored energy as compared to other common fuels. To illustrate this fact, the energy










equivalent of 10 gallons (37.85 liters) of gasoline would require a tank size of 175 gallons

(662.4 liters) for gaseous hydrogen at 3000 psig and 37.6 gallons (142.3 liters) for liquid

hydrogen. Another issue with hydrogen storage in regards to its use as a motor fuel is the

combined weight of the container, safety equipment and any required insulation.

Container weights (including fuel) for several hydrogen storage methods are given for an

energy equivalent of 7.93 gallons (30 liters) of gasoline in Figure 2.2.



Hydrogen: Liquid 18 Ib fuel, 161 Ib total

Hydrogen: Gas at 5000 18 Ib fuel, 450 Ib
psig 8 18 lb fuel, 630 lb
Hydrogen: Gas at 3000
pi B 18 Ib fuel, 630 Ib total

Hydrogen: Metal hydride*
18 Ib fuel, 1700 Ib total
Methanol 108 Ib fuel, 125 Ib total

Gasoline 49 Ib fuel, 60 Ib total

0 500 1000 1500 2000
Fuel U Container Total Weight [Ib]

Figure 2.2. Fuel and total weight of several hydrogen storage systems. *Storage capacity by
weight approximately 1.1%.

There are several methods of hydrogen storage currently available or being

researched. They are summarized as follows:

Metal hydrides. Metal hydrides are specific alloys consisting primarily of

granular magnesium, nickel, iron, and/or titanium. These alloys are capable of adsorbing

hydrogen (1% 8% by weight) at high pressure and moderate temperature and releasing

it under low pressure and elevated temperature. Metal hydrides are characterized by de-

adsorption temperature. Low-temperature (< 200 F) hydrides operate at higher pressures

to prevent hydrogen release at ambient temperatures. These hydrides typically adsorb 1 -









2 percent of their weight in hydrogen. Higher temperature (> 250 F) hydrides hold 5 -

10 percent hydrogen by weight, but require significant amounts of heat to attain the

temperatures required to release the stored hydrogen. (Sunatech Inc., 2001).

Metal hydrides provide the safest means of storing hydrogen. Because the

hydrogen is stored in a solid-state media, it cannot be ignited until released. In addition,

the hydrogen is released at low pressures and moderate temperatures; therefore, no

specialized storage tank is required to deal with high pressures or cryogenic temperatures.

Despite these advantages, metal hydrides are undesirable for use in transportation.

Large, heavy, and costly storage units are required to hold equivalent amounts of energy

as current gasoline tanks, as shown in Figure 2.2. Common hydrogen impurities such as

oxygen and water reduce the ability of the tank to store hydrogen as they bond

permanently to the metal. Additionally, vibrations due to typical driving conditions can

result in particle attrition that also reduces the tank's useful life.

Compressed hydrogen. Compressed hydrogen is the simplest and one of the most

common methods of hydrogen storage and transportation. Even at 10,000 psig, however,

compressed hydrogen contains nearly 8 times less energy per unit volume than gasoline

(not including the energy expended in compressing the hydrogen). Cylinders tend to be

heavy because of the robust construction necessary to withstand the high pressures and

impacts. These factors make compressed hydrogen storage suitable for only short ranged

applications or as a reserve fuel for liquid hydrogen powered vehicles.

Liquefied hydrogen. Liquid hydrogen is formed by cooling hydrogen gas to -423

F (-253 C) at atmospheric pressure. Storage of such low temperature fluids is achieved

using a dual-walled cylinder with an evacuated space between the cylinder walls









(Dewer's flask). Due to the relatively high surface to volume ratio typical of the small

tanks used in transportation applications, additional multi-layered radiation insulation

sheets are also employed (Flynn, 1997).

There are several technological challenges that must be overcome in order for

liquefied hydrogen storage to come into widespread use. First is safe tank design to

reduce weight and hydrogen boil off due to heat infiltration. The imperfect insulation of

the inner tank supports, among other factors, causes a typical boil off rate of 3% per day

(Clean Energy Research Center, 2003). Furthermore, improved methods of hydrogen

liquefaction must be developed to reduce LH2 cost. Today, about 30% of the energy

contained in LH2 is consumed by the liquefaction process (Fuel Cell Store, 2003).

Lastly, re-filling stations must be developed such that the public can operate them safely.

Liquefied hydrogen (LH2) is currently the optimum hydrogen storage method for

vehicles in terms of tank size/weight and energy density. LH2 has the highest volumetric

energy capacity of any commercially available storage system being only four times less

than gasoline; and because hydrogen bums more efficiently than gasoline, LH2 tanks are

not necessarily four times the size of typical gasoline tanks for a given vehicle range.

This allows automobile manufactures to continue using current vehicle designs, easing

the transition into a hydrogen economy.

Carbon nanotubes and glass microspheres. Carbon nanotubes store hydrogen in

microscopic surface pores and within the tube structures via adsorption. The mechanism

by which they store and release hydrogen is similar to metal hydrides, however carbon

nanotubes are lighter, cheaper, and are capable of storing 4.2 to 65% hydrogen by weight

(Fuel Cell Store, 2003). Carbon nanotubes are still under research and development and









currently store between one and ten percent hydrogen by mass (Clean Energy Research

Center, 2003).

Glass microspheres are currently being researched as a potential hydrogen storage

method. Hydrogen is stored by first warming the tiny glass to increase their surface

permeability and then immersing them in high-pressure hydrogen gas. The spheres are

then cooled, locking the hydrogen inside of the glass balls. Increasing the temperature of

the spheres reverses this process. Experiments to increase hydrogen release rates by

crushing the spheres are also being performed. The key advantage of glass microspheres

is storage at ambient temperature.

The technology exists today for the introduction of hydrogen-powered vehicles;

however, the size, weight, and/or cost limitations imposed on storage systems by the low

energy density of hydrogen must first be overcome. Liquid hydrogen holds the greatest

promise for hydrogen-powered vehicles. These storage systems have the lowest weight

and volume of those commercially available, and with improved tank design and

hydrogen liquefaction methods, the relatively high costs will lessen over time.

Electrolysis of Water

English scientists William Nicholson and Sir Anthony Carlisle first discovered that

the application of an electric current to water produces hydrogen and oxygen in 1800.

The principle of electrolysis was later formulated by Michael Faraday in 1820. Since

then electrolysis has played only a minor role in worldwide hydrogen production;

recently contributing to only 4% of total global production (National Hydrogen

Association, 2004). Current electrolytic hydrogen production is limited to low-cost

electricity sources such as hydroelectric or small-scale onsite generation in which purity

is essential.









The importance of electrolysis in a future hydrogen economy is two fold: First, as

discussed previously, electrolysis powered by wind or the ammonia water combined

power/cooling cycle is projected to be the most cost efficient hydrogen production

method by 2020. Second, it provides a practical link between hydrogen and renewable

resources through electricity generation. In this manner, electrolysis can indirectly utilize

any energy source that can be used to produce electricity. Furthermore, when powered by

electricity generated from renewable sources of energy, electrolysis does not require

fossil fuels and has zero polluting emissions.

Process Description

Electrolysis is defined by McMurray and Fay as the use of an electric current to

drive a non-spontaneous chemical reaction (1998). Electrolysis of water consists of a

pair of oxidation/reduction reactions driven by a DC voltage applied across two

electrodes as described by equations 2. la 2.1 c.

Cathode: 2H20 + 2e -H2 + 20H (2.1a)
Anode: 20H 1 02 +H20 +2e (2.1b)


Overall: H20 H2 + 02 (2.1c)


Water is reduced at the cathode to form hydrogen gas and hydroxide ions (OH ). The

OH ions migrate toward the anode where they are oxidized to form oxygen, water, and

two free electrons. The free electrons are then attracted to the positively charged cathode,

thus completing the circuit. A schematic of a simple electrolyzer and the overall

electrolysis process is given in Figure 2.3.

Each electrode is isolated from the other with an ion-conducting diaphragm to keep

the product gases separate; and an electrolyte is used to make the solution conductive.








The electrolyte is chosen such that its reduction and oxidation potentials are less than that
of water. In this manner, the electrolyte is conserved because it acts only as an ion-
conducting substance.




1O2




Figure 2.3. Process diagram of a simple alkaline electrolyzer (adapted from Mirabal)
Energy and Efficiency
The voltage required for reversible or isentropic electrolysis is proportional to
Gibb's free energy of reaction as defined by Faraday's Law:
AG = -nFE (2.2)

where AG is Gibb's free energy of reaction
n is the number of electrons transferred in the reaction
F is Faraday's constant, 9.648531 x 104 Culomb ol
E is the cell voltage
A negative sign is included on the right hand side of Equation 2.2 because by

convention voltage input is considered negative (McMurray and Fay, 1998). The
spontaneity of a given reaction is determined by the sign of the Gibbs free energy of
reaction (from hereon referred to as GFR). GFR is positive for non-spontaneous
reactions and negative for spontaneous ones. For water at standard temperature and

pressure, (25 C and 1 atm), the GFR is 50,941 Btu/lbmH2 (14.93 kW-h/lbmH2) and the

corresponding reversible voltage is 1.23 V. The electrical energy required to drive the









electrolysis reaction is equal to the GFR (Casper, 1978). The enthalpy of reaction (higher

heating value) of hydrogen, however, is 61,451 Btu/lbm (18.01 kW-h/lbmH2).

Conservation of energy dictates that the remaining 10,510 Btu/lbm (48.89 kJ/mol) must

be supplied as heat. For a reversible process, this heat would be obtained from the

surroundings, and the electrolyzer would double as a refrigeration unit.

The second law of thermodynamics states that entropy always increases for any real

process. Entropy production in electrolysis increases the required cell voltage as

described by equation 2.3.

AH = AG + TAS = -nFE + nFT E(2.3)
OT )


where AS Aa nF ) fromFaraday'sLaw


The entropy produced is liberated as heat, which supplies the additional 10,510 Btu/lbm

necessary to form hydrogen. The voltage required for isothermal electrolysis (defined as

the thermoneutral voltage) is 1.47V. This result is obtained by replacing AG in Equation

2.2 by the HHV of hydrogen. In reality, the thermoneutral voltage is the lowest that can

possibly be achieved.

Real electrolyzers require greater than the thermoneutral voltage due to additional

overvoltages independent of the entropy generation. Overvoltage is defined as the

difference between the applied voltage and the reversible 1.23V and is proportional to the

amount of current passed through the cell (Casper, 1978). These overvoltages include:

ohmic resistance of the electrolyte, concentration polarization (changes in the

concentration of H+ or 02+ or water near the electrodes), voltage gradients at the

electrode/electrolyte interface due to the slowness of reaction (proportional to cell









operating temperature), and wire and component resistance (typically about 2% of total

loss) (Casper, 1978). The primary source of electrolyte resistance is the formation of

vapor bubbles on the electrodes (Wendt, 1990). Additional energy losses occur (typically

5% of total energy consumption) within each subsystem including AC to DC

rectification, cooling water system, feed water pumps, and electrolyzer pumps (if

necessary) (Casper, 1978).

The majority of electrolyzer manufacturers have taken steps to reduce these

overvoltages. Concentration polarization can be avoided by adequate mixing of the

electrolyte through circulation or by natural gas lift. One method developed to reduce

electrolyte resistance is zero gap cell geometry in which porous electrodes are pressed on

either side of the diaphragm, forcing the product gases to leave from the rear (Wendt,

1990). Another technique is to increase the cell operating temperature and pressure in

order to speed up reaction kinetics and reduce electrolyte resistance. However, this also

enhances corrosion of the electrodes and shortens operating lifetime.

The figures of merit measuring electrolyzer performance are current,

electrochemical, and thermal efficiencies. Current efficiency measures deviation from the

hydrogen yield predicted by Faraday's law at 1.47 V and 1000 A-h due to extraneous

electrode reactions (Casper, 1978). For most electrolyzers, this number approaches

100%. Electrochemical efficiency is defined as the reversible voltage divided by the

operating voltage. The maximum electrochemical efficiency under isothermal conditions

is 83.7%. Thermal (1st law) efficiency is the ratio of the isothermal voltage to the

operating voltage or the HHV of hydrogen divided by electricity input as given by

Equation 2.4.









