Green Machine Organic Rankine Cycle Field Test May December 2013

Transcription

1 Green Machine Organic Rankine Cycle Field Test May December 2013 Principal Investigator: Chuen-Sen Lin Project Manager: Daisy Huang Other Participants Gwen Holdmann Vamshi Avadhanula David Light Thomas Johnson Ross Coen David Pelunis-Messier University of Alaska Fairbanks PO Box Fairbanks, AK August 27, 2014

2 DRAFT REPORT: PROJECT TITLE: An Analysis of Organic Rankine Cycle Green Machine System Performance in Alaska Green Machine Organic Rankine Cycle Laboratory and Field Test COVERING PERIOD: November 2011 December 2013 DATE OF REPORT: August 27, 2014 RECIPIENT: PROJECT LEAD: PROJECT MANAGER: OTHER PARTICIPANTS: Alaska Energy Authority 813 West Northern Lights Boulevard Anchorage, AK Chuen-Sen Lin Department of Mechanical Engineering University of Alaska Fairbanks P.O. Box Fairbanks, AK Daisy Huang, Alaska Center for Energy and Power, UAF Department of Mechanical Engineering University of Alaska Fairbanks P.O. Box Fairbanks, AK Gwen Holdmann, Alaska Center for Energy and Power, UAF Vamshi Avadhanula, UAF David Light, Alaska Center for Energy and Power, UAF Thomas Johnson, Alaska Center for Energy and Power, UAF

3 Acknowledgement We would like to thank the following individuals and organizations for their contribution to this project: David Pelunius-Messier, Tanana Chiefs Conference for your collaboration throughout this project and many others; The Alaska Energy Authority, Denali Commission, Alaska Department of Environmental Conservation, and US Environmental Protection Agency for funding various aspects of this project; Alaska Power and Telephone and in particular Ben Beste, Vern Neitzer, and the rest of the crew at the Tok powerhouse for field testing the unit; Devany Plentovich, Program manager for the Alaska Energy Authority for her guidance and eminent patience with this effort particularly the final report; McKinley Services for assisting with the installation; Mike Ruckhaus, UAF Facility Services and Chilkoot Ward and the staff at the UAF Powerplant for allowing us to test the Green Machine at their facility; Maria Richards from the Southern Methodist University Geothermal Lab; Mark Hall with the Heat is Power Association; and David Sjoding of Washington State University for peer reviewing this report. ii

4 Executive Summary This study involved testing and evaluation of a 50 kw, model Block 1 Green Machine (GM), an Organic Rankine cycle (ORC) system capable of generating power from low-quality heat sources, such as heat rejected from a diesel engine. The testing was conducted in two phases. Under Phase I, the GM was installed at the University of Alaska Fairbanks power plant and operated under controlled conditions to determine power output under different heating and cooling rates. It was also operated at full load for 600 hours for reliability testing and 400 subsequent hours for other testing. Under Phase II, the GM was installed at the Alaska Power and Telephone (AP&T) power plant in Tok, Alaska. A major goal of the Phase II study, the ORC field test, was to determine the potential for critical issues to occur during real-life application of the GM in village diesel power plants and to assess performance and economic potential based on the field test data. The GM as installed in the Tok power plant ran for 1,137 hours. Altogether, laboratory and field tests combined, the GM ran for approximately 3,000 operating hours. During that time, the system was considered reliable, and maintenance requirements were considered minimal. The system performed close to its nameplate capacity of 50 kw under controlled laboratory conditions and about half that under field conditions, averaging 22 kw net power output. This lower output was expected because of the operating conditions in the field deployment (379 kw heat supplied to power unit evaporator). After 1,137 hours of operation in Tok, the system developed a refrigerant leak due to an incompatibility between the seal material and the lubricant used in the refrigerant. The manufacturer, ElectraTherm, reports that this problem is known and has been fixed in the current comparable updated version of the GM, which is a Block 5, 65 kw ORC system. Currently, insufficient funds prevents repair of the GM in Tok, and it remains idle in the AP&T powerhouse. As of this writing, AP&T is in the process of negotiating with ElectraTherm to replace the unit with an updated model, the Green Machine While the unit was operational, AP&T saved an estimated 1350 gallons of diesel fuel, or $6740 during the period of time the unit was operational. Assuming a full year of operation and the use of a more efficient cold water pump, total fuel savings could be 12,370 gallons of diesel fuel, which would save the utility $61,850 per year based on current (2012) fuel prices. A more realistic scenario in which the unit would be operated annually from April 1 st though November 15 th would result in a savings of 7,730 gallons of fuel, representing $38,640 annual savings. Under this scenario, the simple payback assuming 0% interest on the capital expense and no subsidy is 10.7 years. We expect this payback would be the same or better for one of ElectraTherm s newer units. ElectraTherm reports that 27 commercial units are in operation worldwide, including two machines with over 17,000 hours run time and an additional six machines with over 10,000 hours run time. Most of these are the newer series 4200, 4400, or However, this study necessarily limits itself to reporting on the performance of the model used in its laboratory and field tests. iii

5 Table of Contents Acknowledgement...ii Executive Summary... iii List of Figures... vii List of Tables... ix Abstract... 1 Introduction Motivation Background Phase I: Laboratory Testing Introduction Laboratory Test Preparation Selection of a Low-temperature Heat Engine for Testing Required Elements for Test Plan and Test Site System Modeling, Simulation, and Test Parameters Selection Heat Source Heat Sink ORC System Modeling Methodology Selections of Component Parameters and Operation Parameters Simulation Case Study Test Component and Measurement Device Selections Using Simulation Data: Test Plan, Design and Selection of Components, and Procurement Test Plan and Design/Selection of Components Procurement Installation and Instrumentation Commissioning Experimental Setup and Test Schedule Experimental Setup Testing Schedule Parameters Measured and Data Reduction Parameters Measured iv

6 2.8.2 Data Reduction Reductions in Emissions and CO Economic Analysis Reliability Testing and Results Preparation Green Machine Setup Parameters Hot Water Loop Setup Parameters Cold Water Loop Setup Parameters Operation Procedure Checklist Reliability Test Results Green Machine Shutdown during the Reliability Test Performance Testing and Results Performance Test Operation Procedure Results Data Analysis, Performance Curves and Example Based on Developed Performance Curves Analysis Results from Reliability Test Performance from Reliability Test Analysis Results from the Performance Test Discussion Based on Performance Test Results Example Based on the Performance Curves Further Discussion of Adopting the GM for a Rural Genset Discussion (Economics, Emissions, Findings) Findings of General Information Related to the GM ORC Findings in GM Performance and Comments on Applications Match between the GM ORC and a Diesel Generator Set Economic (Payback Period) and Performance Estimates Proposed General Policy in GM Application Conclusions from Laboratory Testing Accomplished Tasks Related to Project Objectives Preparation for Phase II Field Testing v

7 3 Phase II: Field Testing Project Preparation and Test Plan Installation at the Tok Power Plant Heat Source Cooling Source Instrumentation and Measurement/Evaluation Equipment Test Schedule and Test Plan Procurement of Required Equipment and Supplies Transport and Installation of the Green Machine Operation and Maintenance Requirements Data Collection Results from Field Testing Field Data Performance and Economic Analysis Discussion Issues Encountered Lessons Learned Summary References Appendix A. Appendix B. Heating and Cooling System Design Green Machine Startup Parameters, 27 Aug Appendix C. Materials Used in Green Machine Installation Appendix IA Survey of Low Temperature Heat Engine Companies (2008) Appendix IB Survey of Low Temperature Heat Engine Companies (2010) Appendix IIA Preliminary Line Diagram of the Testing System Appendix IIB Preliminary Components Selected for the Testing System Appendix IIIA Methodology Proposed for Stage 2 and Stage 3 Modeling Appendix-IIIB Expressions for Single Phase and Two-Phase Heat Transfer Coefficient of Fluids in Plate Heat Exchangers Appendix IVA Available Floor Space for Test System Installation Appendix VI Estimated additional kw that could be generated using an ORC unit for all Alaskan communities vi

