LOW TEMPERATURE POWER GENERATION USING HFE-7000 IN A RANKINE CYCLE
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A Thesis Presented to the Faculty of San Diego State University _______________
In Partial Fulfillment of the Requirements for the Degree Master of Science in Mechanical Engineering _______________
by Austin D. Reid Fall 2010
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Copyright © 2010 by Austin D. Reid All Rights Reserved
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ABSTRACT OF THE THESIS Low Temperature Power Generation Using HFE-7000 in a Rankine Cycle by Austin D. Reid Master of Science in Mechanical Engineering San Diego State University, 2010 This thesis presents the modeling, construction, and testing of a cost-effective Rankine cycle that utilizes a low-grade heat source (100°C); a non-toxic, environmentally benign working fluid (heptafluoropropyl methyl ether, or HFE-7000); and a scroll-type expander to generate electric power. A thermal efficiency of 3.1% was achieved, and a clear path forward has been presented to achieve 9.7% efficiency. This experiment validates the use of HFE-7000 as a working fluid in a Rankine cycle – both in a custom application or as a dropin replacement for less-desirable fluids. Such a system has potential uses that include bottoming cycles for industrial processes and inexpensive non-concentrating solar-thermal power plants.
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TABLE OF CONTENTS PAGE ABSTRACT ............................................................................................................................. iv LIST OF TABLES .................................................................................................................. vii LIST OF FIGURES ............................................................................................................... viii LIST OF ABBREVIATIONS .................................................................................................. ix ACKNOWLEDGEMENTS ..................................................................................................... xi CHAPTER 1
INTRODUCTION .........................................................................................................1 Review of Literature ................................................................................................2 Proposed Research ...................................................................................................7
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SYSTEM MODELING .................................................................................................8
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SYSTEM DEVELOPMENT .......................................................................................14 Heat Source ............................................................................................................15 Steam................................................................................................................15 Internal Combustion Engine ............................................................................15 Water Heater Source ........................................................................................16 Heat Rejection Sink ...............................................................................................18 System Components...............................................................................................19 Heat Exchangers ..............................................................................................19 Pump ................................................................................................................19 Expander ..........................................................................................................20 Multi-Vane Expander.......................................................................................20 Screw Expander ...............................................................................................21 Scroll Expander ................................................................................................23 Turbo Expander ...............................................................................................25 Expander Selection ..........................................................................................26 Compressor Conversion ...................................................................................27 Plumbing Configuration.........................................................................................31
vi Electrical Load .......................................................................................................33 Measurement ..........................................................................................................35 Working Fluid ........................................................................................................36 Performance .....................................................................................................36 Environmental Impact ......................................................................................37 Safety ...............................................................................................................39 Fluid Selection .................................................................................................39 4
RESULTS AND DISCUSSION ..................................................................................41 Experimental Data .................................................................................................41 Update to Theoretical Model .................................................................................42 Data Comparison ...................................................................................................43 Discussion ..............................................................................................................44 Pump Performance ...........................................................................................44 Expander ..........................................................................................................46 Heat Source and Sink .......................................................................................47 Overall Efficiency Gains..................................................................................47 HFE-7000 .........................................................................................................48 Cost Analysis ...................................................................................................49
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CONCLUSIONS..........................................................................................................51 Summary ................................................................................................................51 Recommendations ..................................................................................................52
REFERENCES ........................................................................................................................53 APPENDIX SCROLL DEVICE PROPERTIES ..............................................................................57
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LIST OF TABLES PAGE Table 1. Theoretical State Calculations ...................................................................................12 Table 2. Expander Comparison Summary ...............................................................................26 Table 3. Environmental Factors for Possible Working Fluids .................................................39 Table 4. Summary of Experimental Data ................................................................................41 Table 5. Experimental State Calculations ................................................................................43 Table 6. Optimized State Calculations.....................................................................................48 Table 7. Comparison of Refrigerant Density ...........................................................................49 Table 8. System Cost Breakdown ............................................................................................49
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LIST OF FIGURES PAGE Figure 1. System schematic. ......................................................................................................8 Figure 2. P-H diagram for HFE-7000. .......................................................................................9 Figure 3. P-H diagram of theoretical ORC states. ...................................................................13 Figure 4. ORC system at outset of project. ..............................................................................14 Figure 5. Heat source installed.................................................................................................17 Figure 6. Heat sink installed. ...................................................................................................18 Figure 7. Multi-vane expander. ................................................................................................20 Figure 8. Screw expansion device (oil injected). [28] .............................................................22 Figure 9. Scroll expander operation. ........................................................................................23 Figure 10. Expander schematic. ...............................................................................................27 Figure 11. Expander with top removed and check valve assembly. ........................................28 Figure 12. Scroll device sides A and B. ...................................................................................28 Figure 13. Compressor flow diagram. .....................................................................................29 Figure 14. Expander flow diagram. .........................................................................................30 Figure 15. Scroll device installed with inlet machined. ...........................................................31 Figure 16. Cap before and after modifications. .......................................................................31 Figure 17. Expander installed in ORC system. ........................................................................32 Figure 18. Pump to pre-heater link before and after modification. .........................................33 Figure 19. Run capacitor wiring. .............................................................................................34 Figure 20. Electrical setup diagram. ........................................................................................35 Figure 21. 3-phase load with clamp-on ammeter and switches. ..............................................35 Figure 22. P-H diagram of actual ORC states..........................................................................44 Figure 23. Pump performance curve. .......................................................................................45 Figure 24. Compressor performance data under various conditions. ......................................58 Figure 25. Compressor specifications. .....................................................................................59
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LIST OF ABBREVIATIONS ALT – Atmospheric Life Time ORC – Organic Rankine Cycle GWP – Global Warming Potential ODP – Ozone Depletion Potential STP – Standard Temperature and Pressure GPM – Gallons per Minute CFH – Cubic Feet per Hour HFC - Hydrofluorocarbon HFE – Hydrofluoroether HCFC – Hydrochlorofluorocarbon LFL –Lower Flammability Limit PV - Photovoltaic P, p – Pressure (kilopascals) T – Temperature (F, C, and K alternately) H, h – Enthalpy (kJ/kg) – Heat Transfer (kW) – Work (kW) S, s – Entropy (J/mol-K) η – Efficiency (%) , mdot – Mass flow rate (kg/s) cp – Heat capacity (J/kg-K) Subscripts: 1,2,3,4 – States e - expander gen – generator H2O - Water
x in – inlet, supply out – outlet, sink p – pump th – thermal
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ACKNOWLEDGEMENTS I wish to acknowledge the work of Terry Ireland, Mike Lester, Bill Lekas, and everyone at the SDSU Physical Plant that helped make the hardware functional.
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CHAPTER 1 INTRODUCTION Many industrial processes produce waste heat that is typically rejected to a lower temperature heat sink, but can instead be recovered to produce useful energy. This waste energy is often rejected at lower temperatures than would be useful for typical energy conversion systems. Recovery of waste heat offers the benefit of increasing overall efficiency in the case of a power generation, or of providing auxiliary power in other waste heat application. The standard Rankine Cycle, often used for primary power generation using a heat source, operates at relatively high temperatures (250-600°C) in order to maximize Carnot efficiency, rendering many lower-temperature heat sources useless as energy sources. However Organic Rankine Cycles (ORC), which use organic working fluids rather than water or other fluids, can operate at low enough temperatures to take advantage of lower-temperature sources. In addition to industrial waste heat, solar sources are also wont to produce lowertemperature energy. To achieve high-temperature from a solar source typically requires high concentration ratios produced by large-area, and thus expensive, reflective concentrators. A lower temperature requirement allows a solar collector to be a smaller or less efficient concentrator, or possible even a direct-heated system without any concentrating at all. The latter would be an extremely low-cost heat source. The state of the art in ORCs has several limitations. Many of the working fluids are ozone depleting (HCFC fluids) or toxic. Components are expensive relative to power produced – a turbine expander is appropriate for use in a steam Rankine Cycle that produces electric power in the MW range, but is too costly for a system that only produces power on the kW scale. Further, the minimum working temperature of many cycles that have been investigated, often above 200°F (Dai, Wang, & Gao, 2009; Hettiarachchi, Golubovic, Worek, & Ikegami, 2007; Hung, Shai, & Wang, 1997), is still too high for many applications, including some waste heat recovery and solar sources.