V HHV
rElec = = --H (2.4)
yact EHlec

Using this definition of efficiency, ideal electrolysis operates at an apparent 120%

efficiency. Thermal efficiency is the most widely used figure of merit by electrolyzer

manufactures, therefore any given efficiency will be thermal efficiency. Commercial

electrolyzers currently operate at efficiencies (excluding subsystems) of up to 85%

(Stuart Energy, 2004)

Electrolyzer Designs

Electrolyzers are typically classified by their electrolyte; the most common of

which is alkaline/water (Casper, 1978). Others include solid polymer (SPE), seawater,

and solid oxide; descriptions of which are given by Casper (1978).

Alkaline/water electrolyzers typically operate with a 30% potassium hydroxide

(KOH) solution at relatively low temperatures of 158 212 F (70 100 C). There are

two varieties of alkaline/water electrolyzer: monopolar (tank-type) and bipolar (filter-

press). A summary of each type highlighting the unique advantages and drawbacks of

each is given below:

Monopolar or tank-type cells are constructed as an alternating set of anodes and

cathodes connected electrically in parallel and hung vertically from gas collectors into a

tank of electrolyte. Mixing of the electrolyte is achieved through simple gas lift. The

cathodes are normally surrounded by a diaphragm to prevent the mixing of gases. This

arrangement results in individual tanks operating at low voltages (typically 1.9 2.5 V)

and high currents (Casper, 1978).

Bi-polar or filter-press electrolyzers are characterized by the stacked design of the

cells. In this configuration, one side of an electrode serves as the cathode and the other as









the anode of an adjoining cell. Electrodes are connected in series such that a desired

operating voltage is achieved by increasing the total number of cells. The geometries of

these cells are relatively thin; therefore, a pump is required to circulate the electrolyte

through the cells. Bi-polar cells typically operate at lower current levels due to higher

operating voltage. Table 2.4 lists the pros and cons of each alkaline/water electrolyzer

design.

Table 2.4. Advantages and disadvantages of monopolar and bipolar electrolyzers
Advantages Disadvantages
Require relatively few, Unable to operate at
inexpensive parts high temperatures because of
Easily maintained heat loss from large surface
individual cells can be areas
isolated for repair with Bulky design requires
minimum plant downtime greater space per unit
No pumps required for hydrogen produced
Monopolar electrolysis circulation Tanks are difficult to
design for pressurized
electrolysis
Relatively high voltage
losses and non-uniform
current density distribution
result from long current paths
(Wendt)
Compact design Requires precise
Capable of operating at fabrication tolerances and
high pressures and additional gaskets due to
Bi-polar temperatures sealing problems
Lower ohmic resistance Maintenance is more
and energy losses difficult if one cell fails, the
entire cell must be shut-down
and dismantled


One of the largest electrolytic hydrogen production plants in North America was

built by Cominco, Ltd. in British Columbia, Canada. Before being shut down due to high

electricity costs, the plant produced 41 tons/day of hydrogen with 3,229 individual tank-









type cells operating at 2.1 V (70% efficient) (Casper, 1978). Since that time, most

manufactures have adopted the more efficient bi-polar design (Wendt, 1990).

Hydrogen Liquefaction

Liquid hydrogen was first produced by James Dewar in 1898; however, up until the

mid 1940s to mid 1950s it remained nothing more than a laboratory curiosity (Flynn,

1997). In the late 1950s, the US Air Force began producing substantial amounts of LH2

for its top secret "Bear" Program. Under contract to the Air Force, Air Products and

Chemicals, Inc., constructed three production plants code named "Baby Bear," "Mama

Bear," and "Papa Bear" to support Air Force aerospace programs. The largest of these

was Papa Bear, which produced 30 tons/day in 1959 (Flynn, 1997). Today, total annual

production of LH2 in North America is nearly 300 tons/day (Dmevich, 2003).

Demand for large-scale liquid hydrogen production was initially sparked by the

Apollo space program. Liquid hydrogen demand has increased and simultaneously

shifted since the 1960s from aerospace to research and industry. Flynn reports that

aerospace accounted for only 20% of total liquid hydrogen demand in 1990 (1997). This

trend is expected to continue with the onset of a hydrogen economy including the

advancement of fuel cell powered vehicles and the development of improved storage

systems.

Hydrogen production companies are already taking advantage of the higher energy

density of LH2 vs. gaseous hydrogen to effectively reduce distribution costs. Where a

full tube-trailer of gaseous hydrogen contains approximately 300 kg of deliverable

gaseous hydrogen, a comparably sized liquid hydrogen trailer carries 4000 kg (Drnevich,

2003). Another benefit of liquefying hydrogen is the ultra high purity that results from

the majority of trace impurities condensing out. High energy density and purity make









liquid hydrogen a well-suited fuel for hydrogen fuel cell powered vehicles in an emerging

hydrogen economy; giving equivalent performance and driving range as today's gasoline

and diesel automobiles.

Process Description

Hydrogen, like all gases, is liquefied by cooling it to its boiling point, -423 F

(-252.8 C). There are several liquefier designs; all of which are derived from the simple

Linde cycle shown in Figure 2.4 and follow the same general process.

The incoming gas is compressed isothermally from 1 to 2 on the diagram to a

relatively high pressure. Heat is rejected to a cold return stream and the cooled gas is

expanded from 3 to 4 on the diagram to atmospheric pressure and cryogenic

temperatures. The two-phase flow that results is separated in a flash tank where the

liquid yield is drawn off and collected and the remaining gas absorbs heat from the

warmer high-pressure stream before it's recycled back to the compressor. The expansion

can be accomplished using either a Joule-Thompson (expansion or throttling) valve or a

work-extracting device.

Isenthalpic vs. isentropic expansion

Joule-Thompson expansion is modeled as isenthalpic by neglecting potential and

kinetic energy changes as well as heat transfer (insulated valve). The effect that a change

in pressure has on the temperature for an isenthalpic process is described by the Joule-

Thompson coefficient given by Equation 2.5. A negative value indicates a temperature

increase with expansion; a positive value indicates a temperature decrease.

9T 9T ah
OJ T ) = (2.5)
ap h Oh p










By substituting the definition of specific heat at constant pressure, c, =\
O T)v


SvT = T and the volumetric coefficient of thermal expansion,
OP OT P

1 yav>
P = v the Joule-Thompson coefficient is given in its more useful form:



1JT = I V{T/ -1}v (2.6)
Cp

Equation 2.6 demonstrates that the sign of the Joule-Thompson coefficient depends

only on the product of Tf At a given pressure, the volumetric coefficient of thermal

expansion and the specific volume are functions of temperature only. Consequently, a

temperature can be identified at which u, = 0. This point is known as the inversion

temperature; and represents the maximum temperature at which a gas can be cooled by

isenthalpic expansion.

Most practical liquefaction systems use an expansion valve to produce low

temperatures (Barron, 1985). In the case of hydrogen, however, the maximum inversion

temperature at STP is well below ambient (-90.7 F (-68 C)). Additional energy is

required to pre-cool the hydrogen below its inversion temperature for isenthalpic

expansion to be effective.

Expansion in a work-extracting or work-producing device is commonly modeled as

adiabatic and reversible (i.e. isentropic). This process is represented by an isentropic

expansion coefficient, Equation 2.7.


u, = (2.7)
OP










By substituting the Maxwell relation applying the chain rule, and using
P a a a

the definitions defined previously, the isentropic expansion coefficient is given in the

same terms as the Joule-Thompson coefficient:


1p = I (vT) (2.8)
p

Equation 2.8 shows that the isentropic expansion coefficient is always positive

(the temperature always decreases with pressure) because the coefficient of thermal

expansion (/8) for gases is always positive (Hands, 1986). This conclusion can also be

arrived at intuitively by considering conservation of energy. If work is extracted from a

fluid adiabatically, the internal energy and hence temperature must decrease.

Thermodynamically, isentropic expansion is more desirable than isenthalpic

expansion. The T-S diagram in Figure 2.5 shows that an isentropic expansion will

always result in a lower final temperature than isenthalpic expansion.

T
Pi


h= const P


T2h --........- -
T2s -------- 2h
2,
s2s


Figure 2.5. T-S diagram comparing isenthalpic and isentropic expansion

Practically, however, expansion devices cannot tolerate an appreciable amount of liquid.

For this reason, expansion valves are necessary in all liquefaction systems (Barron,

1985). The Claude cycle discussed later seeks to combine the benefit of isentropic









expansion with the necessity of isenthalpic expansion as an efficient means of liquefying

hydrogen.

Ortho/para conversion

Another challenge to simple liquefaction systems is the unique sub-atomic structure

of hydrogen. Hydrogen exists in two different molecular forms: ortho-hydrogen and

para-hydrogen. Each form is distinguished by the relative spins of its protons. The

protons of ortho-hydrogen spin in the same direction whereas the proton spins of para-

hydrogen oppose one another. Hydrogen at STP (i.e. normal hydrogen) is composed of

74.928% ortho and 25.072% para hydrogen. At the normal boiling point of hydrogen

(-423 F (-293.4 C) at 1 atm) the equilibrium ortho/para composition is .21%/99.79%

(Flynn, 1997).

Converting from ortho to para hydrogen is an exothermic process, releasing 302.4

Btu/lbm (703.3 kJ/kg) of heat at STP (Barron, 1985). The conversion process is

relatively slow and the resident time of the hydrogen within the liquefier is short, so the

liquid hydrogen essentially retains its room temperature ortho/para composition.

Conversion gradually takes place in the storage tank resulting in boil-off losses because

the heat of conversion exceeds the latent heat of vaporization (190.5 Btu/lbm or 443

kJ/kg) (Barron, 1985). The heat liberated during the conversion process is sufficient to

evaporate nearly 70% of the original amount of hydrogen liquefied (Flynn, 1997).

Storage time is a major issue with regard to liquid hydrogen as a motor fuel so it is

important that the boil-off losses due to ortho/para conversion are minimized.

Catalysts are used to speed up the conversion reaction allowing the heat to be

absorbed by the liquefier. This alleviates boil-off in storage, but at a penalty to the









overall efficiency of the liquefier. The most efficient method of conversion is to have the

process take place simultaneously as the hydrogen is cooled. This is not possible in

practice but can be simulated by cooling the hydrogen to liquid nitrogen temperatures

(-320.4 F or -195.6 C) and passing it through an adiabatic converter then repeating this

procedure in a step-wise manner (Flynn, 1997). Common materials proven effective as

catalysts are ferric hydroxide gel, chromic oxide on alumina particles, and nickel silicate;

all of which provide nearly 100% conversion to para hydrogen within a few minutes

(Hands, 1986).

Claude cycle

The Claude cycle is the most commonly used system for large-scale hydrogen

liquefaction (Hands, 1986). The performance of the cycle is enhanced by pre-cooling the

compressed hydrogen gas to liquid nitrogen (LN2) temperatures. Adding catalysts in the

LN2 and LH2 baths provides a convenient and effective means of absorbing the heat of

conversion. Figure 2.6 shows a schematic of this variation of the Claude cycle with

labeled state points and flow paths.

Hydrogen gas typically enters the cycle at 1 atm and 80.6 'F (27 'C) (state 1). It is

compressed isothermally isothermall compression is achieved through multistage

compression with inner-cooling and after-cooling) to state 2, typically 20 to 40 atm

(Barron, 1985). The pressurized gas then exchanges heat with the return hydrogen and

nitrogen streams (state 2a) before entering the LN2 bath where it is cooled to -320.4 F

(-195.6 C) and where the first step of ortho/para conversion occurs (state 2b). At this

temperature, the equilibrium concentration of para hydrogen (assuming 100%

conversion) is 50%.
