8 List of Figures Fig. 1. Schematic of Organic Rankine Cycle system... 9 Fig. 2. Efficiency of ORC system with varying screw expander inlet quality for heat source temperature of 200 F (93 C) Fig. 3. Efficiency of ORC system with varying screw expander inlet quality for heat source temperature of 250 F ( C) Fig. 4. Design line diagram of the testing system Fig. 5. GM Setup parameters screen Fig. 6. GM Startup Parameters screen Fig. 7. GM Options parameters screen Fig. 8. GM Machine Defaults parameters screen Fig. 9. GM PLC I-O parameters screen Fig. 10. GM Veris Setup parameters screen Fig. 11. Green machine HMI screen-shot during reliability test operation Fig. 12. Hot water and cold water supply temperatures to ORC power unit during reliability test Fig. 13. Net power generated, power consumed by power unit pump and hot water pump during reliability test Fig. 14. Heat input to power unit evaporator vs. hot water flow rates at different hot water supply temperatures and cold water flow rates Fig. 15. Heat rejected to cold water in power unit condenser vs. hot water flow rates at different hot water supply temperatures and cold water flow rates Fig. 16. System operating power output vs. hot water flow rates at different hot water supply temperatures and cold water flow rates Fig. 17. Heat input vs. hot water supply temperature Fig. 18. Heat rejected vs. hot water supply temperature Fig. 19. System operating power output vs. hot water supply temperature Fig. 20. System operating efficiency vs. hot water supply temperature Fig. 21. Payback period vs. hot water supply temperature Fig. 22. CO 2 reductions vs. hot water supply temperature Fig. 23A. Pay period at 0% interest rate on capital for different Green Machine ORC power outputs, fuel prices, and capital costs Fig. 23B. Pay period at 10% interest rate on capital for different Green Machine ORC power outputs, fuel prices, and capital costs Fig. 24. Schematic line diagram of heating and cooling to the Green Machine vii

9 Fig. 25. Well Number Fig. 26. GM as installed in the Tok power plant Fig. 27. GM cold- and hot-water piping Fig. 28. Data acquisition system for Green Machine field test in Tok Fig /40 propylene glycol/water supply and return temperatures to Green Machine Fig. 30. Green Machine operating power output, Green Machine pump power, and coldwater pump power consumption Fig. 31. Comparison of Green Machine UAF lab test results and Tok field results for net power output for same heat source supply temperatures Fig. 32. Payback period at 0% interest rate on capital for different Green Machine ORC power outputs, fuel prices, and capital costs (for full year of operation) Fig. 33. Payback period at 10% interest rate on capital for different Green Machine ORC power outputs, fuel prices, and capital costs (for full year of operation) Fig. 34. Green machine net power output and net efficiency versus hot-water supply temperature viii

10 List of Tables Table 1. Thermodynamic properties and environmental date of R-245fa... 8 Table 2. Operation and performance parameters for data acquisition Table 3. Selected components for heating and cooling loops Table 4. Various hot water and cold water flow rates at which GM will be tested Table 5. Tier 4 interim EPA emissions standards for non-road diesel engines Table 6. Total component cost incurred in building the experimental system Table 7. GM Setup parameters table with range and default values Table 8. GM Startup Parameters table with range and default values Table 9. GM Options parameters table with range and default values Table 10. Checklist for GM, hot water and cold water loops Table 11. Reliability test results at three different times of the test Table 12A. Various hot water and cold water flow rates at which GM was tested Table 12B. Actual Input Conditions for GM Performance Testing Table 13A. Performance results for HW Temp = 155 F; HW flow rate = 120 gpm to 300 gpm; CW Temp 50 F and CW flow rate = 120 gpm, 160 gpm, and 200 gpm Table 13B. Induced performance results from measured readings of Table 13A Table 14A. Performance results for HW Temp = 175 F; HW flow rate = 120 gpm to 300 gpm; CW Temp 50 F and CW flow rate = 120 gpm, 160 gpm, and 200 gpm Table 14B. Induced performance results from measured readings of Table 14A Table 15A. Performance results for HW Temp = 195 F; HW flow rate = 120 gpm to 300 gpm; CW Temp 50 F and CW flow rate = 120 gpm, 160 gpm, and 200 gpm Table 15B. Induced performance results from measured readings of Table 15A Table 16A. Performance results for HW Temp = 215 F; HW flow rate = 120 gpm to 300 gpm; CW Temp 50 F and CW flow rate = 120 gpm, 160 gpm, and 200 gpm Table 16B. Induced performance results from measured readings of Table 16A Table 17A. Performance results for HW Temp = 225 F; HW flow rate = 120 gpm to 300 gpm; CW Temp 50 F and CW flow rate = 120 gpm, 160 gpm, and 200 gpm Table 17B. Induced performance results from measured readings of Table 17A Table 18A. Performance results for HW Temp = 155 F to 220 F; HW flow rate = 120 gpm to 300 gpm; CW Temp 68 F and varying cold water flow rate Table 18B Induced performance results from measured readings of Table 18A Table 19. Reliability test results Table 21. Diesel engine specifications ix

11 Table 22. Estimated ORC performance for operating on waste heat recovery from diesel engine Table 23. Instrumentation equipment and components for data collection from the GM 50 kw field test at Tok, Alaska Table 24. Parameters measured and instrumentation used during Phase II testing Table 25. Reduced form of the recorded Tok field test data and generated Green Machine ORC performance data during the field test period Table 26. ORC performance during continuous operation period from 10/02/2013, 11:00 A.M., to 11/19/2013, 7:00 A.M. ( hours) Table 27. Estimated ORC performance in Tok, Alaska, for a full year of operation (8,760 hours) x

12 Abstract This report describes the testing and evaluation of an ElectraTherm 50 kw, model Block 1, Organic Rankine cycle Green Machine as completed by the Alaska Center for Energy and Power at the University of Alaska Fairbanks. The Green Machine was tested in two phases. Under Phase I (Laboratory Testing), the Green Machine was installed at the University of Alaska Fairbanks (UAF) power plant and run under controlled conditions to determine power output at different heating and cooling rates. It was also run under full load for reliability testing for a total test time of over 1,000 hours. Following the UAF tests, the unit was deemed suitable for deployment in a village power plant, and under Phase II (Field Testing) the community of Tok was selected to host the demonstration. The unit was installed at the Alaska Power and Telephone power plant in late summer of After its commissioning on October 2, 2013, the Green Machine operated for 1,138 hours before a seal on the screw expander failed, which caused refrigerant to leak. Prior to the failure of the seal, the Green Machine performed as expected. This report describes the study s objectives, the testing plan, the design and fabrication of the testing system, data collection, the Green Machine abbreviated field test performance and an economic analysis based on that performance, the lessons learned, and recommended guidelines for Organic Rankine cycle system selection and application for rural Alaska village diesel generators.

13 1 Introduction This study involved testing and evaluation of a 50 kw, model Block 1 Green Machine (GM), an organic Rankine cycle (ORC) system, and took place in two phases. Phase I involved laboratory testing under controlled conditions at the University of Alaska Fairbanks (UAF) power plant. The GM was run for 50 hours under different prescribed conditions of heating and cooling and for 1,000 hours at full load. Phase II involved field testing under real-world conditions at the Alaska Power and Telephone (AP&T) power plant in Tok, Alaska. The major goal of the Phase I study was to operate the GM in a controlled environment to determine its reliability and gain first-hand information about its performance. It was intended that the results would be used by the funding agency, the Alaska Energy Authority (AEA), as well as utilities and vendors to ascertain whether this technology has the potential to improve overall plant efficiency by using rejected heat from diesel generators to generate additional electricity by employing ORC technology. Phase I results suggested that in circumstances where excess heat is available and a sufficient difference in temperature can be achieved between the excess heat utilized from the diesel engines and a cooling side for operating the evaporator, the GM may be an economical option worthy of consideration. Based on these results, funding for a field test Phase II was sought and awarded. The primary goal of the Phase II study, the ORC field test, was to evaluate the performance of the GM during operation in a village diesel power plant. Other goals included gathering and analyzing performance data, evaluating operation and maintenance requirements, analyzing the economics of potential power generation given the operating parameters of the local power plant, and determining factors that should be considered when deciding whether to deploy ORC technology in an Alaska diesel power plant. 1.1 Motivation Isolated rural villages in Alaska annually consume about 30 million gallons of diesel fuel to generate 370,000 MWh of electrical energy, produced by individual diesel-fired generator sets. In general, the ratio of electrical power produced to fuel energy consumed is less than 40%. The rest of the fuel energy is lost as heat. While some power plants in Alaska s rural villages use a portion of this energy for other beneficial uses, such as space and water heating, most of this energy is wasted. The goal of adding ORC products to an existing power cycle is to reclaim some of this heat to generate more power, thus increasing the overall fuel efficiency of the power plant. While ORC technology is mature for larger-scale power generation, the products for smaller-capacity generator sets, appropriate for the typical size of Alaska village power plants, are still new to the market or are in the prototype phase. Many villages have been approached by ORC product developers to invest in this new technology, so the Alaska Center for Energy and Power (ACEP), together with its funding partners, the Denali Commission and AEA, established a program to test the viability of ORC products in Alaska. 1