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REVIEW OF LITERATURE Much research has been done to investigate the use of Rankine cycles to capture lowquality heat. While alternative fluids have been investigated, such as Zeotropic mixtures (Wang & Zhai, 2009), the vast majority of the research focuses on organic working fluids. Organic working fluids offer better performance at lower temperatures than typical inorganic working fluids, namely water (Dai et al., 2009; Yamamoto, Furahata, Arai, & Mori, 2001). Many of the studies performed are purely theoretical, with no real-world validation of thermodynamic models. Hettiarachchia et al. (2007) evaluated design criteria such as working fluid, evaporation and condensation temperatures, and system cost to optimize the ORC for geothermal sources, concluding that ammonia is the primary candidate, followed by HCFC 123 for heat sources nearing 100°C. Liu, Chien, and Wang (2004) performed an analysis of several working fluids at various evaporator temperatures and found that evaporator efficiency is maximized by fluids with low enthalpies of vaporization, and thus these fluids are most appropriate for an ORC. Mago, Srinivqasa, Chamra, and Somayaji (2008) used an exergy analysis to increase theoretical efficiency in an ORC using regenerative heating, concluding that the evaporator contributes most to exergy destruction in an ORC. Some research has been done to improve the accuracy of existing models using computational and finite element methods. This higher resolution can be used to improve the design of a Rankine system without incurring the cost of a real-world experiment. Vaja and Gambarotta (2010a; 2010b) performed extensive modeling of the ORC system, producing complete Simulink libraries of components for future system modeling. Particular focus was placed on the evaporator and condenser, resulting in an accurate parametric finite difference model. Their model consists of “state determined” (evaporator and condenser) and “not state determined” (pump and expander) to avoid algebraic loops and iterative numerical solutions. A set of detailed simulation results is produced for HCFC123, but no validation has been done to date. The theoretical system operates at 200°C, producing greater that 275kW at 18% thermal efficiency. Toffolo, Lazzarettoa, Manentea, and Rossi (2010) took a similar approach by decoupling the heat transfer portions of the cycle (evaporator and condenser) from the other components, and using the “HEATSEP” model to characterize the heat transfer. Working
3 fluids isobutene and R134a were considered with heat source temperatures down to 130°C, and R134a was found to have higher exergy recovery coefficients. Much research has gone into the selection of working fluids for various cycles based on fluid characteristics. One such characteristic is the fluid’s saturated vapor curve. Many papers focus solely on “dry” working fluids – fluids for which the saturated vapor curve is positive in a temperature-entropy (T-s) diagram. Dry fluids are preferable to “wet” fluids because they do not condense after the sudden enthalpy drop through the expander. Isentropic fluids have an approximately vertical vapor saturation curve. Hung et al. (1997) analyzed a series of wet, dry, and isentropic fluids over various temperatures and pressures, and more recently Hung (2001) focused on dry fluids in “Waste heat recovery of organic Rankine cycle using dry Fluids,” concluding that that isentropic fluids are most appropriate for use with low temperature heat sources. Mago et al. (2008) also focused on dry fluids, while still incorporating regenerative heating, in “An examination of regenerative organic Rankine cycles using dry fluids”, concluding that regeneration is appropriate for an ORC, increasing efficiency and decreasing irreversibilities. Mago et al. also show that superheating dry fluids adds irreversibility to the system – with no risk of condensation in a dry fluid, the fluid should enter the expander as saturated vapor rather than a superheated vapor. In “Investigation of the criteria for fluid selection in Rankine cycles for waste heat recovery”, Siddiqi and Atakan (2010) suggest that T−H diagrams are the best and most intuitive way to judge the performance or the weakness a fluid for a particular heat source and heat sink. However many of the fluorinated hydrocarbons were found to be inappropriate for the high temperature heat sources considered in the research. Many fluids are common for use in ORCs, but some papers investigate novel or previously overlooked fluids. Much of this is driven by environmental laws that phase out heritage fluids. In “HFC-245fa Working Fluid in Organic Rankine Cycle - A Safe and Economic Way to Generate Electricity from Waste Heat”, Zyhowski, Brown, and Achaichia (2010) evaluate HFC-245fa as a potential safe and benign working fluid. They conclude that the fluid could perform efficiently in an ORC due to its high heat capacity and gas density. The fluid is found to outperform water in systems operating with an approximately 150°C heat source and creating 50kW.
4 Husband and Beyene (2008) address environmental effects and component costs in “Low-grade heat-driven Rankine cycle, a feasibility study.” The paper demonstrates the feasibility of a low-grade heat recovery system that can produce electrical power using a relatively benign working fluid. Specifically, a hydrofluroether – heptafluoropropyl methyl ether (HFE-7000 or Novec 7000) has low global warming potential (GWP), zero ozone depletion potential (ODP), low ALT, low toxicity, and zero flammability. A theoretical efficiency of 11% for a 10kW system was shown. Further, a scroll device was suggested as an expander. The type of expander used in the cycle is another common area of research. Papers have focused on several common expanders, including multi-vane, which performs well and is low-cost (O’Callaghan, Badr, Probert, Bell, & Patel, 1985); turbo, which is high cost and high performing, but largely unproven for low temperature power generation (Teagan & Clay, 1973); and screw, which performs well with wet fluids though the cost remains somewhat high (Smith, Stosic, & Kovacevic, 1999). The benefits and drawbacks of each of these devices are described in further detail in the Expander section below. Some novel devices have been investigated as well. Giampaolo and Stefano (2010) research the Wankel device in “Model of a steam Wankel Expander.” Though in its infancy in this application, the Wankel expander shows some promise in low-temperature applications, achieving a theoretical system efficiencies greater than 10% with water working fluid at 150°C. Another device investigated by Giampaolo and Leonardo (2010) is a reciprocating engine operating with a heat source between 100 and 150°C. Results of the model indicate that the expansion device may be promising in a low-pressure (7 bar) steam application producing power up to 10kW. The scroll device is one particular expander that has been the subject of myriad research, both theoretical and experimental. Clemente, Micheli, Reini, and Taccani (2010) have developed a numerical model of a scroll device capable of estimating the performance of the device as both expander and compressor. Theoretical calculations have been compared to a set of experimental data which validate the accuracy of the model for a range of input parameters such as expansion ratio and shaft speed, as well as specific common working fluids.
5 Yanagisawa, Fukuta, Ogo, and Hikichi (2001) found that the most dominant factor affecting scroll expander efficiency is not leakage loss but mechanical loss from the orbiting motion; pressure in the scroll pocket drops due to a throttling effect when the wrap opens before the expansion occurs. Zanelli and Favrat (1994) detail the conversion of a 1 to 3.5 kWe scroll expander from a compressor, and the use of an organic Rankine cycle test facility for expanders up to 10 kWe. Expander efficiencies reached with HFC-134a reach 65% with shaft speeds between 2400 and 3600 RPM. Low shaft research has been performed by Kim, Ahn, Park, and Rha (2007), achieving experimental and theoretical efficiencies of 34 and 65% respectively at shaft speeds below 1400 RPM. Kane, Favrat, Gay, and Andres (2007) measured efficiency of an ORC at about 7% with a heat source at around 90°C for a 7kWe system. This research, though large in scale, has many of the features in common with research presented in this thesis. Saitoh, Yamada, and Wakashima (2007) report a total thermal efficiency of 7% in a solar powered system using a scroll expander and a concentrating parabolic reflector. Peterson, Wang, and Herron (2008) research a lower power system that produces from 187 to 256 W. At nominal ambient outside temperatures (22.5 °C), the system efficiency was 7.2%. The authors found that the expander limited the overall system efficiency with efficiencies between 45 and 50%. Lemort and Quoilin of the University of Liege, Belgium are responsible for significant experimental work, performing extensive testing with ORC systems utilizing a scroll expansion device. Between two studies (Lemort & Quoilin, 2009; Quoilin, Lemort, & Lebrun, 2010), over 40 steady state performance points were gathered for HCFC-123 and water evaluating the effects of boiler temperature, mass flow rate, condenser temperature, and expander rotational speed on shaft power, the expander isentropic effectiveness, and the cycle efficiency. Expander efficiencies ranging from 42% to 68% were observed yielding system efficiency up to 7.4% with heat source temperatures down to 120°C. The research concludes that HCFC-123 offers superior performance with the scroll device relative to water due to higher volumetric performance and the lower under-expansion losses. The authors further recommend the use of HFC-245fa as HCFC-123 is phased out.
6 Overall, it is observed that much research has been performed with scroll expansion devices, with a wide range of input temperatures, fluids, and system sizes. Experimental ORC systems of the approximate size and temperature range as the system of interest in this thesis report efficiencies between 7 and 8%. Additional validation and real-world experimentation has been done to investigate the operation of the ORC. Yamamoto et al. (2001) created hardware to validate their fluid of choice, HCFC-123, versus water. They were able to achieve 1.25% efficiency. An intensive empirical study was performed by Quoilin (2007) in “Experimental Study and Modeling of a Low Temperature Rankine Cycle for Small Scale Cogeneration.” Several variables were considered, including hot air source temperature, expander rotational speed, and refrigerant charge. Refrigerant charge was found to have a significant impact on the performance of the cycle. One commercial system that has been installed is a low-temperature but large-scale power plant installed at Chena hot springs in Alaska (Cogswell, 2006). Operating with a heat source of only 73°C, the system uses turbo expanders to generate 200kW at a net thermal efficiency of over 8%. Further, the cost of this system is minimized by using commercial refrigeration components. The expansion device is a converted in-line centrifugal compressor. Several papers have specifically investigated the cost effectiveness of ORC systems. Husband and Beyene (2008) perform component research and economic considerations and conclude that a 10kW solar driven ORC may be cost competitive when weighed against a comparable PV system. Tchanche, Quoilin, Declaye, Papadakis, and Lemort (2010) perform an exhaustive cost and performance analysis for a variety of fluids and parameters in “Economic Optimization of Small Scale Organic Rankine Cycles.” The authors found a strong correlation between pressure ratio and both cost and power output, however the respective minimum and maximum did not match. The operating point for maximum power does not match that for minimum investment cost. The total investment cost includes costs of a scroll expander, evaporator, condenser, fluid pump, pipes, working fluid charge, water cooling pump, liquid reservoir, control system, miscellaneous hardware and a labor cost equal to 10%
7 of the total material cost. Many fluids were considered including HCFC-123 and HFC245fa, which both achieve investment costs below 5000 Euro/kWe (6500 $/kWe). Smith, Stosic, Kovacevic, and Langson (2007) also performed a cost analysis for a larger system, up to 20kW, and found that the installed cost per kW was as low as 2500$/kWe, much lower than the cost estimated by Tchanche, but at a much larger scale. Such a system was found to be approximately 30% cheaper than a 200kW turbo expander system quoted for the research, and could continue to be cheaper up to 500kW. The research investigated heat source temperatures down to 90°C, and focused on the screw expansion device. The primary goal of the Chena project, and one which was achieved, was to reduce the cost of the installed ORC system to 1300 $/kW (Cogswell, 2006). All of these areas: fluid selection, expansion device, theoretical modeling, experimental demonstration, and cost effectiveness will be considered in this thesis.
PROPOSED RESEARCH This research will build upon the work done by Husband and Beyene. Using the theoretical model as a starting point, an existing system optimized for use with HCFC-134 working fluid will be modified to run with HFE-7000. The operational system will be measured and analyzed to establish overall system efficiency, sources of irreversibility, and potential areas for improvement. Successful research will provide four immediate benefits:
The use of an HFE fluid in an ORC will be demonstrated for the first time.
The use of a low-temperature power source (less than 100°C) will be demonstrated.
The existing mathematical model for HFE-7000 will be validated or modified, allowing future systems to be designed using more ideal parameters such as temperature and pressure, or other components.
Inexpensive components will be used to demonstrate cost effectiveness of the ORC relative to photovoltaic systems. This research will require procurement of components and materials (including HFE-
7000), construction and assembly of the system, leak-checking and fluid-charging of the system, location and connection of heat source and sink, and finally taking of temperature, pressure, and electric measurements during system operation.
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CHAPTER 2 SYSTEM MODELING The theoretical analysis of the fluid is a state analysis – at certain points in the system the state of the fluid is calculated. Each state includes all properties of the fluid such as temperature, pressure, enthalpy, entropy, and specific volume, but can be described with only two – in this case, temperature and pressure. Through each component, certain properties of the fluid change. Through the pump, pressure increases; through the condenser, temperature decreases; etc. Although the analysis is theoretical, a system has already been assembled, so certain component properties are already known. This differs from a typical analysis in that properties are imposed on the model rather than using the model to select components with ideal properties. The system schematic can be seen in Figure 1, with the states labeled.