Liquid nitrogen bath


We --

Compressor












Liquid hydrogen

bath


Catalyst bed 2


Figure 2.6. Claude cycle with liquid nitrogen pre-cooling and ortho/para catalyzation

It is desirable to perform the maximum amount of conversion at this stage because liquid

LN2 is less expensive to produce then liquid oxygen. The hydrogen is further cooled in

the first heat exchanger to state 3. At this point, a portion of the flow (typically 60 to

80%) is diverted and expanded isentropically through a work-extracting device and used

to pre-cool the compressed hydrogen. The expander work is used to offset the

compressor work requirement, increasing the overall cycle efficiency. The remaining

flow continues through the next two heat exchangers and into the liquid receiver. Here

the flow streams are halved and throttled through expansion valves. The liquid yield from

the first stream (state 9a) is collected in the receiver and used sacrificially to absorb the

heat of conversion from the second catalytic bed. The second stream (state 9b) is passed
heat of conversion from the second catalytic bed. The second stream (state 9b) is passed









through the catalytic bed where it is ideally converted to 99.789% para hydrogen and

extracted.

Ammonia-Water Combined Power/Cooling Cycle

The ammonia-water combined power/cooling cycle proposed by Goswami (1995)

utilizes a binary ammonia/water working fluid to produce both power and refrigeration.

The cycle is a combination of an ammonia-water refrigeration system and an ammonia-

based Rankine cycle.

An ammonia-water mixture is used because of its desirable thermodynamic

properties. Binary mixtures have varying boiling points depending on the concentration

of the more volatile species. This characteristic gives a good thermal match with a

sensible heat source, thereby reducing the irreversibility associated with heat transfer

(Hasan, Goswami, 2003). Additionally, the low boiling point of ammonia allows the

utilization of low temperature heat sources such as low-grade waste heat from industrial

processes, solar water heaters, and geothermal sources. In a theoretical investigation

performed by Tamm et al., the cycle is shown to operate with heat source temperatures as

low as 116.6 F (47 C) albeit with low first law efficiency (- 5%). When operating with

a heat source temperature of 224.6 F (107 C) and idealized parameters, however,

second law efficiencies greater than 65% are possible (2003).

The unique ability of this cycle to produce both power and refrigeration gives rise

to two advantages for use in a hydrogen economy. First, the cycle can utilize low-grade

renewable heat sources such as that available from inexpensive flat plate solar collectors

to produce the power needed to drive an electrolyzer and liquefier. Second, the cooling

produced by the cycle can be used to pre-cool hydrogen prior to liquefaction, thereby












reducing the power requirement of the compressor. In this manner renewable energy


source utilization is improved compared to technologies such as wind or P.V.


electrolysis.


Process Description


Figure 2.7 gives a schematic of the cycle showing state points and flow paths.





Rectifier
Column
-CWR
Superheater
CWS -HHWR

HHWS


______ HHWS HHWR
Vapor
Generator
wit a ta Recovery Heat e
Exchanger
Ex-li1 Expander W T
te wk s e Absorber




Solution
Pump Cooler

CWS CWR


CHWR CHWS



Figure 2.7. Combined cycle flow diagram


The fluid leaves the absorber at state 1 as a saturated solution at the cycle low pressure


with a relatively high ammonia concentration. It is pumped to the system high pressure


(state 2) before traveling through the recovery heat exchanger where it absorbs heat from


the weak solution returning to the absorber. The solution is then partially boiled in the


vapor generator by the heat source producing saturated ammonia vapor and relatively


weak concentration ammonia-water saturated liquid. The weak solution leaves the vapor


generator at state 4 and rejects heat to the high concentration stream before it is throttled









to the system low pressure and sprayed into the absorber. The rectifier cools the

saturated ammonia vapor to condense out any remaining water. The vapor is then

superheated to state 7 and expanded to produce work. The sub-ambient exhaust vapor

(state 8) provides refrigeration before returning to the absorber where it is re-absorbed

into the weak solution. The heat of condensation is rejected to the low-temperature source

and the cycle repeats.

The power output and cooling capacity of the cycle under given operating

parameters is highly dependent on the expander efficiency. Irreversibilties due to friction

and leakage decrease the amount of work extracted from the fluid. Because less work is

extracted, the expander exhaust temperature is higher and the cooling capacity is reduced.

Losses in the expander have the greatest impact on the overall cycle efficiency (Tamm et

al., 2003), so it is important to select an optimal design.

The main criteria for expander selection are operating pressures and temperatures,

flow rate of ammonia vapor and material compatibility with ammonia. Ammonia is a

corrosive substance that reacts with metals such as copper, brass, and bronze, all of which

are commonly used as bearing or bushing material. The expander selected for use in the

combined cycle must be sized correctly for the flow rate and for the operating pressure

ratio for maximum power production and refrigeration capacity. It must also be

constructed out of steel, aluminum, or any other material compatible in an ammonia

environment.

Expander Design

An expansion device extracts mechanical energy from a fluid by expanding it from

a high to a low pressure and converting it into shaft work. Various expander designs

using unique expansion methods exist throughout industry. These designs can be









organized into two categories, positive-displacement and turbo-machinery, based on the

method of fluid displacement.

Positive-displacement expanders

Positive-displacement machines such as reciprocating and rotary piston, rotary

vane, and screw operate by expanding a fixed volume of fluid per oscillation. Torque

pulsation is a common phenomenon due to the inherent discontinuity associated with the

finite number of pistons or lobes and fixed displacement. Reliability is an issue with

positive-displacement machines because of a greater number of moving parts (i.e. piston

linkages, sliding vanes); and in the case of pistons, a lubrication system to reduce leakage

encountered in the gap between the moving seals and volute.

Turbo-machinery

Turbo-machinery, comprised of axial and radial flow turbines, utilizes the pressure

differential across a series of radial blades to provide a "lift" force to turn the rotor,

thereby producing shaft work. In this manner, a continuous power output is provided.

Reliability is improved over positive-displacement expanders because the rotor is the

only moving part.

Turbines are designed with a clearance between the blade tips and the volute to

allow free rotation; however, leakage at the tips (windage loss) is the primary cause of

irreversibility in the expansion process. Blade tip clearances remain approximately

constant for varying turbine size. As turbine size is decreased, the loss due to windage as

a percentage of the work output becomes increasingly significant. For this reason,

positive-displacement expanders are more suited for small-scale operations.

The amount that the blade tip clearances can be reduced is limited by the

centrifugal force and/or thermal expansion of the blade material. Typical turbine









operating speeds range from a few thousand up to tens of thousands RPM. Centrifugal

force is dependent on blade tip speed, which is function of the RPM and the rotor

diameter. As a result, larger turbines suffer greater radial blade deformation and are less

suited for blade tip clearance reduction.

Scroll compressor/expander

The scroll compressor was first invented by Leon Creux in 1905 (Gravesen and

Henriksen, 2001). Commercial interest in the technology wasn't strong until the

introduction of computer numerically controlled (CNC) machines in the 1970s. CNC

machines provided the basis for machining the precise elements needed for a scroll

compressor to operate efficiently and quietly (Copeland corp., 2001).

A scroll compressor consists of two identical spiral elements assembled with a 1800

phase difference. During operation, one scroll remains stationary and the other is

attached eccentrically to a motor shaft. This configuration allows the scroll to rotate in an

orbiting motion within the fixed scroll. The phase difference between the two scrolls is

maintained using an anti-rotation device, typically an Oldham coupling (Copeland corp.,

2001).

The fluid flow path within a scroll compressor or expander is described by Figure

2.8. As the rotating scroll (green) orbits about the fixed scroll (red), the outer periphery

forms a line of contact with the fixed scroll, capturing a crescent shaped volume of gas

(step 1). The gas is forced toward the center discharge port in steps 2 thru 5 and

compressed due to the decreasing volume of the crescents. This is indicated by the

brilliance of the yellow color representing the gas pocket. Because several of these gas

pockets are being compressed simultaneously, as depicted in step 6, torque pulsation









common with other positive-displacement machines is low. Scrolls compressors have

been widely adopted by the HVAC industry because of the advantages they offer,

including: simplistic design (i.e. fewer moving parts), low friction, low torque pulsation,

and compliance.







r6. 1. 2.








5. 4. 3.

Figure 2.8. Flow path of a single fluid pocket through a scroll compressor (Adapted from
Gravesen and Henriksen, 2001)

Because of their unique geometry, scrolls do not require valves or valve actuators;

furthermore, there are no linkages or sliding vanes. The relative rolling motion of the

contact points offers less resistance than sliding friction. Additionally, the rolling

contacts provide a seal such that large volumes of oil used as a sealant are not required

and leakage is reduced (Copeland corp., 2001). Continual compression process of the

scroll results in a smoother power output and consequently less noise and vibration than

piston-type devices. Compliance mechanisms balance the dynamic pressure and

centrifugal forces in order to maintain proper sealing. These loading mechanisms correct

tolerances as the scroll surfaces wear and allow the scroll elements to separate slightly in

the axial or radial directions in response to a sudden pressure spike (axial compliance) or









the presence of small amounts of debris or liquid (radial compliance). Taken together,

these attributes contribute to the fact that scroll compressors typically have 10% higher

mechanical efficiencies than comparably sized piston compressors (Wells, 2000) and less

leakage than other compressors in its class (Schein and Radermacher, 2001).

Literature suggests the potential use of a scroll compressor as a high efficiency

expander (Wells, 2000). Copeland compressors have been used successfully as

expanders with R-134A and R-245FA refrigerants as the working fluid. Efficiencies over

70% were demonstrated when operated with pressure ratios between three and five

(Warner, Wayne Copeland Corporation, Personal Conversation, 10 May 2004). Scroll

expanders have also been utilized in an organic Rankine micro combined heat and power

system patented by Yates et al. in 2002 (US Patent and Trademark Office, 2002).

5 kW Prototype

The applicability of the ammonia-water combined cycle for small scale power

generation utilizing low temperature heat sources is currently being studied at the

University of Florida's Energy Research Park. A prototype producing 5 kW of electrical

power has been designed and is under construction.

Heat source and sink. The low-temperature heat source is simulated using a

liquid-propane-fired boiler to heat water to 180 OF. The heat sink for the cycle is cooling

water, which is continually circulated through a 500,000 btu/h cooling tower.

Temperature control is accomplished using a combination of 3-way automatic control

valves and several shell and tube heat exchangers.

Absorber and solution pump. The absorber is a falling-film type. This design

offers a combination of sufficiently high heat transfer rates and large surface areas for









absorption. The fluid leaving the absorber is saturated, therefore no net positive suction

head (NPSH) is available for the pump, leading to cavitation. For this reason, a roller-

type positive-displacement pump is used.

Vapor generator and rectifier. The vapor generator and rectifier are integrated as

a single unit such that no separator is required. The vapor generator is a shell and tube

heat exchanger with hot water on the tube side; the rectifier is a packed column. As the

ammonia bubbles out of solution, it travels through the rectifier and the remaining

effluent drips back down into the vapor generator where it is re-boiled.

Electricity production and cooling capacity. The maximum power output of the

expander is 5.6kW. This work is used to run an electric generator that produces 200 Vrms

single phase AC at 400 Hz. A frequency converter switches the frequency from 400 to

60 Hz required by the electrolyzer. The maximum equivalent cooling capacity of the

system is 1.25 kW; this is demonstrated by cooling a fixed volume of water.















CHAPTER 3
ANALYSIS METHODOLOGIES

This chapter outlines the analytical procedure developed to find the expected

energy requirements for electrolysis and hydrogen liquefaction, as well as the heat and

work interactions of the combined cycle at steady state. An analysis on impact of the

combined cycle expander efficiency on the cooling capacity and the liquid hydrogen

yield is discussed as motivation for an experimental study.

Hydrogen Energy Requirements

Electrolysis of Water

The electrolyzer model used in this study is based on the Stuart Energy

Vandenborre IMET Electrolyzer. The IMET is selected for two reasons: its relatively

simple design due to pump-less electrolyzer circulation, and its high thermal efficiency

(operating at a cell voltage of approximately 1.7V) (Stuart Energy, 2004). It utilizes an

alkaline electrolyte in a filter-press arrangement and can deliver hydrogen at pressures of

up to 363 psi (25 atm), which reduces the compressor power required for liquefaction.

The analysis determines the total electrolyzer power consumption per unit mass hydrogen

produced including the power required to operate the sub-systems of the electrolyzer,

namely the cooling water system, feed water / deionization system, and AC/DC rectifier.

Equation 3.1 defines the thermal efficiency of the electrolyzer, assuming 100%

current efficiency (Casper 1978).