14 ElectraTherm s Green Machine (GM) was identified as one of the devices with the highest potential for success. The GM discussed here, an early Block 1 model, is designed to generate up to 50 kw of power using the ORC, which is a process that can obtain energy from lowervalue (lower-temperature) heat sources than are commonly used for power generation. (As of this writing, the current market-ready version of the GM 65 kw machine is the 4400; the Block 1 50 kw machine described in this report is no longer being sold by ElectraTherm.) 1.2 Background In 2010, ACEP partnered with Tanana Chiefs Conference (TCC), a nonprofit consortium of 42 communities in Interior Alaska, to obtain and test a 50 kw GM. In 2011, the unit was purchased by TCC with grant funding from the Denali Commission. In November, ACEP installed the unit in the coal-fired power plant at the University of Alaska Fairbanks (UAF), using heat from the plant s steam loop tempered with potable water. The GM was operated for 50 hours under different controlled heating and cooling conditions, and for 1,000 hours of reliability testing under full load. TCC facilitated communication between ACEP and the villages under its umbrella to select an Alaska community whose power company would be willing to field test the GM. The community of Tok was selected by TCC, due to its road accessibility and relative proximity to Fairbanks, availability of waste heat at least seasonally, and the competence, expertise, and interest of the staff managing the local powerplant, operated by Alaska Power and Telephone (AP&T). 2

15 2 Phase I: Laboratory Testing 2.1 Introduction In rural Alaska, approximately 180 villages consume about 370,000 MWh [1] of electrical power annually using isolated diesel generator sets. In part because diesel fuel is imported from long distances often just once a year and then stored in bulk, the cost of fuel, and hence the cost of generated power, is very high. If waste heat from the diesel generators can be captured and used for either space heating or supplemental power generation, the fuel savings are significant. Many applications for low quality heat, such as heat rejected from a diesel engine, have been demonstrated. Examples include general heating (e.g., space heating and city water temperature maintenance), direct thermal-to-electricity conversion, heat-to-power conversion using a heat engine, refrigeration, and desalination. Of these applications, waste heat for heating is considered the most efficient application and is commonly practiced seasonally in Alaska. However, in some cases, waste heat for heating is not practical for reasons that include lack of proximity between the powerhouse and buildings, or prohibitively high construction costs and incompatibility of building heat systems. Waste heat for heating in Alaska village diesel generators has been discussed in detail [2]. Waste heat for power through heat engines is recommended under appropriate circumstances because of its acceptable efficiency (i.e., close to 10%), flexibility in heat utilization, and expected low maintenance (similar to steam engine or refrigeration systems). In addition, unlike heating, power is needed year-round. Power usage in many of Alaska s rural villages is about or below 1 MW. For these generator sets, the power produced using waste heat is expected to be below 100 kw. For waste heat engines belonging to this category, the power-to-cost ratio is expected to be very high if the heat engine is facilitated with a radial turbine (a type of expander) for heat-to-power conversion. Many different thermodynamic cycles and different types of heat-to-power expanders have been used to lower the cost. Examples of thermal cycles include the ORC and the ammonia/water (or Kalina) cycle. Examples of heat-to-power conversion expanders include the screw expander, scrolling expander, and piston expander. All of these examples are still in the prototype and proving stage or the prototype fabrication stage. It is well understood that the performance of a heat engine depends on conditions of the heating source and cooling source, both of which largely rely on the load pattern and waste heat properties (e.g., exhaust, jacket coolant) of the diesel generator set and the cooling source available in the village. Therefore, to estimate the performance and economic impact of any waste-heat engine on an individual diesel generator set, the performance data of the heat engine under various heating and cooling conditions are needed. In general, these data are obtainable by testing the heat engine under controlled heating and cooling conditions. The Phase 1 study had four objectives. The first was to demonstrate that an improvement of the efficiency of the diesel power plant by about 10% (i.e., about 4% of fuel efficiency) is achievable through the use of an Organic Rankine cycle (ORC) system, which uses waste heat contained in diesel engine jacket water and exhaust. The second objective was to evaluate the 3

16 feasibility, operation and maintenance requirements, and payback time of applying a selected ORC system. The third objective was to develop guidelines for ORC system selection, operation, and maintenance, and to evaluate the potential impact of applying waste-heat ORC systems on rural Alaska economy, fuel consumption, and emissions and greenhouse gas reductions. The fourth objective was to compare the performance and economics of two ORC systems: (1) a 50 kw system that uses a screw expander and is an emerging technology, and (2) a Pratt & Whitney (P&W) 250 kw unit that uses a radial turbine and is a comparably well-developed technology. 2.2 Laboratory Test Preparation Preparation for Phase I included the selection of an appropriate low-temperature heat engine, the layout of required elements for the testing plan (critical parameters, etc.), and the selection of a testing site (e.g., utility, heat source, and heat sink) Selection of a Low-temperature Heat Engine for Testing The proposed project began by surveying (in 2008 by Jared Kruzek) accessible manufacturers who are involved in the low-temperature heat engine industry about their industrial applications and potential and willingness to deliver, within a reasonable time, a lowtemperature heat engine with a power capacity between 10 kw and 100 kw. Eighteen manufacturers were contacted, and their general product information was reviewed. The manufacturers selected for further consideration included an ammonia/water system manufacturer and an ORC system manufacturer. The ammonia/water system manufacturer was contacted because of its previous credible working experience with the Alaska power industry, because its ultra-small size of 10 kw could be applied in Alaska s smallest and most high-cost communities, and because it showed its complete design layout and willingness to spend its own funds for fabricating a prototype system. In 2010, pre-shipment testing conducted by the ammonia/water system manufacturer showed that a major component (the heat-to-power conversion unit) of the system had two major drawbacks, which hampered overall system performance and caused repeated delays in delivery. While the manufacturer continued work on improving the product, no timeline for delivery could be established. The drawbacks included (1) migration of lubrication fluid from the power unit into the loop of working fluid, which lowered the heat transfer efficiency significantly after a short period of operation, and (2) much lower than expected heat-topower-conversion efficiency of the power unit. Also in 2010, the ORC manufacturer showed promising test results for its ORC prototype. The After conducting an updated market survey to assure no other promising technologies were overlooked, ACEP decided to move forward with procuring and testing the ORC, which was a 50 kw rated unit manufactured by ElectraTherm under the model name Green Machine, Block 1 model. 4

17 2.2.2 Required Elements for Test Plan and Test Site The test plan included testing the ORC system for reliability and performance. The heat engine required a heat source, which provides driving energy to the heat engine and emulates the heating conditions (i.e., temperatures and flow rates) received from the waste heat generated by the village diesel generators. A heat engine also needs a cooling source, which absorbs the dissipated heat from the heat engine and emulates the variety of cooling conditions possibly provided by the cooling sources existing in the villages (e.g., surface and ground water, radiator, and cooling tower). In addition, heat energy transmitting devices (e.g., heat exchanger, pipe, pump, and valve) are needed to transmit heat between the working fluid in the ORC system and the media of the heating and cooling sources. Other examples of required system elements include devices for safety and reliable performance. The purpose of the reliability test was to observe the endurance in operation and consistency in performance of the ORC system under the rated operation condition. The purpose of the performance test was to look at performance details of the ORC system and its components (i.e., efficiencies, energy consumptions) under numerous operation conditions of heating and cooling. There were more parameters to be measured and measurement devices to be installed for the performance test than for the reliability test. To guarantee that all measurement components needed for this project were included in the final testing system design, the preliminary design line was diagrammed, showing all the parameters to be measured. The line diagram of the preliminary design and required components are given in Appendix IIA along with information on components (Appendix IIB). Based on the preliminary design and components information, the requirements for space, facilities, and utilities of the test site were then estimated. The selection of the final design concept needed to be conducted based on the existing facility and resources of the testing site, available overall test budget, desired operation and maintenance requirements, and time constraints. 2.3 System Modeling, Simulation, and Test Parameters Selection In order to find the performance of the ORC system and the performance of its individual components, sensors and measured data of physical properties (i.e., temperatures, pressures, and flow rates) of the working fluid pertinent to the components are needed. The collected data are then analyzed to give the performance results. Preliminary designs and selection of appropriate testing components (e.g., sizes and types) are obtained through the process of system modeling and simulation. Modeling and simulation may also help determine operation parameters that are critical to system performance. By applying testing data, the model can be further improved and simulation results may become good enough to be useful in predicting long-term performance and applying the ORC system to any individual diesel generator. The sections that follow describe the model construction process for different stages of modeling and corresponding functions, and a constructed first-stage model for performance simulation and testing components selection. The model includes three components: heat source, heat sink, and the ORC system. The fluid used in heating and cooling loops was water; the working fluid used in the ORC system was R-245fa refrigerant. 5