Figure 1. System schematic.
9 HFE-7000 is unique in the ORC application because of its high boiling point. A typical refrigerant used in a similar application, such as R-134, will exist in the vapor phase at standard temperature and pressure (STP). However HFE-7000 will not boil at atmospheric pressure until 96.5°F. The pressure-enthalpy diagram is shown in Figure 2 with lines of constant temperature, specific volume, and entropy. The positive slope of the saturated vapor curve can be clearly seen, placing the fluid definitively in the “dry” category – as the fluid proceeds through the expander and experiences a sudden reduction in pressure, the fluid should not condense.
Figure 2. P-H diagram for HFE-7000. The state analysis used to calculate the states of the ORC was created in a spreadsheet and is semi-automated. Certain inputs are defined and the spreadsheet calculates all the states by looking up relevant values in HFE-7000 property tables. With the automated spreadsheet theoretical goals can be quickly achieved by adjusting inputs. For example, to
10 achieve a higher thermal efficiency, it is easy to see precisely what input temperature would be required, or what expander efficiency, or what combination of these and other factors. In the model, the following inputs are required to fully define the system:
State 1 pressure
Thermal sink temperature
Pump head
Pump efficiency
Energy input (defined as thermal reservoir temperature in, temperature out, and flow rate)
Expander efficiency In order to define these inputs and to improve the accuracy of the model, the
following assumptions were made (some of these values will be validated or corrected with the system experiment):
The pressure at State 1 (p1) is ambient – 101kPa.
The thermal sink temperature is ambient (25°C), and the reservoir is large enough to assume it is isothermal throughout the experiment.
The flow rate of the working fluid is 0.38 L/s (6 GPM) – pump head and efficiency are obtained from the performance curve in Figure 20.
Thermal reservoir temperature is 100°C, with a drop in temperature of 14°C at 1 kg/s flow rate – this provides approximately 60kW of power.
Expander efficiency is 50% - this is based on isentropic efficiencies reported for the Copeland ZR108KC-TFD compressor that range from 42.7% to 72.7% [see appendix].
There is a 5% pressure drop through the boiler and the condenser.
There is a 6°C difference in the heat exchangers between the working fluid and the reservoir fluid.
Work done by the pump and work done through the expander are both isentropic. For each state, only two properties are required. The properties used, and the
methods used to obtain them, are described below. State 1 – Condenser to Pump – Pressure is assumed to be atmospheric, and temperature is the same as the assumed condenser temperature plus an offset. (1)
11 (2) State 2 – Pump to Boiler – Since it is isentropic work, the entropy at state 2 is the same as at state 1. Pressure is found by adding head obtained from the pump curve (Figure 20) to pressure in State 1. (3) (4) State 3 – Boiler to Expander – Enthalpy (H) is calculated from the heat addition and fluid flow rate. Pressure is assumed to drop 5% from P2. (5) (6)
State 4 – Expander to Condenser – Expansion is treated as isentropic, so entropy at State 4 is the same as that at State 3. Pressure is assumed to drop to 5% greater than the pressure after the condenser (P1). (7) (8) Efficiency of the expander and pump have not yet entered into the calculations, as they do not have a direct bearing on the working fluid. Rather, the overall system efficiency depends on the efficiency of both components: (9)
, where , and .
(10) (11)
Where ηe and ηp represent the efficiency of the expander and pump, respectively. State calculations using the above techniques, assumptions, and component characteristics, together with data provided by 3M can be found inTable 1.
12 Table 1. Theoretical State Calculations State 1 - After Heat Rejection kPa TH2O K mdotH2O kJ/kg J/molK State 2 - After Pump kPa
p1 T1 h1 s1
103 304 67.735 52
Head Efficiency p2 h2 T2
300 0.2 403 kPa 67.83 kJ/kg 304 K State 3 - After Heat Addition 382.85 kPa H2OT1 269.7 kJ/kg mdotH20 368 K T2 170.24 J/molK Qin State 4 - After Turbine 108.42 kPa 0.5 250.0 kJ/kg 342 K
p3 h3 T3 s3 p4 Efficiency h4 T4
Power Efficiency mdot
298 K 1.00 kg/s
374 1 360 58.52
K kg/s K kW
Analysis 2.863 kW 4.67% 0.29 kg/s
Values in the table that are bold and highlighted yellow represent inputs; all other cells represent calculated values. The model predicts that the system will operate at a thermal efficiency of nearly 5% while producing 2.9kW of electric power. This analysis is intended to lend confidence to the system testing as well as to provide theoretical goals for each state. Knowing the ideal fluid states at various points in the system should take the guesswork out of the testing. These points can be validated using thermocouples and pressure gauges mounted on the system. This data will then be incorporated into the model to increase its fidelity. The pressure and enthalpy of theoretical system states overlain on the saturation curve of HFE-7000 can be seen in Figure 3.
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Figure 3. P-H diagram of theoretical ORC states.
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CHAPTER 3 SYSTEM DEVELOPMENT The vast majority of work that was done in this project went into modifying and troubleshooting the existing Rankine Cycle system to bring it to an operable condition. Damage to the system and components that were not appropriate to the fluid under investigation had to be overcome in order to perform an actual test. The system at the outset of the project can be seen in Figure 4.
Figure 4. ORC system at outset of project. Components that made up the system are as follows:
3 Flatplate heat exchangers
Condenser – Flatplate model C35
Boiler – Flatplate model CH50
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Pre-heater – Flatplate model FP5X12-100
Centrifugal pump – Magnatex MPT201
Air motor functioning as expander – N/A
Electrical generator – N/A
Oil separator – N/A
Oiler – N/A
Reservoir tank – One gallon pressure tank
Various valves, pressure gauges, and thermocouple wells
2 charge ports
HEAT SOURCE Finding an appropriate heat source to provide energy to the ORC was critical to the project. Any adequate heat sources must be near 100°C, measurable, and safe. In addition, it should be controllable and large enough to power the system for several minutes. Several options were investigated: steam, engine waste heat, and water.
Steam Steam was an obvious candidate initially as it could be easily regulated, it was readily available in the building heating system, and it supplied the temperatures required by the ORC. However, it became apparent that the amount of steam required to run the system at the desired power level was more than could be spared. Steam also presented a potential safety hazard. For these reasons steam was abandoned as a potential energy source.
Internal Combustion Engine The majority of the experiments performed on the ORC utilized the waste heat from an internal combustion engine as an energy source. The primary benefit of this heat source was that unlike the steam, it offered total control over when heat could be provided. Drawbacks to the automobile engine heat source included excessive noise and exhaust fumes, but the primary limitation was the lack of control over the quantity and quality of heat that the engine was capable of producing. Initially the waste heat was harvested from the exhaust by diffusing it through an airto-liquid heat exchanger (in this case, an automobile radiator). However, this method did not
16 provide enough energy to run the system at the desired levels. Ultimately, the heating fluid line was linked directly with the cooling system of the engine with an in-line pump. In this way, energy was removed from the engine block as the cylinders were firing and pumped directly into the heat exchanger to heat and boil the working fluid. This setup was successful in that it managed to raise the heating fluid to above 180°F, though not any higher. However, there were several critical shortcomings that made the heat source unacceptable. First, the flow rate was unknown – because of the thermostat in the engine, the coolant pump would turn on and off depending on the amount of energy removed by the ORC. Although the in-line pump was continually circulating, the flow rate would vary depending on the state of the engine’s pump. Also, the engine would frequently overheat, causing the pressure relief valve to open on the radiator, adding further variation to the already erratic flow rate. Second, temperature was unknown – the thermometer on the engine would fluctuate, requiring temperature measurement at the inlet of the boiler heat exchanger, which is an inaccurate one. Lastly, energy input was unknown – the engine was manually throttled up and down to provide more or less energy to the ORC, but there was no way to quantify the energy. Because the use of the engine as an energy source made the accurate gathering of data impossible, it was abandoned in favor of a more controlled solution.
Water Heater Source In the end a water heater was chosen as an energy source for the ORC. The water heater offers the ability to easily control temperature, monitor the amount of energy entering the system, and maintain a consistent flow rate. A water heater had been previously considered but it was thought that safety features would prevent it from achieving the temperatures that the system required. In fact, the temperature is only limited by the boiling point of the fluid, which is water in this case. The water heater selected has a 150 L (40 gallon) tank and two electric heating elements that each draw approximately 3600 Watts. The primary shortcoming of the water heater is its limited energy capacity – if the system produces 2 kW at 5% thermal efficiency, it would require 40kW of power input from the water heater. Rejecting energy at this rate, even with the heating elements powered, the 150 liters of water would lose about 3°C/minute, which would reduce the water to a temperature
17 below what is required to boil the HFE-7000 at the operating pressure within 6 minutes. Further, because of inefficiencies in the heat exchanger, the water must be several degrees warmer than the working fluid. Without multiple tanks, experiments on the ORC can only last 2 or 3 minutes before the heat source is exhausted. Energy is transferred from the water tank to the heat exchanger using a hightemperature circulator pump. After a 1/12 HP pump failed to achieve adequate flow, a 1/8 HP pump was installed. The larger pump is capable of generating a flow rate of approximately 1kg/s in the ORC system. All tubing between the tank and heat exchanger is insulated. The water tank with pump and insulated tubing is shown in Figure 5.
Figure 5. Heat source installed.
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HEAT REJECTION SINK The theoretical analysis performed above shows that the ORC achieves an overall thermal efficiency of less than 5%, which means that 95% of the energy put into the system (minus minimal losses) needs to be removed at the condenser. The simplest approach to remove this amount of energy is with a large thermal sink. In this case, a plastic tub was filled with 340 L (90 gallons) of water, and a ¼ horsepower sump pump was utilized to circulate the water through the condenser heat exchanger (see Figure 6). This approach works well at first because the temperature of the water is low enough to condense the HFE7000. However, after the rejected heat warms the thermal sink, the working fluid is condensed less efficiently. In the previous example of a 2kW system with 5% efficiency, the tub warms at a rate of approximately 1.6°C/minute. In order to maximize system efficiency, the sink must be maintained at a low temperature. This can be accomplished with ice, with frequent replacement of the water, or by leaving long intervals between tests for the bath to cool to the ambient temperature.