V HHV
Sctu (3.1)
17attual Eplec









The losses that occur in the electrolysis process are dissipated as heat. A cooling

water system is employed to remove this heat and keep the electrolyte temperature

relatively low. At temperatures above 302 F (150 C), the corrosiveness of the alkaline

electrolyte causes significant electrode corrosion (Wendt, 1990). The cooling load is

determined using the definition of thermal efficiency and the higher heating value (HHV)

of hydrogen as shown in Equation 3.2.

Qcoo,.1 = HHVH2 (I- h, ) (3.2)

Using a typical COP value of three for many refrigeration systems, the work required to

produce the cooling water is estimated by:


cW C (3.3)
COP

The cooling water volumetric flow rate, given by Equation 3.4, is found by applying

conservation of energy and specifying a 10 F (5.56 C) temperature drop across the

electrolyzer.


eQec (3.4)
S Cw .cpc (AT)

Pump work is calculated using Equation 3.5, assuming a pressure drop of 10ft of water

and a pump efficiency of 70%.


WP = w (3.5)
rp

The feed water required for electrolysis is obtained by assuming the reaction takes

place in stoichiometric proportion. From the overall chemical reaction of Equation

(2.1c), one mole of water is required for every mole of hydrogen or 9 lbm of water for

every lbm of hydrogen. On a volumetric basis, this equates to 1.0825 gal/lbm H2. The









maximum energy required for deionization of water is assumed to be 10% of the energy

required for electrolysis as suggested in the literature (Casper, 1978).

HHV
EFW = 0.1 x -- H (3.6)
7thEle

Casper reports the typical efficiency of an AC/DC rectification system to be 95%

(1978). The total energy consumed per unit mass of hydrogen by the electrolyzer and

sub-systems is given by Equation 3.7.

HHV
Eee H +Wcw + EF (3.7)
th1ec Urect

Hydrogen Liquefaction

The Claude cycle is analyzed to determine the total liquefaction energy per unit

mass hydrogen liquefied. The inlet pressure and temperature, as well as the expander

mass flow ratio are varied independently to develop a family of performance curves used

to gauge each parameter's effect on liquid yield and the total specific liquefaction energy.

Each configuration is then evaluated based on its figure of merit (FOM).

The figure of merit (FOM) is used to measure the performance of liquefaction

systems. It is defined as the ratio of the work required by an ideal liquefier to the work of

an actual liquefier.


FOM WVdea3 (3.8)
W

Ideal liquefaction. Ideal liquefaction is described by the first two processes of a

reverse Carnot cycle: isothermal compression followed by an isentropic expansion

(Barron, 1985). Additionally, all gas that enters the cycle is liquefied. Figure 3.1 shows

the T-S diagram of the process.









T P2
Pl
2






s

Figure 3.1. T-S diagram of ideal liquefaction process

Applying the First Law to the entire cycle (neglecting changes in potential and

kinetic energy) yields:

Wie = c +m h(h -h,) (3.9)

For a reversible isothermal compression process, the heat rejected is given by the Second

Law as:

Qc = ihT,(s, S2) =MT,(s- Sf) (3.10)

Substituting this result into Equation 3.9 gives the ideal work requirement per unit mass

gas compressed.

Wne- e= (hf h2 )-T (S -s ) (3.11)


Claude cycle. The assumptions for the Claude cycle analysis are listed below:

* Heat transfer from the environment is negligible
* Heat exchangers and liquid baths are 100% effective
* Negligible pressure drop through pipe, fittings, and heat exchangers
* Negligible loss in power transmission from expander to compressor
* Tio = T1, Tioa = T2b
S T7 = T8 = Teto minimize irreversibility upon mixing (Hands, 1986)
* T3 = -350 OF
* Compressor efficiency, q, = .75
* Expander efficiency, 7e = .85









* Electrolyzer produces 100% pure normal hydrogen (74.928% ortho, 25.072% para)
* Ortho/para conversion proceeds to equilibrium within the liquid nitrogen (LN2)
bath


In this model, ortho-para conversion takes place in two isothermal stages. First, the

gas is cooled to LN2 temperatures (-320.4 F, -195.6 C) and passed over a catalyst bed.

Equilibrium concentration of para hydrogen at this temperature is 50.5%. This

corresponds to an approximate 25.43% conversion from normal hydrogen, releasing

75.28 btu/lbm (175.1 kJ/kg) of heat (heat of conversion at -320.4 F is 296.07 btu/lbmH2

(688.62 kJ/kg)). The second stage takes place in the liquid hydrogen-receiving tank at

liquid hydrogen (LH2) temperatures (-423 F, -252.8 C). The heat of conversion from

normal to para hydrogen at -423 F is 302.38 btu/lbm (703.3 kJ/kg). The heat released in

proceeding from 50.5% to 99.789% para hydrogen is 134.56 btu/lbm (312.97 kJ/kg).

The liquid yield of the cycle per unit mass hydrogen compressed is found by

applying the First Law to a control volume including the three heat exchangers, Joule-

Thompson valve, and liquid hydrogen-receiving tank (subscripts refer to Figure 2.6).

f AHC2 + mh2b + eh = ( m-m hf ho +m + hM +m fhf (3.12)

where AHC2 is the heat of conversion in the second stage
i is the inlet mass flow rate of hydrogen

Dividing by Mi, introducing the liquid yield per unit mass hydrogen compressed,

mnf m
y -', and the expander mass flow ratio, x = e, and solving for y gives:
m m

h1oa h2b + 17, e (h3 he, (3.13)
y = + (3.13)
hoa -hf + AHC2 hloa hf + HC2

The amount of liquid nitrogen required to pre-cool the compressed hydrogen and absorb









a portion of the heat of conversion is determined by applying the First Law to a control

volume encompassing the three-stream heat exchanger, liquid nitrogen receiver and the

control volume from the previous analysis.

lhAHCl + rhfAH +C2 h, +2lh +MN h = (h-lhf)hl0 + lN h + ihA +li hf

Dividing by Mi, defining the mass ratio of liquid nitrogen to compressed hydrogen as

m2
z = and solving for z yields Equation 3.14


Z c h h, hh h h h1i hf + AHc2
z=-- ---- --x +y (3.14)
h hA hA hA e hc -hA h -hA

where AHc, is the heat of conversion in the first stage

Dividing Equation 3.14 by the liquid yield, y, gives the hydrogen requirement in terms of

unit mass hydrogen liquefied. Based on the literature, the specific energy required to

produce liquid nitrogen is assumed 766.82 btu/lbm-N2 or 0.225 kW-h/lbm-N2 (Gross et

al., 1994).

An energy balance on the compressor, including work contributed from the

expander, gives the specific power required per unit mass hydrogen to drive the cycle.

Wc (h2 hi)- T(s2 ) 7 x(h (3.15)
r-- exe(h, -hg) (3.15)


Dividing this result by the liquid yield ratio gives the compressor work per unit mass

hydrogen liquefied. Total liquefaction energy is the summation of compressor work and

the liquid nitrogen power requirement.

The expander mass flow ratio, x, is varied from 0 to 0.9 with four other

independent parameters: expander and compressor isentropic efficiency, and compressor

inlet pressure and temperature in individual cases to determine their influence on the









cycle performance. In cases one and two, the expander and compressor isentropic

efficiencies are decreased from 1.0 to 0.4 in 0.2 increments to gauge their effect on the

cycle performance. Case three looks at a range of compressor inlet pressures (1 to 25

atmospheres in increments of five) at a fixed inlet temperature of 80 F (26.7 C) to

simulate the operating pressure range of the IMET electrolyzer. In case four, the

compressor inlet temperature is varied from 0 to 80 F (-17.8 to 26.7 C) in twenty-

degree increments; representing the pre-cooling effect of the combined cycle. Plots are

created displaying the temperature, pressure, and component efficiency dependence of

the key liquefaction parameters: total specific work, liquid yield, liquid nitrogen required,

and figure of merit.

The critical state points required to calculate the performance parameters given by

Equations 3.13 thru 3.15 are defined based on the inlet temperature and pressure (state 1)

as well as the zero pressure drop assumption and the isentropic efficiencies of the

compressor and expander. A computer program has been developed to assist in

calculating the state properties and performance parameters for each iteration as well as

for plotting the data. A detailed description of the program including a portion of the code

follows in Appendix A.

Ammonia-Water Combined Power/Cooling Cycle

The ammonia water combined power/cooling cycle of this study is based on the

experimental system under construction at the University of Florida's Energy Research

Park. This particular system is designed to provide 5kW of electrical power from a heat

source temperature of 180 F in order to simulate temperatures attainable from

inexpensive flat-plate solar collectors. Additionally, the maximum pressure is









constrained such that high-pressure fittings are not required, thereby reducing the capital

cost. Other assumptions and/or specifications made in the design are listed below:

* Fluid exiting the absorber and vapor generator is saturated liquid/vapor
* Absorber operating temperature is 100 F
* Vapor generator operates at 170 F
* Cycle high and low pressures are 110 psia (7.58 bar) and 40 psia (2.76 bar),
respectively
* Rectification is 100% efficient (100% pure ammonia vapor at state 7)
* Recovery heat exchanger has a 85% effectiveness, E
* Weak and strong solution streams have equal specific heats
* 75% electric generator efficiency, 7,
* 5 F approach temperature in the cooler
* Negligible pressure drop through pipes, fittings, heat exchangers, and other
components


Binary mixtures differ from pure substances in that knowledge of three

thermodynamic properties is needed to completely define a state (two under saturated

conditions). As such, by specifying the operating temperature and pressure of the

absorber, and assuming saturated conditions exist at the exit, the mass fraction of

ammonia in the strong solution stream is fixed. The mass fraction of ammonia in the

weak solution stream leaving the vapor generator at state 4 is determined in a similar

matter.

The next step in the analysis is to find the mass flow rate of ammonia vapor

through the expander. Equation 3.16 is obtained from an energy balance on the expander

including the electric generator efficiency.


m We (3.16)
17g (h7 h8)

The strong and weak solution mass flow rates follow from species and mass balances on

the vapor generator as described by Equations 3.16 and 3.17.









h rNH, (XNH, XNH3,strong)
NH3weak (X -~ XNH3.) (3.17)
VH NH, strong ~' NH3,weak)

where X is the mass fraction of ammonia

hNH3,strong = NH3,weak NH3 (3.18)

The temperatures of the cold (state 3) and hot (state 5) exit stream are found from

the definition of heat exchanger effectiveness. Because the specific heats of the two

streams are approximated as equal, the equations become a ratio of only temperatures and

mass flow rates.

SX InNH3 weak V4 -T2 (3.1
T3 = +2 (3.19)
mNH3strong

T, =T4 -(T -T) (3.21)

where E is the heat exchanger effectiveness

Heat and work interactions of the absorber, pump, and cooler are calculated from

energy balances on all inlet and outlet streams. The four equations summarizing this

process are given below:

Qab = rnNH3,weak h6 + NH3 h9 'NH3,stronghl (3.21)

W, = NH3,strong (h2 -h,) (3.22)

Cooling capacity is dependent on the temperature of the cooled fluid. It is assumed that

hydrogen at 90 OF is being cooled; therefore, T9 is 85 F (assuming a 100 approach

temperature).

Qc = NH3 (h8 h9) (3.23)

The total heat input to the cycle is determined by "black boxing" the vapor generator,

rectifier, and superheater and considering state points 3, 4, and 7.









Qvg = NH3,weakh5 + I NH h7 IlNH3,strongh3 (3.24)

Lastly, the cycle thermal efficiency is computed from the work and heat interactions as

shown in Equation 3.25. The cooling affect is accounted for by scaling it with the same

coefficient of performance used in the electrolyzer analysis.


We + COP
r7th,cycle = (3.25)


Properties at each state point are estimated using the Gibbs energy method combined

with pure fluid correlations as described by Tamm (2003).

This procedure is repeated for a fixed power output and varied expander

efficiencies. These data are plotted to study the effect on the cycle cooling capacity, heat

input, and pump work and to relate these quantities to the liquid hydrogen yield.

Additionally, the effect of trace quantities of water in the expander inlet stream on cycle

efficiency and cooling capacity is analyzed.