18 2.3.1 Heat Source The physical heat source loop for the new test site was expected to include a hot water source from a steam/water heat exchanger, a VFD pump, and pipes and fittings. Other components for measurement and control are included. Heating fluid enters the heat source heat exchanger of the ORC system through pipes, and transfers heat to the working fluid, the R-245fa refrigerant. The loop controls the temperature and flow rate of the existing hot water from the steam/water heat exchanger. All important information from the steam and heating water loop (e.g., fluid temperatures, pressures, flow rates, VFD rpm, pump power consumption) for each operating condition was collected for system and components performance analysis. The model constructed corresponding to the heat source included all important operation parameters (i.e., temperature and flow rate control and pump power evaluation) of all the function features. The model could be modified easily to cope with different types of heat source and components Heat Sink At the selected test site, the physical heat sink loop included a cooling fluid source from a fire hydrant and its manual flow rate control valve, pipes, and fittings, temperature and flow rate measurements, and control devices. The cooling fluid entered the condenser of the ORC system through the pipes. The loop had limited controllability in temperature and flow rate of the cooling fluid entering the condenser. Information on fluid properties along the pipes, as well as power and water consumption corresponding to conditions of each operation, were collected. The model constructed corresponding to the cooling source included all important operation parameters (i.e., temperature and flow rate control and pump power evaluation) of all the function features. In addition, the model could be modified easily to cope with different types of cooling sources and components ORC System A general ORC unit includes at least a pump, an evaporator, a heat-to-power converter, and a condenser. Other components needed in modeling depend on the versatility of the physical construction of the ORC system. For example, one known property of the ORC system with a screw expander is its ability to allow mixed vapor/liquid working fluid in the heat-to-power conversion unit (in this case the screw expander), so that the system can add a component to control the working fluid flow rate and/or quality of fluid entering the expander to optimize system performance. The physical ORC system used for this project was an integrated unit, which may not be practical for conducting accurate measurements of working fluid properties for performance analysis of individual components without modifying the system (modifying the system may result in losing the warranty). Also, detailed engineering information on individual components may not be available due to concerns about intellectual property. The system does allow the addition of more sensors to access approximate working fluid properties related to performance of many of the components. This feature is helpful in getting better analysis results of the components. 6

19 2.3.4 Modeling Methodology Detailed engineering data for individual components may not be needed for testing ORC system reliability and physical performance related to system power generation under different heat source and sink conditions. Detailed engineering data for components needed for ORC optimal net power generation also may not be needed. In other words, to achieve the objectives listed in the Introduction, detailed engineering data of the ORC system components may not be critical. However, if the purpose of a test is to check that the operation condition (input to the ORC) is optimal and to offer comments on system design, detailed engineering data of components are required. To model the system with reasonable accuracy, the modeling plan was divided into two to three stages, depending on how feasible and desirable it was to know the details of the performance parameters of the ORC system and its components. The first stage involved modeling the ORC system using simple values to represent system performance parameters based on specifications of the ORC system and its components. The purpose of this stage was to qualitatively understand the effects of the operating conditions of the heat source and heat sink on ORC system performance and to find approximate ranges of operation of the parameters of system components. The results would be used for test planning and selections of test system components and measurement devices. The second stage of modeling was to fit the system and component parameters using limited measurement data obtained from limited experimental cases (i.e., heating and cooling conditions). If system simulation results obtained using fitted (approximate) values of system parameters can qualitatively match experimental results (although without appropriate accuracy), extra experimental cases could yield a more complex and detailed model as the third stage. The purpose of the second and third stages of modeling was to develop a simulation model that can predict system performance at any system operation condition. Since modeling was at the component level, the performance prediction was acceptable to the operation conditions, which were moderately outside of the operating conditions covered in the test. The methodology of the first stage is described in detail here. The methodologies of the second and third stages are given in Appendix IIIA for reference purpose only. In the first stage, the selected ORC system was able to optimize the system net output for all heating and cooling conditions, limiting the total output to a maximum of 50 kw. There are other features related to conventional constraints for performance regulation and system protection, such as constraints in maximum temperature and maximum pressure. In a physical prototype, these constraints are imposed by physical mechanisms, such as mechanical and electrical devices. To simplify the mathematical model, some of these functions can be emulated using a conditioning computation statement or neglected because they have much less influence than other functions on the gross performance of the ORC system. The simplified model (Figure 1) includes the evaporator, a screw expander, a condenser, a VFD pump, and controlled heating and cooling sources. The model also allows the quality of the working fluid entering the expander to be adjusted by varying the flow rate of the working fluid for optimal ORC system performance. Simulation results are useful for test planning and test system component or sensor selections. The expander was modeled by a single efficiency at this stage 7

20 and will be modified as more information becomes available in publications and through experimental data. The pump was modeled with varying efficiency based on its operation condition, and the evaporator and condenser were modeled by their respective flow and heat transfer properties and heat transfer areas. The evaporator has the capability to model liquid, liquid/vapor mixture, and vapor flows. Currently, the condenser is modeled as a single section unit. If heat loss to the atmosphere is found significant, heat loss will also be included in the model Selections of Component Parameters and Operation Parameters Ranges of values of operation parameters used for simulation were based on specifications of the components (e.g., hot water flow rate and temperature limits, pressure and temperature limits of the ORC system), properties of the fluids (i.e., heating, cooling, and ORC working fluid), performance of sub-components (e.g., heat exchanger performance versus flow rate), and properties of heat and cooling sources. Known limits, which mostly are based on the recommendation of the manufacturer, included maximum pressure of the ORC system (150 psi), estimated heat source temperature (235 F), controlled heat source capacity (2.4 MMBtu), flow rates of pumps (250 gpm for heating, 375 gpm for cooling), and cooling sink capacity (3.0 MMBtu recommended by an ORC engineer). The values for ORC component parameters adopted from publications (limited data available) and conventional application practice [3] for system simulation included expander efficiency (e.g., 0.78), pump efficiency (e.g., 0.70), heat transfer coefficient of evaporator (e.g., 1500 W/m 2 - K or 265 Btu/ft 2 - F), evaporator area (e.g., 100 ft 2 ), heat transfer coefficient of condenser (e.g., 1400 W/m 2 - K or 247 Btu/ft 2 - F), and condenser area (e.g., 200 ft 2 ). Other limitations included maximum heat source temperature (225 F) and minimum cooling source temperature (50 F). Some of the values were adjusted based on the match between the simulation results and the published and experimental data. The heat exchanger simulation model was based on standard practice [4, 5]. Since the working fluid (R-245fa) properties would affect the ORC system performance and the temperature and pressure limits used for testing, some of the physical properties were obtained (see Table 1). Detailed properties of R-245fa can be found in NIST documents. The results of the first stage simulation included system performance (e.g., net efficiencies, expander power) of the ORC system, and net efficiencies of the test system as functions of operation parameters of the heat source and heat sink. The results include the effects of sizes of heat exchangers and efficiencies of expander and pumps on system performance. The first stage results were used to help design the test plan for the ORC system. Table 1. Thermodynamic properties and environmental date of R-245fa Safety Vaporization Heat (1atm.) Non Kj/Kg Flammable (355.5 Btu/lb) * Ozone depletion potential ** Greenhouse warming potential Boiling T.(1atm.) 14.6 C (58.3 F) Critical Point 154 C (309.2 F) 36.4 bar (527.9 psi) Saturation Slope ODP* GWP** 100 year Isentropic Besides simulating results in system performance, the process also helped with estimating sizes and capacities of components and measurement devices needed for the testing system. 8