Figure 6. Heat sink installed.
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SYSTEM COMPONENTS The components that came with the original system were for the most part well-sized for use with HFE-7000. Specifically the pump and the air motor were capable of delivering the amount of power desired from the modified system. The heat exchangers were sufficiently large to transfer enough energy to and from the working fluid to operate the system at the desired power levels. The original system also included a large oil separator immediately downstream of, and directly above, the expander. The majority of the time spent on the project left the system as-is, simply connecting heating- and cooling-fluid circuits, charging it with HFE-7000, and measuring power output. However, due to several factors, no output was ever achieved. First, the lack of an adequate heat source made delivery of sufficient energy to the system impossible. Second, the expander had a substantial leak that was either caused or exacerbated by exposure to HFE-7000. Lastly, the oil expander acted as a large thermal sink, causing the fluid to condense and fall back into the expander, inhibiting its function. The last two factors necessitated a replacement or repair of the expander and a reconfiguration of the system.
Heat Exchangers The heat exchangers are comprised of an evaporator rated to 220 kW (50 tons) at 7.6 L/s (120 gpm) and a condenser rated to 154 kW (35 tons) at 6.6 L/s (105 gpm). The actual flow rate through the boiler and condenser is approximately 1 L/s, yielding a capacity of 37kW and 29kW, respectively. The heat exchangers are well-sized for this system which absorbs 30kW and rejects approximately 29kW. All heat exchanger types are flat plate type which require less volume for equivalent capacities.
Pump Based on its pump performance curve (Figure 23, p. 56), the Magnatex MPT201 is able to provide the appropriate amount of head at the desired flow rates. Further, the pump housing is magnetically coupled, which requires no seals around the shaft and reduces the possibility of leaks.
20
Expander A variety of expander types are available for this particular application, most notably the multi-vane expander, screw expander, scroll expander, and micro-turbine. Each has advantages and drawbacks that must be considered before any one is selected.
MULTI-VANE EXPANDER The multi-vane expander (MVE) is a positive displacement expansion device invented by Charles C. Barnes in 1874, originally intended as a compressor. It utilizes a rotor turning eccentrically inside a cavity. Vanes on the rotor maintain contact with the cavity walls (see Figure 7). Advantages to the MVE include flat operating efficiency curves over a wide range of conditions, low speeds (approximately 3000rpm) that can match generator speeds without a gearbox, the ability to operate in the presence of liquids and wet vapors, minimal maintenance with little lubrication requirement, and proven operation with organic working fluids (O’Callaghan et al., 1985).
Figure 7. Multi-vane expander. The authors performed an extensive analysis of two different varieties of MVE (circular and non-circular), taking into account the major modes of losses such as breathing, internal leakage, and friction. The non-circular MVE was a converted commercially available refrigeration compressor. The circular MVE was a prototype designed by the authors in which the stator cylinder comprised two circular arcs, one acting as a sealing arc.
21 The theoretical simulation the authors performed was compared to experimental data gathered using a low-grade energy ORC with R-113 as a working fluid. The fluid was mixed with 5% oil by weight to keep the expander lubricated during operation. A number of factors were modeled and measured, including flow rate, rotational speed, cell pressure, torque, mechanical efficiency, power output, and isentropic efficiency. The theoretical model often aligned with experimental results with errors of less than 10%. Testing was limited by power available at the boiler – 85kW at approximately 125°C. While mechanical efficiency was measured at over 80%, the isentropic efficiency of either type of MVE never exceeded 60%. This value corresponded to the peak power point and occurred at a rotational speed of 3200 RPM. Work has continued over the years with the investigation and implementation of MVEs as prime movers in low-temperature ORCs due to their wide availability, reliable performance over a range of conditions, and compatibility with a variety of fluid types. A multi-vane expander would be a reasonable choice for use in the present research.
SCREW EXPANDER The helical screw expander is likewise a positive displacement device. It is comprised of two meshing, helical, counter-rotating screws (see Figure 8). As the fluid expands and moves along the screws, it drives (rotates) them apart to increase volume. This work is geared to a single shaft which can be used to power a generator. Since the 1970s two-phase screw expanders have been investigated and utilized for the recovery of power from liquid or low-dryness geothermal brines by expanding steam or even pressurized hot water. Smith et al. (1999) worked to use the two-phase expansion process with organic working fluids for which exit volume flow rates and expansion ratios are significantly reduces for similar heat utilization at certain temperatures. The authors achieved an expander efficiency of greater than 75% in an organic Rankine cycle using R-113 as a working fluid. Fundamental improvements to the expander design were required to achieve this high efficiency. Like all other positive displacement devices, the seal is critical to prevent internal leakage. In order to prevent direct contact but also achieve a seal between the lobes of each rotor in a screw expander, two methods of lubrication are device types have been developed:
22
Figure 8. Screw expansion device (oil injected). Source: Smith, I. Stosic, N., & Kovacevic, A. (1999). Power recovery from low cost two-phase expanders. Centre for Positive Displacement Compressor Technology, City University, London. oil injected and oil free. The oil injected machine, which relies on oil carried by the compressed gas in order to lubricate the rotor motion, seals the gaps and reduce the temperature rise during compression without requiring internal seals. The oil injected type of machine is simple in mechanical design, cheap to manufacture, highly efficient, and widely used as a compressor. The oil free machine separates any oil from the working fluid by preventing contact between the rotors with lubricated meshing timing gears outside the working chamber. Internal seals are required against the bearings and chamber walls. These additional parts and requirements cause the oil free machine to be considerably more expensive than its oil injected counterpart. Smith et al. found that the amount of oil required for the oil injected machine could not be carried by the fluid vapor, but that the standard oil free machine was cost prohibitive and so were forced to develop an alternative. Using a new lobe profile that nearly eliminated any sliding motion between rotors in favor of rolling motion, and reduced contact stress, significantly reduced the need for oil to the point where a small amount of injected oil could be used. Further, this new design did not require a timing gear as a typical oil free design would. These improvements resulted in
23 a relatively inexpensive expander device that was compatible with organic working fluids that can carry only a small amount of oil, yet highly efficient. Screw expanders depend on precise numerically-controlled machining to achieve a leak-resistance fit, especially with oil free types. Because of the tight sealing requirements, they tend to work better with wet fluids. Because the authors worked with a two-phase system, their results that show high efficiency will not likely transfer to the dry system under investigation in this paper. A dry fluid system would still require seals that greatly increase the cost of the machine.
SCROLL EXPANDER The scroll expander is the last type of positive displacement device that was considered in this study. The device is comprised of a stator and a rotor scroll (resembling spirals). The rotor orbits the stator eccentrically without rotating, trapping pockets of fluid against the stator. As the rotator processes against the stator, this volume of trapped fluid expands (see Figure 9). In reality, the fluid itself is under pressure, so its expansion drives the rotation of the rotor, creating useful work. The scrolls and shaft are designed to minimize the gap between rotor and stator as this is a source of leakage.
Figure 9. Scroll expander operation. The scroll expander is an inexpensive device due to its widespread use and lowcomplexity. Scroll compressors, which operate on the same principle but in reverse, are commonly used in both automotive and building air conditioning systems. Several recent studies have been performed characterizing myriad aspects of the scroll expander.
24 Lemort et al. (2009) present the results of an experimental study carried out on a prototype of an open-drive oil-free scroll expander integrated into an ORC working with refrigerant HCFC-123. The authors identify eight parameters of a scroll expander for semiempirical modeling. Then, relevant parameters are such as the mass flow rate, the delivered shaft power and the discharge temperature are determined. Also, secondary variables such as the supply heating-up, the exhaust cooling-down, the ambient losses, the internal leakage and the mechanical losses were obtained. The maximum deviation between the predictions by the model and the measurements was 2% for the mass flow rate, 5% for the shaft power and 3 K for the discharge temperature. The validated model of the expander was used to quantify the various losses and to introduce possible design alterations to achieve better performances. It was concluded that internal leakages and, to a lesser extent, the supply pressure drop and the mechanical losses are the main losses affecting the performance of the expander. Lemort and Quoilin (2009) have recently evaluated the efficiency of scroll expanders in organic and inorganic Rankine cycles with promising results. Between two studies, over 40 steady state performance points were gathered for R123 and water evaluating the effects of boiler temperature, mass flow rate, condenser temperature, and expander rotational speed on shaft power, the expander isentropic effectiveness, and the cycle efficiency. Expander efficiencies ranging from 42% to 68% were observed yielding system efficiency up to 7.4%. Loss analysis of scroll expanders using R134a refrigerant was investigated by Shigemi, Toshio, and Akira (2009). A method was proposed to estimate losses using P-V diagram. The authors suggest that leakage in suction process was large, so tip seals of scroll clearance are very important to obtain high efficiency. A new suction port system was designed which decreased the leak by about 10%, and in turn increased expander efficiency by about 10%. Other relevant results were also obtained by the authors, such as the effect on efficiency of pressure ratio, mass flow rate difference between measured and theoretical values versus the increase in the rotational speed and the decrease in the pressure ratio of expander, and maximum shaft power and expander efficiency. An expander efficiency of 74% was obtained at the pressure ratio of 3.7 and rotational speed of 50 Hz. Mathias, Johnston, Cao, Priedeman, and Christensen (2009) tested application of gerotor and scroll expanders to determine their applicability in producing power from lowgrade energy using R123. Their experimental study showed that both types of expanders
25 were good candidates to be used in an ORC. The experiment produced 2.07 kW and 2.96 kW, and had isentropic efficiencies of 0.85 and 0.83, respectively. The authors also presented results incorporating some strategies to improve cycle efficiency. One such change pertains to matching the flow rate of the working fluid with the inlet pocket volume and rotational speed of the expander. The volumetric expansion ratio of the expander was also matched to the specific volume ratio of the working fluid across the expander. The network was calculated after subtracting the power required by the pump and the condenser fan – which gave 6271 W of net power at an overall energy efficiency of 7.7%. Scroll devices are readily available and inexpensive, partly due to their widespread use as air conditioning compressors. Scrolls customized for use as expanders are also commercially available. A 3kW oil-free scroll expander from Air Squared costs $3,300 and can achieve an isentropic efficiency of over 80% (http://www.airsquared.com).