A MatLAB program is developed to calculate all state points of the combined

cycle, equations 3.16 thru 3.24, and the optimum liquid hydrogen yield for each value of

expander efficiency. A detailed description of the program and portions of its code are

presented in Appendix A.














CHAPTER 4
EXPERIMENTAL SETUP AND DESIGN

The potential application of a scroll compressor as a high-efficiency expander for

small-scale power generation (i.e. the 5kW combined cycle) is discussed in this chapter

as background for the experimental study. A detailed description of the compressor and

testing apparatus is given followed by an outline of the experimental methods.

Scroll Machines as Expanders

Scroll compressors have been proven as viable expansion devices. Copeland has

performed limited research on scroll expanders using their refrigeration scroll compressor

with R-134A and R-245FA as the working fluid. Results show that efficiencies of

greater than 70% are attainable (Warner, Wayne Copeland Corporation, Personal

Conversation, 10 May 2004). Other publications have investigated the use of scroll

expanders in small-scale solar driven Rankine cycles (Wells, 2000). To date, however,

no known research has been conducted with an ammonia working fluid.

Ammonia offers particular challenges to the design or selection of any expander.

One of which is corrosiveness. Ammonia is corrosive to copper and copper-containing

alloys present in the bearings and motor stators of hermetically sealed compressors like

those manufactured by Copeland. Additionally, ammonia is a small molecule and thus

has relatively low density compared to R134-A (0.0433 lbm/ft3 vs. 0.2622 lbm/ft3), so

leakage losses become more prevalent.

Small-scale, high-efficiency expanders are desired for the 5kW ammonia-water

combined power/cooling cycle because its overall performance and cooling capacity is









highly dependent on the expander efficiency as discussed in later sections. For a

designed power output, increasing the expander efficiency reduces the required mass

flow through the system and hence reduces the total energy consumption. Individual

component and pipe size is reduced as well.

At the 5kW size, the scroll design offers several advantages over turbines as

explained in the background and theory. Ammonia turbines in the 5kW range are

inherently inefficient due primarily to leakage loss at the tips. Tom Revak of Revak

Industries reports that the efficiency of a 5kW is likely to be approximately 40% whereas

Sam Ni of Scroll Labs predicts an isentropic efficiency of 67% for a comparably sized

scroll expander. Custom-design is cost prohibitive however; with the design and

fabrication cost of the aforementioned scroll expander being $280,000.

The objective of the experiment is to test an "off-the-shelf" unit with air and predict

its performance with ammonia from the data obtained. From these observations, an

indication of whether the scroll expander is feasible in the combined cycle is determined

and recommendations for design improvements are made. This experiment also lays the

foundation for further research of scroll expanders for use in the ammonia-water

combined cycle and other small-scale power generation systems.

Testing Apparatus and Instrumentation

The Sanden TRS-90 automotive scroll compressor (shown in Figure 4.1) was

selected as the test compressor for three reasons: it operates in the 5kW range, the scroll

elements and the housing is constructed of aluminum and the bearings and clutch of steel

(ammonia compatible), and it has a pulley and clutch assembly convenient for testing.

The only modification necessary to run the compressor in reverse is the removal of a

reed-type check valve located beneath the stationary scroll element within the housing.









The compressor is designed to operate at a pressure ratio of approximately six with R-

134A refrigerant. Displacement of the compressor is 85.7 cc/rev.




4








2






Figure 4.1. Sanden TRS-90 automotive scroll compressor and test stand

The expander is connected to compressed air source at the suction port (1) using 1"

I.D. plastic tubing. The discharge port (2) is 1/4" I.D. and is vented to the atmosphere.

Also shown in Figure 4.1 is the pulley and clutch assembly (4). The clutch is on/off

modulated by applying 12 volts DC at point 3. Figure 4.2 shows the 5-Hp compressor

and tank used as the compressed air source. The compressor has a maximum pressure of

125 psig and a pumping capacity of 15.7 scfm at 90 psig. A 110-psig regulator is used to

adjust the expander inlet pressure.

Temperatures measurements are taken from thermocouples inserted into the inlet

and exit flows at points 1 and 2 as shown in Figure 4.3. The signal from each

thermocouple is calibrated and conditioned to lmV/F using two thermocouple-to-analog

converters (3) and recorded from a pair of multimeters.
























Figure 4.2. Piston compressor with integrated tank and regulator


-


Figure 4.3. Thermocouple locations and flow meter
The volumetric flow rate of compressed air is measured in standard cfm (standard

conditions are 1 atm and 70 F) using an in-line acrylic gas rotameter (number 4). The









reading is adjusted to actual cfm using the ideal gas relation with the observed inlet

temperature and pressure as described by Equation 4.1.

Figure 4.4 shows the pony brake used to measure the torque output of the expander

and the back pressure gauge (1). The pony brake frame is constructed of wood with

ordinary go-cart brake pad material employed as the friction material. An enlarged view

of the pulley showing the brake material is seen in Figure 4.5. This material has the

added advantage in that it acts as an insulator, protecting the wood from the excessive

heat. The frictional force applied to the pulley is varied by adjusting a pair of wing nuts

(2). The force exerted by the expander torque is measured 14.125" from the centerline of

the expander shaft (3) using a Pelouze 5-pound scale. Rotational speed is measured in

RPM from the center of the pulley with a handheld tachometer (not shown).


















3
iiiiiiiiiiiii


Figure 4.4. Pony brake and back pressure gauge and valve

















Tachometer
placement


Brake pads


Figure 4.5. View of expander pulley showing the brake pads used as frictional surfaces

A detailed component list of the experimental apparatus including the range and

resolution of each instrument (if applicable) follows in Appendix C.

Experimental Methodology

Procedure

Startup:

1. Activate the voltage supply, multimeters, and thermocouple-to-analog
converters.
2. Close the compressor valve.
3. Start the compressor and allow it to charge to 125 psig.


Test:


1. Cap the expander exit port.
2. Crack the compressor valve and allow system to charge.
3. Select the desired source pressure by adjusting the tank regulator.
4. Once pressure is selected, close the compressor valve and open the
backpressure valve to discharge the system.
5. Close the backpressure valve and remove the expander exit port cap.









6. Loosen the wing nuts on the pony brake to ensure that testing begins with
minimum brake force.
7. Initiate the test by fully opening the compressor valve.
8. Record rotational speed (RPM), inlet and exit temperature, flow rate,
backpressure and arm force.
9. Tighten the pony brake wing nuts about 1/8 of a turn and repeat step 8 for
each trial.
10. Continue until the expander is stalled.
11. Terminate the test by closing the compressor valve.
12. Allow 15-20 minutes between each test for the compressor motor and
expander clutch assembly to cool.

Data Analysis

Experimental data is collected in an Excel spreadsheet programmed to perform the

conversions and calculations necessary to complete the analysis. Each calculation

performed in the spreadsheet and the formulas used for them are explained below.

The corrected volumetric flow rate for the given inlet pressure and temperature is

related to the indicated value by treating the air as an ideal gas (Equation 4.1).


corrected indicated o (4.1)
poT)

where po, To are at standard conditions (1 atm and 70 F)

The mass flow at standard conditions is found by multiplying the fluid density by the

corrected volumetric flow rate as described by Equation 4.2.


m = corrected (4.2)



where is substituted for the density
RT

Mass flow is corrected to the actual inlet conditions using Equation 4.3 (Holman, 2001).


Corrected T (4.3)
So









Shaft power output is defined by Equation 4.4, the product of the force measurement and

the expander rotational speed.

What = Force x z (4.4)

The volumetric efficiency quantifies the amount of tip leakage encountered during

operation. It is defined as the ratio of flow usefully expanded to the total flow through the

expander (Equation 4.5).


= (4.5)


where m is the rotational speed (RPM)
d is the expander displacement per revolution

Inlet and exit enthalpies are computed from the measured temperatures and pressures and

are used in Equation 4.6 to calculate the isentropic efficiency.

h ho-
= out (4.6)
h, hout,














CHAPTER 5
RESULTS AND DISCUSSION

The electrolyzer and its sub-systems are analyzed to find the specific energy

consumption, thermal efficiency, and cell voltage. Following the electrolyzer

investigation, simulations of the Claude cycle are made to determine the effects of

component efficiencies and compressor inlet conditions on specific energy consumption.

Results of each test are presented in tabular form with several graphs displaying the

important trends. The analysis concludes with the selection of the optimum operating

parameters.

The ammonia-water combined cycle simulation examines the dependency of the

boiler heat input, pump work, and cooling capacity on the expander efficiency for a fixed

output and establishes the motivation for the scroll expander performance study. The

influence of trace amounts of water in the vapor stream on cycle performance is also

investigated. The analytical portion of the results concludes with the calculation of the

maximum rate of hydrogen production.

Results of the scroll expander performance study are examined to predict the

expander's behavior with ammonia and to determine its feasibility for use in the

combined cycle. Several trends are developed to describe the performance of the scroll

expander. The data is compared to a performance chart of the same unit operated as a

compressor in order to determine if such information can reliably predict expander

performance.









Hydrogen Production and Liquefaction

Electrolysis of Water

Specific energy requirements for the electrolysis of water are displayed in Table

5.1. The majority of the electrical energy is required by the electrolyzer itself with the

subsystems representing only 16.2% of the total. Cooling water pump work is found to

be negligible compared to the energy consumed by the cooler (0.005 kW-h/lbm-H2

compared to 0.884 kW-h/lbm-H2). Including all subsystems, the total specific energy

required to electrolyze water is 24.839 kW/lbm-H2 (54.76 kW-h/kg-H2). Contrasting

with the energy requirement of thermoneutral electrolysis (17.865 kW-h/lbm-H2 (39.385

kW-h/kg-H2)), the electrolyzer has a thermal efficiency of 85.8%; however, the efficiency

drops to 71.9% when all subsystems are considered. At 85.8% electrolyzer efficiency,

the cell voltage required to drive the process is 1.713 V.

Table 5.1. Specific energy requirements of the IMET electrolyzer
Energy Requirements
kW-h/Ibm-H2 kW-h/kg-H2
Electrolyzer 20.814 45.886
AC/DC Rectifier 1.095 2.415
Cooling Water 0.844 1.860
Feed Water 2.081 4.589
Pump 0.005 0.010
Total 24.839 54.760

The amount of cooling water and feed water corresponding to their energy

consumption are 1.726 gpm/lbm-H2 (6.534 Lpm/lbm-H2) and 1.085 gal/lbm-H2 (4.107

Lpm/lbm-H2), respectively.

Hydrogen Liquefaction

Initial inspection of equations 3.13 and 3.15 indicate that the liquid yield and work

per unit mass hydrogen compressed are proportional to the expander mass flow ratio.

This is evidenced more clearly by defining the work per unit mass LH2 (Equation 5.1).









W (h2 -h) -Tl (s2 -s1) 7, (h3

wfhe=) n ---1-,, -h,+A )+-fN (5.1)
y hI -h2b + h3-h y 1

(State points referenced from Figure 2.6).

Equation 5.1 shows that increasing the expander mass flow ratio, xe, always

reduces the specific work for a given set of operating conditions; however, the amount of

liquid yield is physically constrained as described by Equation 5.2.

xe + y <1 (5.2)

The liquid yield continues to increase as defined by Equation 3.13 until the constraint is

met at which time it becomes a monotonically decreasing function of xe and T5. This

implies that an optimum value of the expander mass flow ratio exists at which the

liquefaction energy is minimized.

The exact form of the constraint is found from an analysis of the third heat

exchanger and the expansion valve. Heat exchanger cold side inlet and outlet

temperatures Tg = -423 F (-252.8 C) and T7 = -402.32 OF (-241.29 OC) are known from

the saturation temperature of hydrogen at atmospheric pressure and by assuming T7 = Te,

respectively. The "hot" side inlet temperature T4 = -402.32 OF (-241.29 OC) is equal to T7

because the flow passes through the 100% effective second heat exchanger as the

minimum capacity stream. The percent of the mass flow through the J-T valve that is

liquefied, k, is initially guessed as 80. T5 is then calculated from Equation 5.3 and used

to find the quality of the expanded stream. The value of k is iterated until convergence is

achieved.