21 The first stage of test system modeling was completed and the effects of flow properties of heat source and heat sink on net efficiencies of the ORC system and the test system (i.e., including parasitical power consumptions through heating loop and cooling loop) were obtained from simulation. In addition, temperature drops of heating flow and cooling flow crossing the ORC system were useful for selections of Btu meters for heating and cooling loops and the steam/hot water heat exchanger. Simulated thermodynamic states of working fluid along the ORC system were useful for selection of pressure gauges for components performance monitoring. A schematic of the ORC system used in this simulation model is shown in Fig. 1. The working fluid used in the simulation model was refrigerant R-245fa. The saturated liquid refrigerant from the condenser was pumped at high pressure to the pre-heater. In the pre-heater the refrigerant was heated to the saturated liquid state; the saturated liquid then went to the evaporator. In the evaporator the saturated liquid was superheated or saturated (may include vapor or vapor/liquid mixture). This high-pressure working fluid was converted to low-pressure liquid or vapor/liquid mixture (to the condenser pressure) using a screw expander, which is connected to the generator to produce power. The low-pressure refrigerant from the screw expander was cooled to the desired state in the condenser, and the liquid portion was pumped back to the pre-heater, and the cycle continued. Fig. 1. Schematic of Organic Rankine Cycle system The ORC system model has three major components: the heat source loop, the heat sink loop (open loop), and the ORC system. In the heat source loop for diesel generator waste heat application, the heating fluid may come from the engine jacket water or from a 50/50 glycol/water mixture exiting the exhaust heat exchanger or both combined. In the heat sink loop, the cooling fluid may be from the cooling tower, radiator, or a large body of water, such as from a nearby river or lake. 9

22 Modeling and simulation was a continuous process. As more system component information was available from experiments and literature, the model was updated to predict more realistic and accurate performance results for given heat and cooling conditions, which were expected to function for the model at the second and third stages. Detailed system component information, such as for the screw expander and for boiling and condensing heat transfer coefficients of refrigerant in the evaporator and condenser, was not available in the literature and could only be evaluated based on experimental data (if experimental data of the ORC system was accessible). The data obtained from the experimental analysis was used to tune the model so that in the future it can be applied to any waste heat source for economic and feasibility analysis of the ORC system. The intention of the second stage and third stage (Appendix IIIA) modeling was to enable the model to be capable of comparing the performance of the ORC system operated under different diesel engine load and environmental conditions, not to reengineer the design of the ORC system. The difference between the second stage and third stage modeling is in the complication of the modeled component (i.e. the model parameters representing the component). In the first stage simulation, five system parameters were being controlled; they included inlet temperature and flow rates for heat source and heat sink input loops and the state of the refrigerant inlet to the expander. The quality of refrigerant inlet to the expander was controlled to investigate its effect on the power output and efficiency of the ORC system for given heat source and heat sink conditions. The system simulation had been performed for different screw expander inlet refrigerant states for given heat source and heat sink inlet conditions. The heat source and heat sink inlet conditions were flow rate and inlet temperatures of respective fluids (here water was considered for both heat source and sink). In the current preliminary simulation, the following assumptions were made: 1. All the ORC heat exchangers, i.e., evaporator, pre-heater and condenser, are 100% efficient. 2. The quality of refrigerant out of the evaporator in the ORC system is controlled. 3. The quality of liquid out of the pre-heater and condenser are saturated liquid. 4. The isentropic efficiency of the screw expander and pump (within the ORC system) are constant at 78% and 70%, respectively. Assumptions affect system performance characteristics. If simulation results do not match the published or measured performance characteristics of the real system, a test plan must be designed to determine which assumptions need to be changed. The model can be modified easily, once better performance characteristics of components are obtained from experiments. In simulating the performance of the ORC system, explicit formulae for heat transfer coefficients of refrigerant on one side of the heat exchanger and water on other side of the heat exchanger should be known. Generally, the heat transfer coefficient of a fluid is expressed in terms of its thermodynamic and transport properties. The heat transfer coefficient also depends heavily on the geometry of the heat exchanger and material of construction. 10

23 All heat exchangers considered in the present case were plate heat exchangers (PHE). A widely accepted expression for heat transfer coefficient of single-phase fluids in a plate heat exchanger is given by Muley and Manglik [6]. In the ORC system, this expression is used for calculating the heat transfer coefficient of hot water and cold water in the evaporator and the condenser, respectively, and the heat transfer coefficient of hot water and refrigerant in the pre-heater. In all of the above cases of heat transfer coefficient calculation, the fluid thermophysical properties were taken at average fluid temperature. An expression for heat transfer coefficient of evaporating refrigerant liquid-vapor mixture in the evaporator is given by Ayub [7]. The expression for heat transfer coefficient of condensing refrigerant liquid-vapor mixture in the condenser is given by Selvam et al [8]. All the above expressions are presented in the Appendix IIIB Simulation Case Study The constructed ORC system model has been used to simulate an example of an ORC system of 50 kw with the system parameters mentioned above, and heat exchanger parameters and computation method listed in Table IIIB-1 of Appendix IIIB. The efficiency versus expander inlet quality is shown in Fig. 2 and Fig. 3. These figures also show the effect of parasitic power and heat sink supply temperature on system efficiency. The parasitic power is the power needed to pump the heat source and heat sink fluids to and from the ORC system. As the heat sink supply temperature decreases (in this case from 21 C to 5 C), the same amount of heat from the condensing refrigerant in the condenser less the amount of cooling fluid must be removed, which may decrease the parasitic power and increase the efficiency of the system. This result may be one of the advantages of using the ORC system during the winter months. 11

24 8 Heat Source Temperature of 200F Efficiency (%) System Pump Work Only All Pump Work (Tc=21C) All Pump Work (Tc=5C) Expander Inlet Quality (kg/kg) Fig. 2. Efficiency of ORC system with varying screw expander inlet quality for heat source temperature of 200 F (93 C) 12 Heat Source Temperature of 250F 11 Efficiency (%) 10 9 System Pump Work Only All Pump Work (Tc=21C) All Pump Work (Tc=5C) Expander Inlet Quality (kg/kg) Fig. 3. Efficiency of ORC system with varying screw expander inlet quality for heat source temperature of 250 F ( C) 12

25 The effect of heat source temperature on the efficiency of the ORC system is also shown in Fig. 2 (200 F or 93 C) and Fig. 3 (250 F or C). As the heat source temperature increases, the efficiency of the ORC system increases. Here in simulating the system for different heat source temperatures, the screw expander inlet pressure conditions (or evaporator exit conditions) were different, though all other system parameters remained the same (i.e., condenser pressure, expander and pump efficiencies, and heat source flow rate). For a heat source temperature of 93 C, the expander inlet pressure was 6.95 bar, and for C, the pressure was 15.7 bar. More work is produced by the expander when it goes from high pressure at the expander inlet to the same condenser pressure, and this may be the reason for an increase in system efficiency for different heat source temperatures. Based on the given model, simulation results show that the higher temperature of the heating fluid and the lower temperature of the cooling fluid give better system performance. Higher flow rates for both heating and cooling flows also give better system performance. However, the increase in system performance may be capped by the size of the respective heat exchangers. For the current model, the efficiencies of the pump and screw expander were considered constant, which indicates a monotonically increasing relationship between the system efficiency and the efficiency of the working fluid pump or the screw expander. However, the ORC system may have a different optimal working fluid flow rate at each operation condition, defined by heating and cooling flow conditions. This suggests that, in order to better simulate system performance, the model of the system needs to include parameters of the working fluid flow rate and other related parameters, such as those that define performance of pumps, screw expander, evaporator, and condenser. These parameters can be estimated from matching the model simulation results and measured data, of which the best possible data come from approximate measurements (i.e., steady state temperatures measured from the outside surfaces of uninsulated tubes of the working fluid). For the structure of this ORC system, the data from the approximate measurements may be sufficient for the purpose of this project Test Component and Measurement Device Selections Using Simulation Data: Based on simulation results, the maximum heating flow (supply side temperature) and minimum heating flow (return side temperature) of the ORC system are 250 F and 143 F, which are useful for selecting an appropriate Btu meter for heat flux measurement. This information is also useful for the selection of an appropriate steam/hot water heat exchanger for hot water temperature and flow rate control. Similarly, simulation results help select an appropriate Btu meter for the cooling side. Based on the simulation results of pressures along the ORC system loop (i.e., working fluid entering the expander at 235 psia, existing the expander at 30 psia, entering the pump at 23 psi, and exiting the pump at 240 psi), pressure gauges (i.e., 0 50 psi for the low pressure side of the ORC loop and psi for the high pressure side) are selected for working fluid pressure measurements. 13