TURBO EXPANDER The micro turbine, or turbo-expander, is the only non-positive displacement device considered. It operates in a similar manner to the expansion portion of a gas turbine engine (combustion and compression removed). A high pressure gas is directed past turbine blades causing them to rotate as the gas expands. This rotational work can be harvested for electrical power. The rotor can be an axial or centrifugal type, with the centrifugal being more common because of its frequent use in automotive turbochargers. The turbo-expander is standard for large-scale Rankine cycles due to its high efficiency. In small-scale applications, drawbacks to the turbo-expander include high cost, high shaft-speed that must be geared down to work with an electric generator, and a low tolerance for varying gas inlet pressures. One such experiment in which a turbo expander was used in a low-power application was performed by Teagan and Clay (1973). Their research investigates the performance of a 3 kW expander and generator set driven by an organic Rankine cycle. The turbine speed of 70,200 rpm was geared down to 3,600 RPM. This gearing was done internally to the sealed system so that a low speed shaft seal with proven performance could be used to transmit the turbine output to the generator. Fluorinol-85 was used as the working fluid. The conditions at the inlet to the turbine were 530F and 480 psi. At the exit of the turbine conditions were 202F and 29 psi.
26 The operating pressure and temperature of the system investigated above are both higher than can be achieved in this study. No documented uses of the turbo expander at pressures and power levels more appropriate for this study have been found. Because of this poor heritage, and because of its high cost, the turbine expander is a decidedly poor choice for the present research.
EXPANDER SELECTION After careful consideration of available expansion devices, a scroll-type device was chosen as a replacement for the air motor. The scroll device has several advantages: few moving parts, wide availability, relatively low cost, and proven performance with several common working fluids. Significantly, extensive research has been performed with scroll devices in similar systems. These studies form a strong basis for continued work with the scroll expansion device, lending confidence to the choice and helping to avoid pitfalls as the device is optimized for this particular system. A qualitative comparison of the different expander types can be seen in Table 2. The quantity, applicability, and recency of research focusing on each expander type, as well as any commercial use, define the “heritage” factor. Table 2. Expander Comparison Summary Type
Performance
Cost
Heritage
Rotary Vane Screw Scroll
+ ‐ ++
$ $$ $
+ + ++
Turbine
++
$$$
‐‐
The scroll device chosen for this project is actually a compressor designed for air conditioning systems – model ZR108KC-TFD made by Copeland (see Figure 10) [specifications in appendix]. Conversion of the Copeland compressor, though not ideal because it is not optimized for this use, has been successfully accomplished before (Ingley, Reed, & Goswami, 2004; Kim et al., 2007). As a compressor, this device is rated at 9 horsepower, or 6.7 kW, which is more than adequate for the amount of power expected from the ORC. It was important to allow flexibility in the system in case more or less power was
27
Figure 10. Expander schematic. desired; the expander selection was intentionally oversized with the knowledge that lower power operation would cause a reduction in efficiency.
COMPRESSOR CONVERSION Since the expander is actually a compressor operating in reverse, substantial modifications were required in order to ensure both its operation as an expansion device and its integration in the ORC system. Before any modification began, the device was tested
28 under normal operation (with 3-phase power supplied) to verify its functionality. Then compressed air was injected into the outlet (which became the inlet of the expander) to test its operation in reverse. It was quickly verified that device check valve was preventing reverse flow, and the case had to be opened (see Figure 11). The scroll device is shown in Figure 12.
Figure 11. Expander with top removed and check valve assembly.
Figure 12. Scroll device sides A and B.
29 After removal of the check valve, an injection of compressed air directly into the scroll device verified its operation in reverse. However another function of the assembly was to route fluid from the scroll outlet to the case outlet. The removal of the check valve allowed gas to move freely from the inlet to the exit of the case, completely bypassing the scroll device. In the compressor, fluid enters at the inlet then is free to diffuse throughout the case before it is compressed through the scroll and routed through the check valve and toward the outlet at the top of the case (see Figure 13). It would have been possible to modify the check valve assembly to allow reverse flow and still maintain proper fluid routing, but three factors made this an undesirable solution: head loss through the modified valve would decrease system efficiency, the tolerances that were required to achieve a seal when the cap was reattached to the case were extremely tight, and using the port lower on the case made system integration difficult.
Figure 13. Compressor flow diagram.
30 Instead of keeping the existing fluid routing, it was decided to route the incoming fluid directly through the top of the cap into the scroll, and then allow the fluid to diffuse through the case after expansion and exit through the original exit port (see Figure 14).
Figure 14. Expander flow diagram. The modification to the fluid routing through the expander required widening the inlet of the scroll device (see Figure 15) in order to accept the 1.5” pipe diameter used elsewhere in the system. The inlet was widened further to accommodate o-rings used for a seal. The previous inlet port was capped since it was no longer needed. The previous outlet port was widened to accommodate 1.5” tubing. The new inlet and exit tubing was welded to the cap. A flange was also welded around both the cap and the case in order to facilitate removal of the cap during troubleshooting. For the same reason, unions were installed on either side of the expander to allow easy removal (see Figure 16 for cap modifications).
31
Figure 15. Scroll device installed with inlet machined.
Figure 16. Cap before and after modifications.
PLUMBING CONFIGURATION The removal of the oil expander opened the window for additional modifications to the system plumbing. First, the expander was moved to the highest point possible in the system in order to eliminate the risk of liquid HFE-7000 falling back into the expander as it did in the original configuration. Figure 17 shows the expander installed in the system.
32
Figure 17. Expander installed in ORC system. Second, extraneous piping was removed in order to minimize parasitic thermal losses. This involved lowering the condenser which required removal of one charge port, and removing stress-relief tubing. The stress relief tubing is a potentially beneficial feature for a system undergoing deep thermal cycles, however it was decided that the low operating temperature of the system made stress-relief unnecessary. Lastly, the line between the pump and the boiler was straightened (see Figure 18), which reduced head loss through this highpressure line and also allowed a charge/drain port to be installed at the bottom of the system. Positioning the drain port at the lowest point is when HFE-7000 is used because it is a liquid at room temperature and cannot be removed like traditional working fluids.
33
Pre heater
From pump
Figure 18. Pump to pre-heater link before and after modification.
ELECTRICAL LOAD The air conditioning compressor that was converted to an expander uses a 3-phase induction motor to power the scroll. This complicated the conversion of the compressor because induction motors utilize current in the stator to induce a magnetic field in the rotor which is used to generate torque. When an induction motor is used as a generator, there is no current provided from an external source, and so no magnetic field can be induced in the rotor. However, if capacitors are wired between stator leads and a magnetic field can be induced somehow, even for a short period, the generator will become self-sustaining. There are two ways to apply a current to the stator – an external power source, such as a car battery, can be used to “jump start” the coils, or the tiny amount of residual magnetism that almost always remains in the copper windings can be exploited to induce small currents which in turn induce more magnetism until the coils build up enough charge to operate at maximum voltage. When using the latter method, which was chosen for this project, it is critical that no load be applied while charge is building, since virtually any load would be too large for these tiny seed currents to be produced. Motor run capacitors of 35 microfarad rated at 370VAC were selected (see Figure 19). These are more capacitive than is strictly required, but it is safer to err on the high side as larger capacitors will not inhibit the charging in any way, but undersized capacitors could short.
34
Figure 19. Run capacitor wiring. Various loads were considered that are capable of dissipating multiple kilowatts of power. Ideally, the generator would be wired to the hot water heater to dissipate any power produced directly back into the heat source. This was tried without success – unbalanced loads disable 3-phase generators and only 2 heating elements were available in the water heater. If a 3rd element were available, a resistive load like this would have ample capacity to dissipate whatever power the generator produces. Also, resistive loads have no reactive element, so all the power measured is real power with no imaginary component that must be factored into calculations later. Ultimately a 3-phase motor was selected as the load; although the reactive element of an electric motor requires some calculation to find the total effective power, a motor offers the benefit of guaranteed functionality with a 3-phase source. Switches were also wired along 2 of the lines to allow manual application of the load once sufficient charge is achieve in the induction generator. No third switch is required because a
35 single line does not complete any circuit through the motor and thus cannot draw any power from the generator. The electric connection between the generator and the load is shown in Figures 20 and 21.
Figure 20. Electrical setup diagram.
Figure 21. 3-phase load with clamp-on ammeter and switches.
36
MEASUREMENT The ORC is equipped with thermocouple wells at critical locations throughout the system, allowing measurement of the working fluid temperature. The system also has pressure gauges mounted in critical locations, but during system charging it was discovered that the offset and/or accuracy of these gauges are poor enough to add considerable uncertainty to system measurements. Though correction factors were added during this impromptu calibration, data from the pressure gauges are suspect and treated as primarily qualitative. Measurement of the heating and cooling fluids is critical for calculating the total energy added to and removed from the system. Measuring temperatures at the inlet and exit of the heat exchangers was done by using a thermocouple and a heat gun to measure the temperature of the fittings on the heat exchanger. This method proved to be inaccurate and was abandoned in favor of a time-averaged bulk reservoir temperature measurement. During system testing, temperatures of the thermal reservoir and sink were recorded at fixed time intervals. Data were also taken in a similar way without the system running in order to account for error caused by thermal losses from the reservoirs. It turned out that both the reservoir and sink lost a negligible amount of energy over the time frame of the experiment – due to effective insulation and relatively large thermal mass. Measurement of the electrical output is done using a clamp-on ammeter around one of the three wires powering the 3-phase motor and a voltmeter probing across one of the capacitors.
WORKING FLUID Many factors must be considered in the selection of the working fluid. Besides thermodynamic performance, these also include environmental impact, safety, cost, and availability.
PERFORMANCE The performance of refrigerants in Organic Rankine cycles has been exhaustively researched. One common characteristic considered in fluid selection is the slope of the
37 saturation vapor curve. Because of the nature of the low-quality heat source, the superheated approach employed by a traditional Rankine cycle cannot be utilized; the fluid will be close to its saturated vapor curve as it enters the expander. This means that wet fluids will condense as they lose enthalpy through the expander, an undesirable characteristic. Conversely, dry fluids become superheated as pressure drops through the expander. This superheated gas can be used as a regenerator to pre-heat condensed fluid. Hung et al. recently found that a wet working fluid with a steep curve outperformed dry fluids (Hung et al., 1997), despite significant work that has been done to investigate dry fluids (Hung, 2001; Hung et al., 1997; Mago et al., 2008). However no regenerator was implemented in the work done by Hung. Isentropic fluids, those with curves that are nearly vertical, seem to perform best in a non-regenerator system (Quoilin, 2007). Another factor that affects performance of a fluid is its heat of vaporization, or latent heat. Denser fluids with greater latent heat can move more energy through the system with less mass flow. This allows a reduction in the size of all components and an increase in thermal efficiency due to lower losses through large heat exchangers and a larger pump. Lastly, boiling point of the fluid is a critical consideration since the intended use is low-grade heat recovery. The fluid must be able to undergo a phase change at a reasonable pressure (approximately 6 atmospheres) at the intended operating temperatures (below 100°C ). This particular criterion narrowed down the available refrigerants significantly.