T,= (- k)(T7 -k T )


(5.3)










Convergence is achieved in only three iterations with k = .725 and T5 = -408.05 F

(-244.47 C) because the temperature change of the supply stream is restricted by the

lower volume of the return stream. This exactly defines the constraint as:

y <.725(1- x) (5.4)
The optimum value of xe occurs when y exactly equals the constraint; an example of

which is seen in Figure 5.1.

8 .5 .

8 I

7.5 -

7 -

65.5 -



Y 5.5 optimum



4.5 -

4-

3.5
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1
Expander Mass Flow Ratio. X
Figure 5.1. Sample output showing the optimum expander mass flow ratio, xe

Prior to analyzing the effect of compressor inlet temperature and pressure

variations on the performance parameters of the Claude cycle, the expander and

compressor isentropic efficiencies are studied independently with regard to motivation

for further research and development of these components.

Expander efficiency. The effect of the expander isentropic efficiency on the

Claude cycle performance is summarized in Table 5.2. As an approximation, the liquid

yield constraint is held constant. In reality, however, the liquid yield is further











constrained with decreasing expander efficiency. At re = 0.4, the percent of the source


stream liquefied is approximately 48% compared to 72.5% for re = 0.85. The simulation

was run with the compressor efficiency fixed as 100% and an inlet temperature and

pressure of 80 F and 25 atm, respectively.

Table 5.2. Claude cycle simulation results for expander isentropic efficiency variation
Wftmin Wideal FOMmax
ie Xe,opt Ymax (z/Ypt kW-h/Ibm-LH2 kW-h/bm-LH F ma

1 0.5890 0.2975 15.386 3.684 1.268 0.3444
0.8 0.6350 0.2646 15.485 3.743 1.268 0.3389
0.6 0.6888 0.2256 15.638 3.835 1.268 0.3308
0.4 0.7522 0.1796 15.904 3.995 1.268 0.3175


Table 5.2 shows that an increase in the expander efficiency from 40% to 100% reduces

the optimum expander mass flow ratio by 21.7% from 0.7522 to 0.589. This shift in xe

increases the maximum liquid yield by 65.4% since it is related by the constraint of

Equation 5.4. The trend between xe,opt and ymax for different expander efficiencies is

observed in Figure 5.2. Table 5.2 also shows a 3.26% reduction of the cycle liquid

nitrogen requirement. The relationship between these two parameters is depicted in

Figure 5.3.


0.8
0.7
0.6
0.5
0.4
0.3
0 .1 -----------------------------_----------------
0.2
0.1
0
0.4 0.5 0.6 0.7 0.8 0.9 1
Expander Isentropic Efficiency

--Specific Liquid Yield --Expander Mass Flow Ratio

Figure 5.2. Specific liquid yield and expander mass flow ratio as functions of the
expander efficiency







58



16

E 15.9

| 15.8

15.7

|| :15.6
0E
S5 15.5

3 15.4
2'
15.3
0.4 0.5 0.6 0.7 0.8 0.9 1
Expander Isentropic Efficiency

Figure 5.3. Required liquid nitrogen vs. expander efficiency

The combination of these effects results in a 7.78% reduction in specific work as

described by equation 5.1. This is shown in Figure 5.4, as well as the shift in xe that

accompanies the decrease in specific work. The ideal work requirement depends only on

the inlet and liquid conditions and hence is unchanged; therefore, the FOM scales directly

with specific work.

8.5

rl=1.0
80.
[= 0.3
7.5 [- = 0.6
0- 0 = 0.4
7








4.5


4

3.5
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1
Expander Mass Flow Ratio, X
Figure 5.4. Specific work vs. expander mass flow ratio for varied qe






59


Compressor efficiency. By inspection of Equations 3.13, 3.14, and 5.1, it is clear

that the expander mass flow ratio, liquid yield, and liquid nitrogen requirement are

independent of compressor isentropic efficiency for fixed inlet and outlet conditions;

provided that the compressor can supply the necessary pressure. The simulation was run

with an expander efficiency of 100% and an inlet temperature and pressure of 80 F and

25 atm, respectively. Results are displayed in Table 5.3.

Table 5.3. Claude cycle simulation results for compressor isentropic efficiency variation
Tic Xe,opt Ymax Wf,min Wideal FOMmax
Tic Xe,opt Ymax (z/y)oPt kW-h/Ibm-LH2 kW-h/lbm-LH2 max
1 0.5890 0.2975 15.386 3.684 1.268 0.3444
0.8 0.5890 0.2975 15.386 3.746 1.268 0.3386
0.6 0.5890 0.2975 15.386 3.851 1.268 0.3294
0.4 0.5890 0.2975 15.386 4.059 1.268 0.3125

The only effect that increasing the compressor efficiency has is to lower the specific work

from 4.059 to 3.684 or 9.2%, as illustrated by Figure 5.5. The ideal work is again

independent of component efficiency and thus scales directly with the specific work.


Figure 5.5.


0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9
Expander Mass Flow Ratio, X
Specific work vs. expander mass flow ratio for varied q,










Comparing the figures of merit for the two cases in Figure 5.6 indicates a more

profound impact of the compressor efficiency on the cycle performance. A 60% decrease

in efficiency from ideal results in a 9.2% reduction in the FOM for the compressor case

compared to 7.8% for the expander. A greater emphasis should therefore be placed on

the development of high efficiency hydrogen compressors to minimize liquefaction

energy.


0.35
0.345
o 0.34
0.335
0.33
o 0.325
0.32
LL 0.315
0.31
0.4 0.5 0.6 0.7 0.8 0.9 1
Isentropic Efficiency
-- Expander Efficiency Effect -- Compressor Efficiency Effect

Figure 5.6. Impact of compressor and expander efficiencies on Claude cycle FOM

Compressor inlet pressure. Compressor inlet pressure is varied to determine the

advantage of using pressurized electrolysis. A simulation was run for an outlet pressure

of 40 atm and compressor and expander efficiencies of 75 and 85%, respectively.

Table 5.4. Claude cycle simulation results for compressor inlet pressure variation
1 X Y (zy) Wfmin Wideal FOMmax
atm eopt max pt kW-h/lbm-LH2 kW-h/lbm-LH2 ma
1 0.6230 0.2733 15.457 6.275 1.772 0.2824
5 0.6230 0.2733 15.457 5.050 1.521 0.3012
10 0.6230 0.2733 15.457 4.521 1.413 0.3125
15 0.6230 0.2733 15.457 4.210 1.349 0.3204
20 0.6230 0.2733 15.457 3.989 1.304 0.3268
25 0.6230 0.2733 15.457 3.817 1.268 0.3357









Table 5.4 shows that the optimum expander mass flow ratio, maximum liquid yield, and

minimum liquid nitrogen requirement are independent of the compressor inlet pressure.

This is true because the inlet temperature and exit pressure are held constant and the

compression process is modeled as isothermal. Furthermore, the liquid yield and

expander mass flow ratio are decoupled from the compressor inlet and exit temperatures

by the 100% effective cooling bath assumption. The functional relationship between the

specific work and inlet pressure is graphically described in Figure 5.7.


6.5

6 6
I 5.5


I4.5
4
S3.5


0 5 10 15 20 25 30
Compressor Inlet Pressure (atm)

Figure 5.7. Effect of compressor inlet pressure on the specific work

Increasing the pressure from 1 to 25 atmospheres reduces the specific work requirement

by 39.2% while also reducing the theoretical work requirement by 28.4%; the net result is

an 18.9% increase in the figure of merit.

Compressor inlet temperature. Compressor inlet temperatures ranging from 0

to 80 F are analyzed to gauge the merits for hydrogen pre-cooling using the combined

cycle. A simulation was run for an inlet pressure of 25 atm, exit pressure of 40 atm, and

compressor and expander efficiencies of 75 and 85%, respectively.










Results of the simulation are summarized in Table 5.5. An interesting result of

this simulation is the reduction in the figure of merit with lower inlet temperatures.

Table 5.5. Claude cycle simulation results for compressor inlet temperature variation
T X, Ymx () Wfmin Wideal FOM
(F) eopt max tz/Y kW-h/Ibm-LH2 kW-h/lbm-LH2 max
0 0.6230 0.2733 15.621 3.802 1.004 0.2639
20 0.6230 0.2733 15.577 3.806 1.069 0.2808
40 0.6230 0.2733 15.535 3.809 1.134 0.2978
60 0.6230 0.2733 15.495 3.813 1.201 0.3150
80 0.6230 0.2733 15.457 3.817 1.268 0.3323

This phenomenon is explained by the increase in the liquid nitrogen requirement due to

the assumption that To = T2. Reducing the inlet temperature of the compressor effectively

decreases To and lowers the enthalpy at state C. From Equation 3.14, the denominator he

- ha is reduced from 187.2 Btu/lbm (435.5 kJ/kg) to 167.3 (Btu/lbm) 389.2 kJ/kg, thereby

increasing the liquid nitrogen requirement from 15.457 to 15.621 lbm-LN2/lbm-LH2.

This effect partially offsets the benefit of pre-cooling the hydrogen gas resulting in a

0.393% decrease in specific work. The ideal specific work simultaneously decreases by

20.8%. As a result, the FOM actually decreases by 20.6% from 80 to 0 OF. The liquid

nitrogen requirement and specific work in relation to the compressor inlet temperature is

presented in Figures 5.8 and 5.9.


15.64
15.62 I
15.6 e
15.58 '
15.56 -$
15.54 E
15.52 E
15.5 -
15.48 3
15.46 g.
15.44 -
80 70 60 50 40 30 20 10 0
Compressor Inlet Temperature (F)

Figure 5.8. Liquid nitrogen requirement vs. compressor inlet temperature






63


Figure 5.9 concludes that the reduction in specific liquefaction energy due to

decreased inlet temperature can be neglected within the specified range. A qualitative

comparison of the figures of merit made between cases three and four show that

increasing inlet pressure is significantly more effective in reducing the specific

liquefaction energy (Figure 5.10).


3.82 E
3.818 -
3.816 S
3.814
3.812
3.81
3.808 _
3.806
3.804
3.802
3.8 |
80 70 60 50 40 30 20 10 0
Compressor Inlet Temperature (F)

Figure 5.9. Specific work requirement vs. compressor inlet temperature

The analysis concludes that the optimum operating point for the Claude cycle under

the assumed component efficiencies (qe = 0.85, rq = 0.75) is P1 = 25 atm. The cycle

parameter values are displayed in Table 5.6 at the optimum design point and at normal

operating conditions. The specific work of liquefaction is reduced by 46.7% and the

figure of merit is increased 18.9%.

The total energy required to produce and liquefy hydrogen is 28.656 kW-h/lbm-H2

(63.175 kW-h/kg-H2). Over 86% of this energy is consumed in electrolysis. Properties

at each state point are given for the optimum design condition in Appendix B. Given the

conclusion that compressor inlet temperature has little effect on the specific work, the










maximum hydrogen production rate of the combined cycle is 0.1775 Ibm/hr or 7.21

gallons/day.


0.34
0.33
S0.32
S0.31
0.3
4 0.29
0.28
i 0.27
0.26
0.25

Increasing Inlet Pressure -- Decreasing Inlet Temperature

Figure 5.10. Comparison of inlet pressure and temperature affect on the cycle FOM

Table 5.6. Claude cycle performance parameters for normal and optimum configuration
Normal Optimum
T = 80F P, = 1 atm T, =80F P, = 25 atm

Xe, opt 0.6230 0.6230
Ymax 0.2733 0.2733
(z/y)opt 15.457 15.457
Wf,min
Btu/Ibm-LH2 21411.70 13025.04
kW-h/Ibm-LH2 6.275 3.817

Wideal
Btu/Ibm-LH2 6046.88 4327.92
kW-h/Ibm-LH2 1.772 1.268
FOMmax 0.2824 0.3357


Ammonia-water Combined Cycle

The ammonia-water combined cycle was analyzed for a fixed power output to

observe the impact of expander efficiency on the heat and work requirements as well as

the cooling capacity. The simulation was run under the assumptions listed in chapter 3

while varying the expander isentropic efficiency from 10 to 100% in 10% increments.










Results of the simulation are given in Figures 5.11 thru 5.15 and a sample output of the

simulation is given in Appendix B.