26 2.4 Test Plan, Design and Selection of Components, and Procurement Test Plan and Design/Selection of Components The test plan included two parts: a 600-hour reliability test and a 50-hour performance test. The reliability test focused on the ORC system s reliability and performance consistency during long-term operation. The performance test focused on identifying system and component performance characteristics under different operation conditions. The required equipment to obtain the desired information for these two tests included (1) a heating source and a cooling source with controllable fluid flow conditions, (2) a power uploading (generated by ORC system) and consuming (e.g., by pumps) circuit, (3) measurement and data acquisition devices for measuring and collecting operation and performance data of critical components in the system, and (4) an Internet connection between the ORC system and ElectraTherm facilities. Other necessary accessories included a piping pressure regulator, safety devices, and filters. The ORC system is a completely self-regulated system; some of its critical components are system performance measurement and control/safety/monitoring devices. The interface between the ORC system and the environment is through the heating fluid inlet and exit ports (for heating fluid), cooling fluid inlet and exit ports (for cooling fluid), the UAF electrical motor center (for power uploading from the ORC system to the UAF power plant, and power consumption of pumps and actuators), and Internet connection box (for remote monitoring and control from the ElectraTherm company). Manual interface devices between human and machine include HMI, an emergency switch, and manually operated valves). The ORC system does not have enough measurement devices to evaluate the performance of all the major components (expander, evaporator, condenser, and pump). If performance evaluation is required for every major component, extra sensors for temperatures and pressures pertinent to the components are needed. Due to the integrated nature in manufacturing the ORC system, temperature sensors cannot be installed in direct contact with the working fluid. Therefore, temperature sensors are installed to the outside surfaces of pipes, immediately next to the fluids at the locations of interest. Since the measurements were for steady-state operation conditions, the deviations between the measured temperatures and the respective true temperatures of fluids were expected to be small enough to have minimal effect on the conclusions of this project. The final selected test site was a vacant space in the UAF power plant building. The UAF power plant has plenty of steam supply for heating, water supply for cooling, and pressurized deionized water for pipe pressure maintenance. All of the available resources made the installation and operation expenses low and affordable. The final design layout of the heating and cooling loops were obtained through several design iterations, and the design was based on the availability of the test space, project budget, accessible resources, codes and standards, and timing constraints. The final testing system includes a closed steam loop coupled with a closed hot water loop for heating, an open cooling water loop for cooling, an electrical circuit for power uploading/consumption, instrument and signal processing circuits for data acquisition/performance monitoring and control, and an Internet connection for remote monitoring/control. A detail of the available floor space for installation is given in Appendix IVA. 14

27 Detail of design layout is given in Fig. 4, and the selected components for heating and cooling are listed in Table 3, with brief information on component selection and specification. The following subsections discuss more details of the individual subsystems mentioned: Heat source loop: Again, the heat source loop can be subdivided into two loops: the steam loop and the hot water loop. The steam-to-hot water heat exchanger is the component that connects both of the loops. This heat exchanger is used to transfer heat from steam to hot water. 1. The main installed components of the steam loop are steam/water heat exchanger, steam flow control valve, steam trap, and condensate piping. 2. The main installed components in the hot water loop are hot water VFD pump, air separator, pressure relief valve, expansion tanks, BTU meter, temperature mixer, and ON/OFF solenoid valves. 3. The steam/water heat exchanger, VFD pump, and temperature control valves are used to control hot water temperature and flow rate flowing into the ORC system to simulate heat source conditions obtained from different diesel engine loads. Temperature mixers are used for maintaining uniform temperature throughout the cross section of the pipe for more accurate heat input measurements. The ON/OFF solenoid valves are used in the case of emergency shutdown of the power unit so that the hot water bypasses and flows back to the heat source instead of the power unit. The BTU meter is used for measuring the flow rate and amount of heat released by the hot water to the evaporator and pre-heater of the power unit. 15

28 # Component # Component # Component 1 Green machine (GM) 12 Pressure relief valve 23 Check valve 2 Expansions joints 13 Steam-to-hot water heat exchanger 24 Steam trap 3 Solenoid valve 14 Steam control valve (may be automatic) 4 Ball valve 15 Steam manual control valve 5 Strainer or Filter 16 Pump (Constant flow rate) 6 Drain 17 Hydrant source (Cooling water source) 7 Temperature mixer 18 Cooling water from GM (Hydrant sink) 8 BTU meter 19 Bypass for temperature control on coolant side 9 VFD Pump 20 Steam inlet 10 Expansion tank 21 Steam condensate outlet 11 Rolairtrol air separator 22 Temperature control valve Fig. 4. Design line diagram of the testing system 16

29 Table 2. Operation and performance parameters for data acquisition Hot water loop parameters Cold water loop parameters GM parameters (sensors installed by our team) GM parameters (sensors come with the ORC system) Electric power parameters Steam Inlet Temperature to Heat Exchanger Cold Water temp into GM Before Check Valve GM Ambient Temp GM Gross Power (kw) - NOT accurate GM net power output (Watts) Steam Condensate temperature Cold Water temp into GM just before GM GM Condenser Inlet Temp (sensor on pipe surface) Temperature difference between hot and cold water supply temperatures (F) GM Pump power (Watts) Hot water temp out of Heat exchanger (to GM) Cold Water temp out of GM after bypass valves GM Condenser Outlet Temp (sensor on pipe surface) GM pump VFD Hz Hot water pump power (Watts) Hot water temp into Heat exchanger (from GM) GM Hot Water Inlet Temp Cold Water temp out of GM condenser before pump Cold water flow rate GM Evaporator Inlet Temp (sensor on pipe surface) GM Evaporator Outlet or Expander inlet Temp (sensor on pipe surface) Expander Hi Pressure (PSIG) Expander Low Pressure (PSIG) Cold water pump power (Watts) GM Hot water outlet Temp Heat rejected to cold water in GM condenser in MWH GM Expander Outlet Temp (sensor on pipe surface) Expander Rear Temp (F) (sensor on pipe surface) Steam Inlet Pressure GM Pump Inlet Temp (sensor on pipe surface) Bearing Temp (F) Steam Condensate Pressure GM Pump Outlet Temp (sensor on pipe surface) CW Supply Temp (F) (sensor on pipe surface) Steam Valve Position GM Refrigerant Tank Temp (sensor on tank surface) CW outlet Temp (F) (sensor on pipe surface) Hot water flow rate GM Condenser Inlet Pressure HW Supply Temp (F) (sensor on pipe surface) Hot water heat supply to GM in MWH GM Evaporator Inlet Pressure HW outlet Temp (F) (sensor on pipe surface) GM Pump Inlet Pressure Generator runtime hours GM Pump Outlet Pressure Lube oil pressure (PSIG) Refrigerant Tank Pressure Generator RPM GM Hot Water Bypass Temp (sensor on pipe surface) GM pump Power (kw) GM Hot Water out of PH before bypass (sensor on pipe surface) Total GM kwh from starting GM Net Power (kw) 17