ENVIRONMENTAL IMPACT Environmental impact of a working fluid is primarily characterized by three parameters: ozone depletion potential (ODP), global warming potential (GWP), and atmospheric lifetime (ALT). ODP is a measure of the relative capacity for ozone depletion compared to trichlorofluoromethane (CFC-11), which has been assigned an ODP value of 1.0. Traditional refrigerants including CFCs and HCFCs have ODPs between .1 and 1. Many modern refrigerants, such as hydrofluorocarbons (HFC) and hydrofluoroethers (HFEs) have eliminated chlorine from their chemical structure due to the high reactivity of chlorine with ozone, and so have ODPs of zero. GWP is a factor that uses a similar method to estimates how much the chemical will contribute to global warming over a fixed time period. In the case of GWP, the reference chemical which has been given a GWP value of
38 1.0 is carbon dioxide. The factor is based on the total amount of infrared absorption, the particular wavelengths of absorption, and the atmospheric lifetime of the chemical. location of radiative factor, and decay rate of a chemical (Environmental Protection Agency [EPA], 2010). The ALT of a greenhouse gas refers to the approximate amount of time it would take for human-caused contribution to an atmospheric pollutant concentration to return to its natural level as a result of either being converted to another chemical compound or being taken out of the atmosphere via a sink. This time depends on the pollutant's sources and sinks as well as its reactivity. The lifetime of a pollutant is often considered in conjunction with the mixing of pollutants in the atmosphere; a long lifetime will allow the pollutant to mix throughout the atmosphere. Average lifetimes can vary from about a week (sulfate aerosols) to more than a century (chlorofluorocarbons, carbon dioxide) (EPA, 2010). For nearly 100 years CFCs have been used as refrigerants because of their desirable performance and safety characteristics. Due to their high ODP they are being temporarily phased out by HCFCs which have a lower ODP. However, even HCFC contribute to ozone depletion to some extend and are being phased out as well (http://www.arap.org). The Montreal Protocol of 1987 eliminates CFC production by 2010 (except for essential uses) and HCFC production by 2030 and 2040 in developed and developing countries, respectively. Alternatives to HCFCs include HFCs and HFEs. HFCs provide many of the required performance characteristics of CFCs and HCFCs and have zero ODP. As a result, they are frequently used as a drop-in replacement fluid. However HFCs have a significant GWP so have been identified by the Kyoto Protocol as a greenhouse gas and thus targeted for reduction (http://www.arap.org). HFEs show promise as an ultimate solution both to GWP and ODP concerns. HFEs have zero ODP as well as very low GWP and ALT compared to HFCs. These properties make HFEs desirable for use as a working fluid in energy conversion systems, but very little investigation has been previously performed into the use of HFE fluids in such an application. A comparison of the three factors discussed for several commonly used refrigerants, and the fluid under investigation, is shown in Table 3.
39 Table 3. Environmental Factors for Possible Working Fluids CFC‐11 HCFC‐22 HCFC‐123 HFC‐134a HFE‐7000
ALT 45 12 1.3 14 4.9
ODP 1 0.034 0.012 0 0
GWP 3660 1710 53 1320 370
SAFETY The American Society for Heating, Refrigeration and Air Conditioning Engineers (ASHRAE) defines the level of toxicity and flammability of a refrigerant as one of 2 and 3 classes, respectively. Toxicity can either be class A if the refrigerant has not been identified as toxic at concentrations less than or equal to 400 ppm by volume, or class B if it has been. The flammability of a refrigerant can be class 1 if there is no flame propagation in air at 21°C and 1 atm, class 2 if its lower flammability limit (LFL) is greater than 0.10 kg/m3 at 21°C and 1 atm and its heat of combustion is less than 19,000 kJ/kg, or class 3 if it is highly flammable as defined by LFL less than or equal to 0.10 kg/m3 at 21°C and 1 atm or its heat of combustion is greater than or equal to 19,000 kJ/kg. These two classes are written together as a two-character figure (for example: A2 or B3). The ASHRAE safety figure does not factor into the performance of the ORC, but is nonetheless important to consider in the selection of a working fluid as it affects operator and facility safety.
FLUID SELECTION Due to largely environmental considerations, this study narrowed the list of possible refrigerant candidates to HFEs. Within that category, boiling point was a discriminating requirement that further narrowed the field. After consideration of all the relevant factors that have been identified, HFE-7000 was selected as the working fluid for this project. HFE-7000 is a hydrofluoroether manufactured by 3M (under the trade name Novec 7000). It is a segregated HFE, meaning it has no hydrogen atoms within the molecule bonded directly to fluorine, but instead has all
40 hydrogen-carbon bonds segregated from fluorine by ether oxygen, reducing its atmospheric lifetime (Owens, 1998). It has a GWP of 370, zero ODP, and an ALT of 4 years. It is a safe fluid with low toxicity and zero flammability (ASHRAE A1) (3M Electronics, 2005). From a performance standpoint, it is a dry fluid of high-density, high latent heat, and a breakdown temperature of 186°C, making it a potentially good replacement for current HCFC and HFC fluids. Other benefits include a high heat of vaporization allowing system components with lower pressure capability, and compatibility with a wide range of materials including brass, copper, and a wide variety of plastics. Husband and Beyene (2008) have demonstrated the feasibility of a low-grade heat recovery system that can produce electrical power using HFE7000 as the working fluid. A theoretical efficiency of 11% for a 10kW system was shown at a pressure ratio of 6 and input temperature of 240°F.
41
CHAPTER 4 RESULTS AND DISCUSSION EXPERIMENTAL DATA The experiment performed with the ORC system successfully produced electric power. The system experiment began with the water heater temperature at boiling and the condenser tank at 80 deg F. Circulator pumps for both reservoirs were on to ensure the heat exchangers were in equilibrium with the fluid and would not cause any transient effects. The 3-phase load was electrically detached from the generator using the two switches (see Figure 18). A timer was started when the compressor was turned on and valves were opened. Almost immediately, the expander began spinning (can be felt and heard). Within seconds, voltage across the capacitors built from less than 1V to nearly 300V. Once maximum voltage was reached and the rotor was properly magnetized, the two switches were closed to connect the load. Temperature data for the two water tanks and electrical data between the generator and motor were taken at 30 second intervals. The system produced 937 W electrical power using an input of 30 kW. This corresponds to a gross thermal efficiency of 3.1%. Results are summarized in Table 4.
Table 4. Summary of Experimental Data Qin Volume
Mass ΔT Δt ΔE P Δ
Qout 20700 in³ 40 89.6 gallons 151.2 338.7 liters 151200 338727 grams ‐12.5 6.5 °F
180 180 s 4390365 5114490 J 30444 27414 W 3030 W
Wout Current Voltage Power
2.6 A 208 V 937 W
ηth
3.1%
ηe
31%
42 The power measured at the output of 2.6A at 208V corresponds to the “real” component of a single line. In order to calculate total power, “reactive” power must be included as well. The theory behind the calculation is not within the scope of this paper, but total power can be derived from real power by multiplying by the square root of 3 (Manhire, 2004). It can be seen from the table that the amount of energy lost between the boiler and the condenser (3030 W) is significantly higher than the amount of energy produced as electricity (937 W). The discrepancy between these numbers primarily represents the efficiency of the expander according to the following equation: (12)
This equation assumes that other losses, such as thermal losses from the pipes and the expander, are negligible; this is a safe assumption because the system is well insulated. From the equation the efficiency of the expander is found to be 31%.
UPDATE TO THEORETICAL MODEL The theoretical state model was updated with data gathered from the system experiment. Data updated include condenser temperature, expander efficiency, and energy addition to the boiler. Using energy addition to the boiler along with the enthalpy at states 2 and 3 allows the calculation of the mass flow rate within the system according to the following equation: (13)
The flow rate is calculated to be 0.15 kg/s. This can then be used, along with the temperature at the pump (T1), to find the efficiency and head of the pump, which is also updated in the model. The new state calculations can be found in Table 5. It should be noted that a mass flow rate of 0.15 kg/s corresponds to a volumetric flow rate of slightly over 400 cubic feet per hour (CFH) at the inlet. Also included in Table 5 is a calculation of the work done by the pump, and the amount of electrical power (Win gross) required to provide this energy at the rated
43 Table 5. Experimental State Calculations p1 T1 h1 s1
103 304 67.735 52
Head Efficiency p2 h2 T2
520 0.07 623 67.89 304
State 1 - After Heat Rejection kPa TH2O K mdotH2O kJ/kg J/molK State 2 - After Pump kPa Win(net) Win(gross) kPa kJ/kg K State 3 - After Heat Addition kPa H2OT1 kJ/kg mdotH20 K T2 J/molK Qin State 4 - After Turbine kPa
p3 h3 T3 s3
591.85 267.35 370 166.14
p4 Efficiency h4 T4
108.42 0.31 242.4 kJ/kg 334 K
Power Efficiency mdot
302 K 1.00 kg/s
0.02 kW 0.33 kW
372 1 365 30.43
K kg/s K kW
Analysis 1.180 kW 2.80% 0.15 kg/s
efficiency. The “efficiency” value near the bottom of the table is thermal efficiency, for easy comparison to data reported in Table 4. A chart of the states overlaid on a pressure-enthalpy (P-H) diagram of HFE-7000 can be seen in Figure 22.
DATA COMPARISON The data between experiment and theory agree fairly well, with the model predicting higher power output and lower efficiency than was actually measured. In a well-insulated system with well-characterized components, the model should accurately predict real-world system performance. In this case, the power output matches within 20% and the efficiency matches within 13%. This comparison serves to validate the state calculations and the fluid property tables, but does recommend fine-tuning of the model for this particular system.
44
Figure 22. P-H diagram of actual ORC states. More accurate system data would be required to determine precisely which assumptions in the model lead to a discrepancy in the output characteristics. However, for a system for which accurate component properties are known (such as head loss through heat exchangers), future models can be optimized before any hardware is built.