The ammonia vapor mass flow rate required to drive the expander and produce

5kW of electricity is a function of only the exhaust enthalpy at state 8; since the power

output and specified temperatures and pressures are held constant. From the definition of

isentropic efficiency, the vapor mass flow scales with l/qe. The weak and strong solution

flow rates follow a similar trend as shown in Figure 5.11 as they are related to the vapor

flow by a constant ratio of the ammonia mass fractions as described by Equations 3.17

and 3.18. The maximum mass flows for the vapor, weak, and strong solutions occur at

the lowest efficiency (10%) and are are: 3586.34 lbm/hr (0.4519 kg/s), 72037.7 lbm/hr

(9.0767 kg/s), and 75624.1 lbm/hr (9.5286 kg/s), respectively.


x 104

Weak Solution
7 Strong Solution
-- Ammonia Vapor
6




6 -


-2



2 -- _1 1
1-

0
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1
Expander Isentropic Efficiency
Figure 5.11. Mass flow rate dependence on expander efficiency










Likewise the minimum mass flows are (in the same order): 358.63 Ibm/hr (0.0452

kg/s), 7203.77 (0.9077 kg/s), and 7562.41 (0.9529 kg/s). The mass fraction of the strong,

weak, and vapor streams are xs = 0.3988, xw = 0.3689, and xv = 1.0 (assumed). A

decrease in mass flow through the system manifests itself in the reduction of work and

heat interactions of the cycle for a given output as seen in Figures 5.12 and 13.


5.5 ...

5-

4.5 -

4 -

1 3.5 -

0 3 -
2\
E 2.5 -
CL ,
2 -

1.5 -



0.5
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1
Expander Isentropic Efficiency
Figure 5.12. Pump work variation with expander efficiency

Minimum pump work, boiler heat input and absorber heat rejection are: 0.724 Hp

(0.540 kW), 331,566 Btu/hr (97.18 kW), and 321,117 Btu/hr (94.11 kW). The ideal

cooling capacity under these conditions is 9130.26 Btu/hr (2.68 kW). Figure 5.14

concludes that at least 60% efficient expansion is required to obtain any cooling capacity.

Below this point, the exhaust temperature of the expander exceeds the assumed

temperature of the substance to be cooled (85 F). This effect is also evidenced in the

plot of the expander isentropic efficiency versus the cycle thermal efficiency (Figure

5.15).










x 10
3.5 I I
-- Boiler Heat Input, Qin
3- Absorber Heat Rejection, Qout


2.5

E 2


( 1.5
1-



0.5 -


0
0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1
Expander Isentropic Efficiency
Figure 5.13. Boiler heat input and absorber heat rejection vs. expander efficiency


10000 -

9000 -

8000

7000 /

S6000

5000 -

S4000

o 3000

2000

1000 /

0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1
Expander Isentropic Efficiency
Figure 5.14. Cycle cooling capacity as a function of expander efficiency

Thermal efficiency increases linearly with expander efficiency; however, a sudden

increase in slope occurs at approximately le = 0.6 at which point the cooling effect









begins to enhance the thermal efficiency of the cycle. The highest achievable thermal

efficiency for the given operating conditions is 7.22%.

0.08 ...i

0.07 -

D 0.06

0.05 -
E

0.03




0.01 /


0 0.1 0.2 0.3 0.4 0.5 0.6 0.7 0.8 0.9 1
Expander Isentropic Efficiency
Figure 5.15. Cycle thermal efficiency vs expander efficiency

The mass fraction of ammonia entering the expander was analyzed more closely to

judge the assumption of pure vapor leaving the rectifier and to determine the impact that

trace quantities of water have on the cycle performance. The analysis was carried out for

an ideal expander. Figure 5.16 shows the profound negative effect on cooling capacity.

The cooling capacity diminishes to zero rapidly as trace amounts of water are introduced

into the expander stream up to only 2.5% by mass. Boiler heat input is reduced from

331,566 Btu/hr (97.18 kW) to 320,934 Btu/hr (94.06); however, thermal efficiency is

reduced 9.26% from 7.22% to 6.55% because of the lost cooling benefit. Another

concern is that the expander exhaust temperature drops below the mixture dew point as

shown in Figure 5.17. At a 2.5% water vapor concentration by mass, the mixture quality

is 0.967. This most likely is not an issue for compliant devices such as scrolls in which a







69


small quantity of liquid can be tolerated, or with high-speed devices such as turbines, in

which the residence time of the fluid is shorter than the time required for condensation to

occur (metastable condition).


Figure 5.16.


10000
9000
8000
7000
6000 .
5000
4000
3000 cO
2000 =
1000 0
0
0.995 0.99 0.985 0.98 0.975
Expander Ammonia Mass Fraction, x7

Effect of trace amounts of water within in the expander inlet stream on
cycle cooling capacity


330
320
310 _
300
290
280 a.
270 E
260
250


1 0.995 0.99 0.985 0.98 0.975
Expander Ammonia Mass Fraction, x7

-- Isentropic Exhaust -*- Dew Point at 40 psia

Figure 5.17. Expander exhaust and dew point temperature at several water
concentrations

These results show that rectifier design is a crucial element for the success of a small-

scale combined cycle in the hydrogen production field in which high efficiency translates


into greater liquid yield per unit energy input.










Scroll Expander Performance Study

Scroll expander performance was measured for inlet pressures of 60, 70, and 80

psig; a range suitable for the 5kW combined cycle. Two tests were performed at each

pressure to verify repeatability of the results. Tests at pressures over 80 psig were not

feasible due to the relatively small tank and the inability of the compressor to supply

compressed air at high flow rates (> 60 scfm). Furthermore, the compressor motor is

equipped with a high-temperature cut-off switch that disconnects power after

approximately 10 minutes of continuous operation. A fan was used to aid in cooling the

motor; however periods of lockout continued to occur, limiting the maximum duration of

each test.

Figure 5.18 shows the results of the repeatability analysis applied to shaft power

measurements at 65 psig. The second set of data indicated by the square points agrees

well with the trend line of the initial data.


0.35

0.3

0.25
= .
0.2

0.15

0 0.1

0.05

0
0 500 1000 1500 2000 2500 3000
Rotational Speed (RPM)

Figure 5.18. Repeatability analysis applied to shaft power output at 65 psig

Results of the study are summarized in Figures 5.19 thru 5.22. Shaft power output

is plotted with respected to expander rotational speed in Figure 5.19. Rather than










beginning at a maximum value and decreasing monotonically with RPM as expected, the

power output reaches a maximum at approximately 1500 RPM before decreasing toward

zero in all three cases. This is thought to occur due to choked conditions at the expander

exit. Flow becomes choked when the port to fitting area ratio is smaller than the critical

area ratio given by the temperature and pressure of the exiting air. The area of the

expander exit port and fitting is 0.375" and 0.25", respectively. Further evidence of

choked flow is given by the fact that the maximum attainable rotational speed is only

3000 RPM at source pressures up to 110 psig, whereas the TRS-90 scroll compressor can

normally achieve speeds of up to 9000 RPM (Sanden engineer, personal conversation).


0.4
0.35

0.35--------------------- -------------
0.3
I X
0.25


0.15
0 0.1
0.05
0
0 500 1000 1500 2000 2500 3000
Rotational Speed (RPM)

*60 psi 70 psi X80 psi

Figure 5.19. Shaft power vs. rotational speed at 60, 70, and 80 psig inlet pressure

A similar trend is witnessed with isentropic efficiency, qe (Figure 5.20). Low

values of qe are attributed to the poor volumetric efficiency, qv, of the expander at low

RPM and relatively high torsional load. Increased torsional resistance raises the pressure

within each pocket of the scroll, enhancing tip leakage and reducing volumetric

efficiency. Figure 5.21 illustrates the relationship between volumetric efficiency and










rotational speed. At each pressure, rv increases asymptotically toward a final value

between 0.8 and 0.9.


0.2
0.18
> 0.16
S0.14 A
S0.12
S0.1 "
o 0.08
S0.06
S0.04

0.02

0 500 1000 1500 2000 2500 3000
Rotation Speed (RPM)

60 psi m 70 psi A 80 psi

Figure 5.20. Scroll expander isentropic efficiency

The volumetric efficiency indicates the percentage of air that passes through

without doing any useful work. This process can be modeled as isenthalpic, with the

approximation of constant temperature (ideal gas). The warmer air mixes with the cold

air, from which work was extracted, within the scroll housing effectively raising its

temperature prior to the measurement location. Furthermore, heat is exchanged from the

surroundings to the fluid through the exit port fittings. This temperature rise causes an

erroneous calculation of the exit enthalpy and thus the isentropic efficiency. However,

trends may still be observed to determine where the point of maximum efficiency occurs.

The exit temperature variation with rotational speed is shown in Figure 5.22. The

points of minimum exit temperature coincide with those of maximum power output as

expected from the First Law of Thermodynamics.












0.9
0.8
> 0.7
C X
a 0.6
X
w 0.5 -




> 0.2
0.1
0
0 -----------------------------

0 500 1000 1500 2000 2500 3000 3500
Rotational Speed (RPM)

*60 psi A70 psi x80 psi

Figure 5.21. Volumetric efficiency variation with expander rotational speed



50
48 -
-46
kL
44
42
)40 A
E38-
S36


32-
30
0 500 1000 1500 2000 2500 3000
Rotational Speed (RPM)

*60 psi m 70 psi A 80psi

Figure 5.22. Expander exit temperature and rotational speed relationship

The maximum power output of 0.368 Hp (0.274 kW) occurred at 1460 RPM for the


80-psig inlet pressure case. The most efficient operating point is 18.2%. Rotational


speed, inlet pressure, and power output at this point is 2000, 80 psig, and 0.282 Hp. The


temperature of the working fluid (excluding leakage) is found at any point using the


volumetric efficiency and flow rate in Equation 5.5.










T Tex CJ- leakageT nlet Txit (- rnlet (5)5)
Cf ~lv ~3

Therefore, with a volumetric efficiency of 0.6092 and temperatures of 71.6 F and 31.5

F at the inlet and exit at this point, the temperature of the working fluid is 4.77 F.

The low value of isentropic efficiency is due primarily to leakage caused by the

density mismatch. The TRS-90 is designed for R-134A with a density of 0.262 lbm/ft3 at

STP whereas the density of air at STP is .07298 lbm/ft3; nearly 3.6 times lower than R-

134A, and the density of ammonia is 0.04333 lbm/ft3; 1.6 times lower than air. The

performance of the expander with ammonia is expected to be worse than with air because

higher pressures are required for a unit volume of ammonia to store an equal amount of

energy as a unit volume of air at a given temperature. This relationship is arrived at by

considering the ideal gas law as a first approximation (Equation 5.6).

P P Pa/, ma.mmonla. 1.68 (5.6)
mT ammonia mT air ammonia air

Higher pressures lead to increased leakage within the scroll and a loss of performance.

Additionally, ammonia is a smaller molecule than air and much smaller than R-134A,

further facilitating tip leakage and reducing efficiency.

Fundamental design changes are required for the scroll concept to be utilized as

an expander. The geometry of each scroll element should be altered such that the total

number of chambers is increased as shown in Figure 5.23. This design reduces pressure

differences between chambers and hence leakage (Hans-Joachim and Radermacher,

2003).








75




44 -44 -
Cr, mp4r Expander



la II







5 -o40 -Ue -e 10 a 0 2 3w 4 60 4O -4 n -. -2 .11 v 10 0 D 4, O


Figure 5.23. Comparison of optimum geometries of a scroll compressor (left) and
expander (right) (Adapted from Hans-Joachim and Radermacher, 2003)














CHAPTER 6
RECOMMENDATIONS

Analytical Study

The analytical study of the electrolyzer, Claude cycle, and ammonia-based

combined power/cooling cycle examined a limited range of operating parameters. By

modeling the overall process with a program such as ASPEN, a greater number of

operating configurations could be analyzed.

ASPEN is a chemical processing software package that allows the user design a

cycle and specify a set or range of operating and boundary conditions. Using algorithms

included in the code for most devices, ASPEN performs a complete thermodynamic

analysis and outputs user specified data in an interactive manner.