30 Table 3. Selected components for heating and cooling loops Green machine (GM) Steam to Hot water heat exchanger Steam flow rate control valve with actuator Steam trap VFD pump for hot water loop Expansion tank Air separator Pressure relief valve BTU meters 3-way butterfly valve Cold water pump Component Size Reason for selection The only commercial unit available in market at that time Power output: Max 50kW, Min 5kW Hot water supply: 180 o F to 250 o which can recover heat to power from diesel engine jacket F. Cold water supply: 40 o F to 110 o water temperatures. Heat transfer media temperature F requirement. i. Heat exchange rate: 4000MBTU/hr ii. Area = ft 2 i. Valve Spec.: Valve flow coefficient (Cv): 160; ii. 4 Normally closed Siemens valve iii. Actuator: SKC62U Specification: SpiraxSarco FT-30; Maximum operating Pressure: 30psig; Maximum Temperature: 45 o F of superheat at all operating pressures i. Bell & Gossett 20hp pump (1750rpm) rated for VFD operation. ii. Rated for 250gpm and head of 116feet of water Extrol SX-40V; Pre-charged to 12psi. Bell & Gossett Rolairtrol R-4F air separator: 300gpm and 4 flange Bell & Gossett 45psi pressure relief (PR) valve Kamstrup BTU meters on hot water loop and cold water loop Triad 4 3-way butterfly valves with double acting actuator 15hp Scott Pump with 3500rpm i. Based on heat source requirement for GM. At 5.5% efficiency (worst case scenario), the GM would need 4000MBTU/hr for upgraded GM of 65kW. ii. Area of HX was selected with communication with heat exchanger manufacturer i. The reason for using steam flow control valve with actuator is to manipulate hot water supply temperature to GM from 155 o F to 250 o F. ii. The valve flow coefficient (Cv) was calculated from standard manufacturer s handbooks. The calculated Cv obtained was about 135. Steam trap was selected based on power plant steam supply pressure range of 18psig to 24psig. i. Pump was selected based on the pressure drop in the hot water piping calculated using pressure drop across various components. ii. Pump was also selected in communication with pump manufacturer. iii. VFD is for various flow rates of hot water. Expansion tank volume selection: ASRAE handbook standard procedure with a factor of safety margin (for high temperature and above atmosphere pressures of hot water). The calculated expansion volume based on handbook was 3.5gal. Air separator selection was based on the maximum hot water flow (300gpm) in the loop. 45psi PR is used to accommodate for above boiling water temperature of 235 o F. For measuring the heat supplied by hot water to GM and heat rejected to cold water by GM. Temperature and flow rate requirements. Two three way butterfly valves were used in the cold water bypass loop to control the cold water supply temperature to GM by recirculation of warm water coming out of GM Cold water pump will only be used during the performance testing of the GM when the recirculation of warm water is needed to test the GM at varying cold water temperatures. Check valve Walworth 3 150psi check valve To prevent cold water flow back into fire hydrant The cable size was selected based on the distance between 3/0 metal-clad (MC) three conductor cable the GM and power plant motor center (point for power uploading to UAF grid), voltage drop, and heat generated due Electrical cable for uploading with ground (Due to unavailability of 1/0 power cable and time concern) to cable resistance. According to standard handbook calculation 1/0 cable was enough to serve the purpose. Grid protection relay Beckwith grid protection relay model # M- 3410A To protect mainly the utility (grid system) from GM and to protect the GM itself. If there is large voltage (±10%) or frequency fluctuations in the utility system the GPR sends a signal for GM shutdown. 18

31 Heat sink loop: 1. The main installed components of the heat sink loop include the hose from the fire hydrant to the power unit, pump (not VFD), BTU meters, check valve, two 3-way flow and temperature control butterfly valves, bypass loop, temperature mixer, and hose from power unit to a ditch. 2. The fire hydrant valve, 3-way butterfly valves, pump, and bypass loop are used for flow rate control and the water inlet temperature of the GM cooling side. 3. The BTU meter is used for measuring the flow rate and amount of heat released by the working fluid to the cooling water in the condenser of the power unit. The check valve is to prevent water flow-back into the fire hydrant. Electrical system: 1. The electrical system installation is the electrical wiring required for various components (ORC power unit, VFD pump, cold water pump, etc.) in the whole test system and for safety concerns. The wiring diagram is given in Appendix IVB. 2. Electrical wiring was successfully completed for uploading the power generated by the GM to the UAF grid. Electrical wiring was also completed for powering the VFD pump on the hot water loop, pump on cold water loop, steam control valve, and two 3-way bypass valves in cold water loop. 3. A grid protection relay was used to protect the grid and the ORC system. Instrumentation: Performance characteristics of the ORC system and its components were obtained from measured data through data analysis. Performance characteristics are defined in terms of various input conditions of the heat source and heat sink. The performance characteristics were then used for selecting appropriate rural Alaska diesel gensets for this ORC system and planning waste heat distributions for optimal returns from waste heat applications. Operation and performance parameters detected and collected for this test are summarized in five groups listed in Table 2. Besides the measurement devices originally equipped with the ORC system, measurement devices for all the other parameters listed in Table 3 needed to become available through purchasing or other means. Remote monitoring and control line: Remote monitoring and control was done from both UAF and the ElectraTherm facility in Nevada simultaneously. For this purpose, the broadband Internet connection cables, static IP address and modem were installed at the UAF power plant experimental site Procurement Considering the integrated testing system (ORC system plus supporting testing system), equipment purchased included the 50 kw ORC system and most of the components used in steam loop, hot water loop, the open cooling water loop, electrical components used for the electrical circuit, and sensors and data acquisition components used for measuring/monitoring and control. Table 3 shows the major components used for steam, heating, the cooling system, and electrical systems. Major components purchased for measurement/monitoring and control 19

32 system included a National Instrument PXI board and SCXI modulus. In addition, steel and black iron hot water pipes, steam rubber hoses, wires, cables, structural materials, and related connection and construction materials were on the purchase list. Before December 2011, all purchases had been completed except the sensors needed for measuring temperatures and pressures related to the performance of the ORC components and two Watt meters for pump power measurements, which were completed in mid-january Some of the components and construction tools (steam trap, steam-to-hot water heat exchanger, expansion tank, pipe grooving tool, cutting tool, welder) were borrowed from UAF Facilities Services and the UAF power plant. One lesson learned is that procurement may become extremely time consuming for a remote city in a remote state. According to our experience, sometimes, even common structural materials may take a few days to become available. 2.5 Installation and Instrumentation Installation and instrumentation started around mid-november 2012 and most of the work was complete within 3 weeks. The rest sensors to the ORC system components and Watt meters to the hot water and cold water pumps was installed in January Detection, rectification, and modifications were part of this task. All the installation work (e.g., wiring, piping) followed existing industrial standards and codes, as well as the requirements specified in Chapter 1 of the ElectraTherm manual (Green Machine Installation Requirements). Since regulations, safety, and reliability are important to this project, experts (e.g., contractors, UAF power plant, UAF Facilities Services, ACEP, and the UAF College of Engineering and Mines [CEM]) in related areas were consulted for piping, wiring, and control system design and installation. All the work was done internally by UAF personnel, including engineers in the units of Facilities Services, the power plant, and ACEP. Other individuals contributed to this work including many students and a few staff and faculty members from the CEM machine shop and Mechanical Engineering Department. The following is a brief list that describes relative important information related to the ORC system installation and instrumentation (also refer to Fig. 4, Table 2, and Table 3): Steam loop: Piping: Steel pipe or steam rubber hose (300 F and 30 psi). Joints: Threaded with sealant or welding. Steam/hot water heat exchanger: Eccentric condensate flange for outlet of steam side. Heat exchanger: A designed frame, used to keep an appropriate height of the water outlet position. Hot water loop: Piping: 4 black iron pipe with Gruvlok joints. Struts structures were used to support the 150 ft pipe system at many critical locations. 20 hp pump: A designed structure was used to support the potential large bending load. Pipe entering and exiting the ORC system and bypass pipe (including a control valve): Steel pipe provided by the ORC manufacturing for entering, exiting, and bypass pipes. A designed 20

33 supporting structure was required for the overhung pipes and the valve. BTU meter (ultrasonic): 3 ft straight pipe distance was required in the upstream side. The flow meter needed to be installed in the cold side pipe. Cold water loop: Piping: 4 rubber hose and fire hydrant clamps were used. A check valve was installed to prevent reverse flow. Pumps: Needed a structure to support the bending load. Inlet, exiting, and bypass pipes (including controlled valves) of the ORC system: Support structure was required for the overhung pipes and the valve. BTU meter (ultrasonic): 3 ft straight pipe distance was required in the upstream side. Electrical circuit: Upload Contactor: 3/0 MC Cable with 3 conductors and a ground. Ceiling tray was used to harness the cable. Pump: Cable has 3 lines and a ground. Instrumentation and DAQ: Sensors installed: Numerous thermal couples were used for temperature measurements, two thermistors for the two BTU meters, 6 manual pressure gauges, 7 digital pressure gauges, 2 Watt meters, and 2 flow meters (i.e., BTU meters). All digital signals connected to the NI SCXI 1001 box. DAQ system: Included a NI PCI-MID-16E-1 board and NI SCXI modules (1102, 1124, four 1120, and two 1121 modules). Internet lines: One line connected to the power plant server and the other directly to the local computer. Appendix V shows pictures of the integrated system after installation. 2.6 Commissioning The commission started on December 12, 2012, and lasted for about a week. Before commissioning, the ORC system (the Green Machine) user needed to install the heat source, heat sink, electrical circuit, and a dedicated Internet line to the GM based on the requirement list provided by ElectraTherm (ET), the manufacturer of the GM. ElectraTherm needed to receive the completed GM Installation Requirement list and a Pre-Commissioning Checklist before an engineer could be dispatched for commissioning. During commissioning, the following was performed by the ET commissioning engineer: Checked the installation of the supporting systems: This included investigation of all the installed heating, cooling, and electrical systems and the Internet connection. There were many discussions between the ET engineer and the UAF engineers. 21