DISCUSSION A 3% thermal efficiency for this system is reasonable to expect; the Carnot cycle efficiency at the experimental temperatures is only 18.8%. However, significant improvements should be possible with further optimization of key components. A goal of one-half the Carnot efficiency – 9.4% – is achievable with HFE-7000. An increase in thermal efficiency can be achieved by addressing several significant issues that arose after analyzing measured data from the system test and comparing the data to the model. Those issues relate to the pump, the expander, the heat source, and the heat sink.
Pump Performance First, the pump was found to be oversized for the flow rate achieved (which is largely dependent on the capacity of the heat source, to be discussed below). The pump efficiency
45 was found using the manufacturer performance curves (see Figure 23) to be 7% – an order of magnitude lower than what should be expected for an application of this size. Even with a greater flow rate, the pump can only achieve an efficiency of 20%.
Figure 23. Pump performance curve. In the model, the low efficiency results in a work input of 330W (“Win(gross)” in Table 5). This represents over a third of the power generated by the system and so reduces the overall thermal efficiency by that amount. With a highly efficient pump – say, 90% – the impact on thermal efficiency would be less than 1%. In practice, the effect of the pump on system efficiency is even worse. The rated power of the motor driving the pump is almost 1.44 kW – 0.5 kW higher than the power output by the ORC. If the pump power were considered in the net thermal efficiency calculation, the system would have negative efficiency. During operation, the pump feels hot, implying additional friction or work that is not being applied toward pumping the fluid. From both a theoretical and a practical standpoint, it is clear that the pump is not appropriate for the system size and HFE-7000. It is oversized for the flow rate and head required, and it is not designed for, and not necessarily compatible with, a fluid such as HFE-
46 7000 that is more dense and less viscous than water. For this application, a smaller centrifugal pump or a positive displacement pump could achieve an efficiency greater than 75%.
Expander The overall efficiency for the conversion of thermal to electric energy between the boiler and condenser was found to be 31%. Some thermal losses must be expected through the heat exchangers, the plumbing, and the case of the expander itself. However, the system is well insulated, so it is unlikely that the conversion efficiency of the expander/generator is higher than 35%. Considering that the scroll device was able to achieve an efficiency of 72.7% [see appendix] when operating as a compressor, a similar performance should be possible when operating in reverse, so long as similar conditions are applied. Further, a compressor converted for use as an expander has been demonstrated to operate at nearly 75% efficiency (Kim et al., 2007). Two possible culprits that could account for the poor performance of the scroll are pressure and flow rate. An improvement to the performance of the scroll could be achieved with an increase in pressure. However, since the pressure of the HFE-7000 is already its vapor pressure at State 3, the pressure cannot be increased given the limitations of the heat source. That is, the pressure of the fluid cannot rise unless its temperature rises as well, but the temperature is already near the boiling point of water, the maximum available temperature with a hot water heating system. The volumetric flow rate of fluid through the expander is the other factor that strongly influences the efficiency of the scroll. The flow rate of HFE-7000 was found to be 0.15 kg/s, which corresponds to a volumetric rate of 400 CFH at the inlet to the scroll, and 2580 CFH at its exit. Per the datasheet for the expander, rated displacement is 1060 CFH. This rated value probably corresponds to the inlet when the scroll acts as a compressor, or the outlet for the expander configuration. So rather than underpowering the scroll as it was originally thought would happen, it is likely that the fluid is either overexpanding, or much of the expansion is taking place in the expander case and the subsequent plumbing, rather than in the scroll itself. Either of these scenarios would adversely affect the performance of the
47 scroll device because of leakage and friction. Further, the electric motor used to drive the scroll when it acts as a compressor is optimized to operate most efficiently at a certain speed (in this case, 3500RPM). Likewise, when acting as a generator, it operates most efficiently at the same speed. The further the rotor speed varies from 3500RPM, either faster or slower, the less efficiently the generator will operate. Both the scroll device losses and sub-optimal generator performance likely account for the poor efficiency achieved during operation.
Heat Source and Sink It became clear during testing that the thermal reservoir was not massive enough even to adequately power the existing system. The 30kW of power that the system consumed quickly decreased the temperature of the 150 L water heater (a 13°F drop in 3 minutes) even with the heating elements powered. Likewise, the heat sink reservoir heated relatively quickly. As the temperature of both source and sink change, the performance of the system changes. From Figure 22, it can be seen that the HFE-7000 is sub-cooled at State 1. This subcooling decreases efficiency because it removes additional energy from the fluid that must be added again at the boiler. An increase in condenser temperature so that the fluid at State 1 is closer to the saturated liquid curve (37°C at p1) could increase overall system efficiency appreciably. Also from Figure 22 it is seen that the fluid is superheated at State 4. This excess energy is currently wasted by being rejected to the heat sink. It could be better utilized in a regenerator to contribute to the heating of the fluid between States 2 and 3.
Overall Efficiency Gains If all of these improvements were to be made to the system: expander optimization, pump optimization, and condenser temperature, system efficiency could reach nearly 10% (see Table 6). The increase in efficiency over experimental results of approximately 7% are made up primarily by expander optimization (5.5%), then by pump optimization (1%), and lastly by condenser temperature optimization (.5%).
48 Table 6. Optimized State Calculations p1 T1 h1 s1
103 312 76.985 58.26
Head Efficiency p2 h2 T2
520 0.75 623 77.13 312
State 1 - After Heat Rejection kPa TH2O K mdotH2O kJ/kg J/molK State 2 - After Pump kPa Win(net) Win(gross) kPa kJ/kg K State 3 - After Heat Addition kPa H2OT1 kJ/kg mdotH20 K T2 J/molK Qin State 4 - After Turbine kPa
p3 h3 T3 s3
591.85 267.35 370 166.14
p4 Efficiency h4 T4
108.42 0.75 242.4 kJ/kg 334 K
Power Efficiency mdot
310 K 1.00 kg/s
0.02 kW 0.03 kW
372 1 365 30.43
K kg/s K kW
Analysis 2.994 kW 9.74% 0.16 kg/s
HFE-7000 The use of HFE-7000 as a working fluid in a Rankine Cycle is clearly demonstrated by this project. The fluid functions as other common refrigerants would with the added benefits of being benign to human health and the environment. A comparison between HFE7000 and two other common refrigerants – R-22 and R-134 can be seen in Table 7. These values are chosen to approximate States 3 and 4 of the system experiment, which is across the expander, and to help understand the poor performance of the expander. The ratios of pressure across the states are similar between all fluids, as are the differences in enthalpy (not shown). One noteworthy difference is the relative density of HFE-7000 – over 2 times higher than the other fluids. At this point, it is unknown how HFE-7000 performs relative to other fluids in a sideby-side comparison. Theoretically they are quite similar, but a test would need to be
49 Table 7. Comparison of Refrigerant Density 3
HFE-7000 HCFC-22 HCFC-134
Fluid Density (kg/m ) State 3 State 4 100C, 600kPa 60C, 100kPa 47.11 7.51 17.6 3.1 20.88 3.73
performed to understand how differences, including density, manifest themselves in a realworld situation.
Cost Analysis The estimated cost of the optimized 3kW system is approximately $11,000 (see Table 8 for breakdown). This does not include the cost of an energy source and sink, as these will vary by application. Table 8. System Cost Breakdown Component 3kW expander/generator HFE‐7000 Pump Piping, valves Boiler Condenser Pre‐heater Controller Misc. hardware Total
Price ($) 4500 1400 1200 1000 800 800 600 500 500 11300
For a standalone solar-powered system, a direct-heated collector (lines of black tubing exposed to the sun) could be utilized as a source and a large shaded bladder full of water could act as sink. To power a 3kW system at 10% efficiency, the collector field would need to be approximately 50 square meters, assuming 800W/m2 insolation and 80% conversion efficiency. The collector and sink could be reasonably expected to add $5,000$10,000 of cost, totaling roughly $20,000 for an installed 3kW system. A photovoltaic (PV)
50 system costs approximately $9W-1 installed, or $27,000 for a 3kW system (SolarBuzz, 2010), and will require a similar area. It can be seen that a solar-powered Rankine cycle operating at 10% efficiency is cost competitive with PV sources. However, an efficiency approaching that level has yet to be demonstrated empirically. For industrial waste heat applications with an existing heat source and sink the cost increase for installation is minimal. A 3kW system that costs $12,000 operating around the clock would have a payback period of 6.5 years at current industrial electricity prices of $0.07 kWh-1 (Energy Information Administration, 2010).
51
CHAPTER 5 CONCLUSIONS All primary goals of this project have been achieved:
The use of an HFE fluid in an ORC has been demonstrated.
The use of a low-temperature power source (97°C) has been demonstrated.
The existing mathematical model for HFE-7000 has been both modified and validated so that future systems can be designed using more ideal parameters such as temperature and pressure, or other components.
Low-cost components have been used to demonstrate potential cost effectiveness of such a system relative to PV in solar powered applications.
SUMMARY The use of HFE-7000 as a working fluid offers the potential to recapture energy from a variety of existing and novel heat sources. The relatively low operating temperature of <100°C makes it appropriate for waste heat recapture applications as well as small-scale solar collectors. Such a system could have residential applications if a sufficiently large heat sink were available (such as a swimming pool). The fluid is advantageous in that it is environmentally benign and can provide useful energy at low risk and potentially low cost to the end consumer. Experimentation has shown that the fluid functions in a system made from off-theshelf components. The system achieved a thermal efficiency of 3.1% with no optimization of components or inputs. Further theoretical analysis shows that such a system could be optimized with alternate components to provide a net thermal efficiency as high as 10%. At such efficiency, and without high-priced custom components, the system becomes cost competitive with PV systems in solar powered applications, and has a short payback period for industrial applications. From a comparison with other commonly-used Organic working fluids, it appears that HFE-7000 would function as a drop-in replacement fluid in other ORC systems. A similar compressibility and heat capacity is observed in all fluids. The primary differences are a
52 higher density and vapor pressure in HFE-7000. The conclusion of interchangeability is further strengthened by the experiment performed which required minimal retrofit to a system that had previously run with HCFC-134.
RECOMMENDATIONS Improvements to the present system could be achieved fairly easily:
Optimization of the expander for various inlet pressures and flow rates could be accomplished in a standalone experiment and greatly increase the efficiency of the expander and overall system.