Additionally, an optimization of the combined cycle for maximum hydrogen

production would indicate the operating conditions, power output, and overall system

size required to minimize energy cost. The economic viability of a large-scale

implementation of this system should be examined through a life-cycle cost analysis.

Scroll Expander Performance Test

The scroll expander used in the performance test was an automotive air-

conditioning compressor modified to run in reverse. Recommendations for future scroll

expander experimentation are:

13. Test the expander in a closed loop system with ammonia vapor
14. Pre-heat the inlet vapor to simulated the combined cycle operating conditions
15. Re-design the compressor housing to allow higher flow rates and eliminate choking
16. Design an oil injection and separation system to reduce leakage losses
17. Use a dynometer or motor to improve control on the applied torque






77



Future work should also include improvements to the scroll design. Manufacturing

the scroll involute using the optimum expander geometry shown in Figure 5.23 would

improve its performance as an expander. Furthermore, the use of low-friction materials

such as those under development at the University of Florida would eliminate the need

for an oiling system, making the scroll an attractive design for the ammonia based

combined power/cooling cycle.














CHAPTER 7
CONCLUSIONS

Global energy consumption is projected to increase 54% over the next 25 years.

With proven oil reserves being called into question beyond 2030 it is important to

develop renewable technologies to sustain the future global energy demand. By

introducing an alternative fuel for transportation only, oil consumption can be reduced by

as much as 20%.

Hydrogen has many characteristics that make it a desirable fuel. It has the highest

energy content per unit mass of any known fuel nearly 3 times higher than gasoline, it

burns cleanly and efficiently, and it can be produced from water via electrolysis powered

by renewable energy. Two major obstacles to the emergence of a hydrogen economy are

the limited means available to efficiency produce mass quantities of hydrogen from

renewable energy sources and the storage issues related to the low energy density of

hydrogen. Liquefying hydrogen provides a solution to its low density; however, the

process requires additional energy.

This thesis explored the possibility of using a 5-kW ammonia-based combined

power/cooling cycle to produce hydrogen from renewable resources and pre-cool it prior

to liquefaction in an effort to reduce the overall energy consumption. The advantage of

this cycle is its ability to utilize low temperature heat sources available from solar and

geothermal resources.

Simulations of the Claude liquefaction process and the 5-kW ammonia-based

combined power/cooling cycle were developed to model the effects of component









efficiencies and operating parameters on the maximum hydrogen production rate and

system energy requirement. Additionally, a performance test of a scroll compressor was

performed to gauge its effectiveness as an expander for the combined cycle.

Conclusions resulting from tests and analyses are summarized below:

1. Pre-cooling hydrogen has little effect on the specific liquefaction energy and is

detrimental to the liquefier efficiency.

2. Pressurized electrolysis is the most effective method of reducing the energy

consumed in liquefaction.

3. The total energy required to produce and liquefy hydrogen is 28.656 kW-h/lbm-H2

(63.175 kW-h/kg-H2); 86% of which is consumed during electrolysis. A maximum of

7.21 gallons (27.3 liters) per day of liquid hydrogen can be produced from a 5-kW

combined cycle.

4. The mass flows as well as the heat and work interactions of the 5-kW combined cycle

scale with inverse of expander efficiency (l/rle). Sixty percent expansion efficiency

is required to extract cooling from the cycle.

5. Cooling capacity of the cycle is extremely sensitive to the vapor mass fraction of the

expander inlet stream. At 2.5% water by mass and for perfect expansion, the cooling

capacity completely diminishes.

6. Results of the performance test indicate that scroll compressors operate poorly as

expanders. Low isentropic efficiencies result from leakage around the scroll tips.

Improvements in the scroll design such as increasing the wrap of each scroll element

and using low-friction material for oil-less operation would make the scroll an

efficient expansion device suitable for the combined cycle.














APPENDIX A
COMPUTER PROGRAM FOR CYCLE SIMULATIONS

Two computer programs were written to assist in the evaluation of thermodynamic

properties and to perform cycle analyses of the Claude liquefaction cycle and the

ammonia-water combined power/cooling cycle. A description of each program is given

below, including portions of the source code.

Claude Cycle Simulation

The program was written to assist in the parametric analysis of the specific work

and efficiency of the Claude cycle. A subroutine was included to evaluate the

thermodynamic properties at each state point coinciding with Figure 2.6. The code has

the flexibility of single point calculations or variable inputs for a parametric analysis.

Thermodynamic Property Evaluation

The property code incorporates portions of RGAS and PSAT, two programs written

by Dr. Roger Gater (2001). Property evaluation is carried out as a subroutine of the

overall cycle simulation. The properties defined by user input and the listed assumptions

are passed into either routine depending on the fluid condition. For saturated conditions,

the pressure is defined; for superheated vapor, pressure-temperature, pressure-enthalpy,

or pressure-entropy is input. Properties are then evaluated using the Redlich-Kwong gas

model and returned to the main program. Critical properties and coefficients required by

the Redlich-Kwong model are listed in Table A. 1.










Program Description

The Claude cycle simulation program is written in MatLAB. It consists of three

sub-routines and a data file: saturation2.m, gas_properties.m, gas_propertiesbase.m,

gas.dat, all of which must be present for the program to operate. The program begins by

reading data from the "gas.dat" file. It then asks for user input of compressor inlet

temperature and pressure; giving the option of English or SI units. From the user input

and given assumptions, the thermodynamic properties at each state point are evaluated by

the "gas_properties.m" subroutine. If saturated conditions are known to exist,

"saturation2.m is invoked. The key performance parameters of the Claude cycle are then

calculated using the equations of Chapter 3. Results are output to the screen in figure

form. Additional aspects of the program are described by the imbedded comments.

Main Program Claude.m

[gas_num gas_name R Tc Pc cpoR a b c Zc A w] = textread('GAS.dat','%f %s %of %f %f %f of f % f %ff
%f %of, 'headerlines', );

units = input('Select Units: 0 = Metric, 1 = English: ');
while (units < 0) (units > 1)
units = input('\nError, Try again: ');
end

if units == 0
T1 = input('\nEnter compressor inlet temperature (K): ');
P1 = input('\nEnter compressor inlet pressure (atm): '); P1 = P1 1.0132;
else
T1 = input('\nEnter compressor inlet temperature (F): ');
T1 = (T1 32)*5/9 + 273.15;
P1 = input('\nEnter compressor inlet pressure (atm): '); P1 = P1 1.0132;
end

P2 = input('\nEnter compressor discharge pressure (atm): '); P2 = P2 1.0132;
etae = input('\nEnter expander adiabatic efficiency: ');
%eta e = .85;
etac = input('\nEnter compressor efficiency: ');
%eta c = .75;
% properties in J/g or kJ/kg
P_stp = 1.0132; %bar
gas = 10; % selects hydrogen gas from GAS.DAT data file

Pe = P_stp; Pg = P_stp; Pf = P_stp; P7 = P_stp; P8 = P_stp;










P9 = P_stp; P10 = P_stp; PlOa = P_stp; PA = P_stp; PC = P_stp;
P2b = P2; P3 = P2; P4 = P2; P5 = P2;

HC1 = 175.1; HC2 = 312.97; %Heats of conversion kJ/kg
WN2 = 1783.623; %kJ/kg-N2 Energy of LN2 liquefaction

T2 = T1;T10 = T1;TC= T1;

%State g and f
%call saturation program
routine = 1;
Ps = Pg; P = Pg;
[Ts,Zf,Zg,vf,vg,hfg,ufg,sfg] = saturation2(Ps,R(gas),Tc(gas),Pc(gas),Zc(gas),A(gas),w(gas));
Tg = Ts; Tf = Ts; T = Ts;
gas_properties;
hg = h; sg = s;
hf= hg hfg; sf= sg sfg;

%State 1
routine = 1; pressuree and temp specified
T= T1;P = P1;
gas_properties;
hi = h; sl = s;

%State 2
routine = 1; isothermall compression
T = T2; P = P2;
gas_properties;
h2 = h; s2 = s;

%State 2b
Ps = P_stp;
gas =13; %sets nitrogen properties
Ts = saturation2(Ps,R(gas),Tc(gas),Pc(gas),Zc(gas),A(gas),w(gas)); %calculates temperature of nitrogen
gas =10; %returns to hydrogen
T2b = Ts; T = T2b; P = P2b;
routine = 1;
gas_properties;
h2b = h; s2b = s;

%State 3
T3 = T2b;
test = 0;
while T3 >= 70 % above critical temperature of hydrogen (asymptotic problems)
routine = 1;
T = T3; P =P3;
gas_properties;
h3 = h; s3 = s;
%state e s
routine = 2; %test to see if saturated conditions exist
P = Pe; Ps = Pe;
se = s3; ss = se;
gas_properties;
he s = h; Te s = T;
ifTe s <= Tf
xe = (se sf)/(sfg);











he_s = hf+ xe*(hfg);
Te_s = Tf; %isentropic temperature
end
if (h3 hes) > test
T3opt = T3; Te_s_opt = Te_s;
delta h opt = h3 he_s;
h3opt = h3; s3opt = s3; he_sopt = he_s;
end
test = h3 he s;
T3 = T3 .1;

end
%state e
routine = 3; % pressure and enthalpy specified
he = h3opt etae*(delta h opt);
hh = he; P = Pe;
gas_properties;
Te = T; se = s;

%state 4
routine = 1;
T= Te;P =P1;
T4 = Te;
gas_properties;
h4 = h; s4 = s;

%state 7 and 8
routine = 1;
T = Te; P = Pe;
T7 = Te; T8 = Te;
gas_properties;
h7 = h; h8 = h;
s7 = s; s8 = s;

%state 10
routine = 1;
T= T10; P =P10;
gas_properties;
hlO = h; slO = s;

%state 10a
routine = 1;
T = T2b; P = P10;
gas_properties;
hl0a = h; sl0a = s;

%state A (saturated liquid)
routine = 1;
Ps = PA; P = PA;
gas = 13;
[Ts,Zf,Zg,vf,vg,hfg,ufg,sfg] = saturation2(Ps,R(gas),Tc(gas),Pc(gas),Zc(gas),A(gas),w(gas));
TA = Ts;
gas_properties;
hgA = h; sgA = s;
hA = hgA hfg; sA = sgA sfg;










%state C
routine = 1;
T = TC; P = PC;
gas_properties;
hC = h; sC = s;
gas =10; %return to hydrogen

%Specific work, liquid yield, liquid nitrogen requirement, and figure of merit calculation
X(1)= 0;
step =.001
for i = 1:/step
X(i+l) = X(i) + step;
y(i) = ((hl0a-h2b) + etae*X(i)*(deltahopt))/(hlOa hf + HC2);
if y(i) >= .725*(1 X(i)) %.725 found from iterative procedure on HX 3
y(i) = .725*(1-X(i));
end
z(i) = (HC1 + (h2 hO0) + etae*X(i)*(delta_h opt) + y(i)*(hl0 hf + HC2))/(hC hA);

if units == 0
W(i) = (((h2 hi) Tl*(s2 sl))/eta c eta e*X(i)*(delta h opt))/3600; %work per unit mass
compressed
Wf(i) = W(i)/y(i) + z(i)/y(i)*WN2/3600; %work per unit mass liquefied kJ/kg
Wideal= ((hf- hl) Tl*(sf- sl))/3600;
else
W(i) = (((h2 hl) Tl*(s2 sl))/eta c eta e*X(i)*(delta h opt))/(2.326*3412);
Wf(i)= W(i)/y(i) + z(i)/y(i)*WN2/(2.236*3412);
Wideal= ((hf- hl) Tl*(sf- sl))/(2.326*3412);
end

FOM(i) = Wideal/Wf(i);
if X(i+) >= .9
break
end
end

z(i+l) = z(i);
Wf(i+l) = Wf(i);
W(i+l) = W(i);
FOM(i+1) = FOM(i);
y(i+l) = y(i);

Xopt = (1 max(y)/.725)
Ymax = max(y)
Wfopt = min(Wf)
FOMopt = max(FOM)
zopt = min(z./y)
W ideal

figure(l)
plot(X,Wf)
title('Work Per Unit Mass LH2 vs. Expander Mass Flow Ratio, X');
xlabel('Expander Mass Flow Ratio, X')
ylabel('Wf [kW-h/lbm-LH2]')