A complete replacement of the pump is recommended. The existing pump is inefficient even at the higher flow rates for which it was designed, causing it to contribute thermal energy to the fluid and complicate measurements and system analysis. A more efficient pump would mitigate these concerns.
The addition of a second water heater would immediately increase the energy available to the system, allowing higher flow rate and increased system performance. Such an increase would require a commensurate increase in the heat rejection capacity of the thermal sink, so a more massive water bath would be required.
Increasing the temperature of the thermal sink would prevent sub-cooling in the liquid phase and increase system efficiency.
Improved measurement instrumentation would improve resolution of system measurements. Thermocouple wells in the hot- and cold-water lines would allow for instantaneous, accurate measurement of the power input to and rejection from the system. This would eliminate the need for time-averaged bulk temperature measurements of the source and sink, which are less accurate because the system changes performance as the water temperatures change. An in-line flow meter to measure flow rate of the working fluid would remove any error caused by the calculation used to find flow rate from the energy input. More accurate flow rate values would allow better correlation to and optimization of the expander performance.
53
REFERENCES 3M Electronics. (2005). Novec 7000 engineered fluid product information. Retrieved April 10, 2010, from http://multimedia.3m.com/mws/mediawebserver? mwsId=66666UuZjcFSLXTtlXftMxMVEVuQEcuZgVs6EVs6E666666. Clemente, S., Micheli, D., Reini, D., & Taccani, R. (2010, June). Numerical model and performance analysis of a scroll machine for ORC applications. In Proceedings of 23rd Annual ECOS Conference. Lausanne, Switzerland. Cogswell, F. (2006). An ORC power plant operating on a low-temperature (165 F) geothermal source. Proceedings of the Geothermal Resources Council Annual Meeting. Curran Associates, Inc. Dai, Y., Wang, J., & Gao, L. (2009). Parametric optimization and comparative study of organic Rankine cycle (ORC) for low grade waste heat recovery. Energy Conversion and Management, 50, 576–582. Energy Information Administration. (2010). Average retail price of electricity to ultimate customers: Total by end-use sector. Retrieved April 10, 2010, from http://www.eia.doe.gov/cneaf/electricity/epm/table5_3.html Environmental Protection Agency. (2010). Emissions. Retrieved April 10, 2010, from http://epa.gov/climatechange/index.html Giampaoloa, M., & Leonardo, M. (2010, June). Model of a steam/organic vapour volumetric reciprocating expander. In Proceedings of 23rd Annual ECOS Conference. Lausanne, Switzerland. Giampaolo, M., & Stefano, P. (2010, June). Model of a steam Wankel expander. In Proceedings of 23rd Annual ECOS Conference. Lausanne, Switzerland. Hettiarachchi, H., Golubovic, M., Worek, W., & Ikegami, Y. (2007). Optimum design criteria for an organic Rankine cycle using low-temperature geothermal heat sources. Energy, 32, 1698-1706. Hung, T. (2001). Waste heat recovery of organic Rankine cycle using dry fluids. Energy Conversion and Management, 42, 539-553. Hung, T., Shai, T., & Wang, S. (1997). A review of organic Rankine cycles (ORCs) for the recovery of low-grade waste heat. Energy, 22(7), 661-667. Husband, W., & Beyene, A. (2008). Low-grade heat-driven Rankine cycle, a feasibility study. International Journal of Energy Research, 32(2008), 1373-1382. Ingley, H.A., Reed, R., & Goswami, D.Y. (2004). Optimization of a scroll expander applied to an ammonia/water combined cycle system for hydrogen production. Retrieved April 10, 2010, from http://www.mae.ufl.edu/NasaHydrogenResearch/ pubprogramreport/Dr.%20Ingley_H2%20Production_for%20submission.pdfvv.
54 Kane, M., Favrat, D., Gay, B., & Andres, O. (2007). Scroll expander organic Rankine cycle (ORC) efficiency boost of biogas engines. Ecos 2007, 2, 1017-1024. Kim, H., Ahn, J., Park, I., & Rha, P. (2007). Scroll expander for power generation from a low-grade steam source. Institute of Mechanical Engineers, 221(5), 705-711. Lemort, V., & Quoilin, S. (2009). Designing scroll expanders for use in heat recovery Rankine cycles. Retreived April 10, 2010, from http://www.labothap.ulg.ac.be/ cmsms/uploads/File/London_paper_proceedings.pdf. Lemort, V., Quoilin, S., Cuevas, C., & Lebrun, J. (2009). Testing and modeling a scroll expander integrated into an organic Rankine cycle. Applied Thermal Engineering, 29(14-15), 3094-3102. Liu, B., Chien, K., & Wang, C. (2004). Effect of working fluids on organic Rankine cycle for waste heat recovery. Energy, 29, 1207-1217. Mago, P., Srinivqasa, K., Chamra, L., & Somayaji, C. (2008). An examination of exergy destruction in organic Rankine cycles. International Journal of Energy Research, 32, 926-938. Mago, P., Srinivqasa, K., Chamra, L., & Somayaji, C. (2008). An examination of regenerative organic Rankine cycles using dry fluids. Applied Thermal Engineering, 28, 998-1007. Manhire, B., (2004). On three-phase reactive power. In Proceedings of the 2004 American Society for Engineering Educational Annual Conference and Exhibition. Salt Lake City, Utah. Giampaolo, M., & Stefano, P. (2010, June). Model of a steam Wankel expander. In Proceedings of 23rd Annual ECOS Conference. Lausanne, Switzerland. Mathias, J., Johnston, J., Cao, J., Priedeman, D., & Christensen, R. (2009). Experimental testing of gerotor and scroll expanders used in, and energetic and exergetic modeling of, an organic Rankine cycle. Journal of Energy Resources Technology, 131(1), 012201 (9 pages). O’Callaghan, P., Badr, O., Probert, S. D., Bell, M. A., & Patel, R. M. (1985). Performance of multi-vane expanders. Applied Energy, 20 (3), 207-234. Owens, J. (1998). Segregated hydrofluoroethers: Low GWP alternatives to HFCS and PFCs. 3M Specialty Materials. Retrieved April 10, 2010, from http://multimedia.3m.com/ mws/mediawebserver?mwsId=TTTTTVv_LdGtMYu75Z0AOy9Tfwvrfdv_IwUTfwUT fTTTTTT Peterson, R., Wang, H., & Herron, T. (2008). Performance of a small-scale regenerative Rankine power cycle employing a scroll expander. Proceedings of the Institute of Mechanical Engineers, 222(3), 271-282. Quoilin, S. (2007). Experimental study and modeling of a low temperature Rankine cycle for small scale cogeneration. Retrieved April 10, 2010, from http://organicrankine.com/ orc_documents/scroll_non_auto/TFE_SQ010607.pdf.
55 Quoilin, S., Lemort, V., & Lebrun, J. (2010). Experimental study and modeling of an organic Rankine cycle using scroll expander. Applied Energy, 87(2010), 1260-1268. Saitoh, T., Yamada, N., & Wakashima, S. (2007). Solar Rankine cycle system using scroll expander. Journal of Environmental Engineering, 2(4), 708-719. Shigemi, N., Toshio, O., & Akira, M. (1999). Scroll expander, internal losses of expander. Transactions of the Japan Society of Refrigerating and Air Conditioning Engineers, 10(2), 123-132. Siddiqi, M., & Atakan, B. (2010, June). Investigation of the criteria for fluid selection in Rankine cycles for waste heat recovery. In Proceedings of 23rd Annual ECOS Conference. Lausanne, Switzerland. Smith, I. Stosic, N., & Kovacevic, A. (1999). Power recovery from low cost two-phase expanders. Centre for Positive Displacement Compressor Technology. Retrieved April 10, 2010, from http://organicrankine.com/orc_documents/screw/screw_expander.pdf. Smith, I., Stosic, N., Kovacevic, A., & Langson, N. (2007). Cost effective small scale ORC systems for power recovery from low enthalpy geothermal resources. Geothermal Resources Council Transactions, 31. Solarbuzz. (2010). Solar Module Retail Price Environment. Retrieved April 10, 2010, from http://www.solarbuzz.com/Moduleprices.htm Tchanche, B., Quoilin, S., Declaye, S., Papadakis, G., & Lemort, V. (2010, June). Economic optimization of small scale organic Rankine cycles. In Proceedings of 23rd Annual ECOS Conference. Lausanne, Switzerland. Teagan, W., & Clay, W. (1973). 3 KW closed Rankine-cycle powerplant with a turbine expander. Thermo Electron Corp. Retrieved April 10, 2010, from http://www.dtic.mil/srch/doc?collection=t3&id=AD0776982 Toffoloa, A., Lazzarettoa, A., Manentea, G., & Rossi, N. (2010, June). Synthesis/design optimization of organic Rankine cycles for low temperature geothermal sources with the HEATSEP method. In Proceedings of 23rd Annual ECOS Conference. Lausanne, Switzerland. Vaja, I., & Gambarotta, A. (2010, June). Dynamic model of an organic Rankine cycle system. Part I – Mathematical description of main components. In Proceedings of 23rd Annual ECOS Conference. Lausanne, Switzerland. Vaja, I., & Gambarotta, A. (2010, June). Dynamic model of an organic Rankine cycle system. Part II – The full model: description and simulation. In Proceedings of 23rd Annual ECOS Conference. Lausanne, Switzerland. Wang, X., & Zhao, L. (2009). Analysis of zeotropic mixtures used in low-temperature solar Rankine cycles for power generation. Solar Energy, 83(5), 605-613. Yamamoto, T., Furahata, T., Arai, M., & Mori, K. (2001). Design and testing of the organic Rankine cycle. Energy, 26, 239-251. Yanagisawa, T., Fukuta, M., Ogo, Y., & Hikichi, T. (2001). Performance of an oil-free scroll-type air expander. IMechE C591, (27), 167-174.
56 Zanelli, R., & Favrat, D. (1994). Experimental investigation of a hermetic scroll expandergenerator. Proceedings of the International Compressor Engineering Conference. West Lafayette, Indiana. Zyhowski, G., Brown, A., & Achaichia, A. (2010, June). HFC-245fa working fluid in organic Rankine cycle - a safe and economic way to generate electricity from waste heat. In Proceedings of 23rd Annual ECOS Conference.
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APPENDIX SCROLL DEVICE PROPERTIES
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Figure 24. Compressor performance data under various conditions.
59
Figure 25. Compressor specifications.