34 Checked the Green Machine: A double check of the information and locations of the labels, installation and functionality of the components, match of the specification received by the customer and the specification of the manufacturer s record, etc. Working fluid transferring: Conducted a pressure test and transferred working fluid from storage bottles into the GM. Details are found in the GM manual (Chapter 5 Fluid Transferring). Parameter setup: Start parameters, operation parameters, default parameters, and other parameters using HMI (Human Machine Interface) were set up. Details are found in the GM manual (Chapter 7 Setup). Startup of the GM: Test of emergency stop, operation mode, alarm mode, and PLC-IO. Details are found in Section 3.4 (Operation) and Chapter 7 (Distributor Manual) of the GM manual. Training: A comprehensive training was held for ACEP engineers involved in this project. The training included the most important materials covered in the GM manual: information on installation, routine maintenance and special tools, operation, background, fluid transferring, tasks and applications of HMI. This part of the training was for the general audience (many of the ACEP engineers, students, faculty, and staff members). Hands-on system setup and operation training: This training was for an engineer and a graduate student who would perform the GM test. In addition, some design issues were discussed. Two important examples are listed below: 1. All the components of the GM were designed for a capacity of 65 kw, but the system was rated as a 50 kw system. This downgrade was due to the lack of an emergency bypass valve to prevent the screw expander from over-speeding while the engine was shutting off. This unit can be modified with a retrofit to make it operate at 65 kw safely. 2. To start the GM, the temperature of the heating fluid entering the GM could not be too high, or the screw expander would over-speed and the engine would shut down immediately. This occurs because of the methodology used to control the starting working fluid flow rate. 2.7 Experimental Setup and Test Schedule Experimental Setup The experimental site was set up at the UAF power plant. Setup included a 50 kw GM heat engine (the ORC system), a heating source system to supply heat to the GM, a cooling sink loop to absorb the expelled heat from the GM, electrical cables for uploading the generated electricity and electrical power consumption to the operation pumps of the heating and cooling loops. Some of the specification information includes electrical characteristics of the GM (i.e., VAC/2 phase/100/60hz, 50 kw, 24VDC), hot and cold water connection (heating source 150 o F 235 o F and 160 gpm, cooling source 40 o F 110 o F and 250 gpm, working fluid R- 245fa), mechanical specifications ( and 7,300 lb), environmental specifications (storage 20 o F 140 o F, relative humidity 0.95, noise 96 db at 3 ), and controls (inputs HMI, physical switch, and emergency stop button; output relay contacts for heating and cooling water supplies, hot water bypass, and dry cooler control; remote control remote monitoring 22

35 and machine control via Internet). Setup for the heating system, cooling system, external electrical system, data acquisition system, and Internet connection are discussed in detail earlier in this chapter (section headings Test Plan, Design and Selection of Components, and Procurement and Installation and Instrumentation ) Testing Schedule The reliability test was a 600-hour test of the GM under full capacity under constant operation conditions (Planned: heating source at 160 gpm and 220 o F; cooling source at 160 gpm and 50 o F). The performance test on the GM was conducted mainly to learn the performance characteristics of the GM and its components (expander, evaporator, condenser, pump) at different hot water and cold water flow rates and temperatures. Another purpose of this test was to create performance characteristic charts of the GM at different heat source and heat sink conditions. Testing of the GM was planned at 5 different hot water flow rates, 5 different hot water temperatures, 3 different cold water flow rates, and 2 different cold water temperatures. The list of hot water and cold water temperatures at which the GM was tested is given in Table 4. The number of GM performance tests planned was 150. The procedure of performing this test (i.e., the method of changing temperatures and flow rate, and data collection) is explained under the heading Operation procedure in Section 2.9. The preparation work and the checklist for the performance test are similar to those for the reliability test, and readers are referred to the description under Reliability test in Section 2.9. Table 4. Various hot water and cold water flow rates at which GM will be tested Hot water temperatures ( o Hot water flow rate Cold water Cold water flow rate F) (gpm) temperatures ( o F) (gpm) Parameters Measured and Data Reduction In this section, two main topics are discussed: parameters measured and data reduction. The parameters measured are the measurements taken from the GM reliability test and performance test. The data reduction description gives the mathematical expressions and methods used to reduce the data obtained from reliability and performance test into conventional parameters to represent the actual performance of the system Parameters Measured The various parameters measured during reliability and performance tests of the 50 kw ORC power unit were (i) hot water flow rate, inlet and outlet temperatures to power unit (, ), (ii) cold water flow rate, inlet and outlet temperatures to power unit (, ), (iii) electrical power output of power unit ( ), (iv) electrical power 23

36 consumed by power unit pump ( ), (v) hot water pump power ( ), and (vi) cold water pump power ( ). Note that the electrical power output of the power unit ( ) already considers the power unit pump electrical power consumption. Cold water pump power consumption was estimated based on hot water pump power consumption due to use of a fire hydrant as a cold water source, which made power consumption of the cold water flow not measurable Data Reduction The mathematical expressions and methods used in obtaining the derived parameters from measured parameters are given in this section of the report. This information was useful in further analysis of the power unit. This section also discusses the procedure and methodology adopted to estimate the reductions in emissions and CO 2 and the economic impact of installing an ORC power unit at a village diesel power plant. Heat supplied ( ) by hot water to the evaporator of the power unit is obtained by (1) Here, density of hot water ( ), inlet enthalpy ( ), and outlet enthalpy ( ) of hot water to power unit were obtained based on evaporator hot water inlet and outlet temperatures and using the NIST REFPROP 8.0 program. is the average density of hot water obtained at the inlet and outlet evaporator hot water temperatures. Heat rejected ( ) to cold water by condenser of power unit is obtained by Here, density of cold water ( ), inlet enthalpy ( ), and outlet enthalpy ( ) of cold water to power unit were obtained based on condenser cold water inlet and outlet temperatures and using the NIST REFPROP 8.0 program. is the average density of cold water obtained at the inlet and outlet condenser cold water temperatures. The screw expander power output ( ), which is given by Eq. 3, is the power generated by the power unit expander without considering any pump power consumption (i.e., without considering the power unit pump (working fluid pump), and hot water and cold water pumps, whereas the system operating power output ( ) is the power generated by the power unit, uploaded to the UAF power system, given by Eq. 4, which considers the power unit pump and cold water pump powers. Here, in calculating system operating power output ( ), only the ORC working fluid pump and cold water pump power consumptions were considered, because in general a stationary diesel engine is equipped with a jacket water pump to dissipate heat to the atmosphere using air coolers. In most of rural Alaska, diesel gensets are equipped with a jacket water heat recovery system, which may have a pump already installed. Taking this into account, the electrical power consumed by a hot water pump is neglected assuming the already installed jacket water pump can be used to overcome the ORC power unit evaporator pressure drop. is used in the calculations on annual diesel fuel saved, emissions reductions, and (2) 24

37 economic outcome discussed in following paragraphs. Here, both and are measured parameters. Screw expander efficiency ( ), ORC power unit efficiency ( (also called ORC net efficiency), and system operating efficiency ( ) are estimated using Eq. 5, Eq. 6, and Eq. 7 respectively: (3) (4) (5) (6) Liters (or gallons) of diesel fuel saved per year ( ) were calculated using Eq. 8, which was based on system operating power output ( ), 363 power unit working days per year with two days of maintenance, and stationary diesel engine specific fuel consumption. Stationary diesel engine specific fuel consumption of 3.7 kwh/lit (14 kwh/gal) is a reasonable value for rural Alaska village diesel gensets. Dollar amount saved on diesel fuel per year ( ) was calculated based on diesel fuel saved per year ( ) and diesel fuel cost of $5.0/gal, which is a reasonable value for rural Alaska stationary diesel generator power plants Reductions in Emissions and CO2 As the ORC power unit was designed to operate on waste heat from a village diesel genset (i.e., a free heating source), it would offset some of the power needs of the village and, in turn, lead to emissions reduction. Annual emissions reductions were estimated based on the annual system operating power output by power unit (363 power unit working days per year with two days of maintenance) and stationary diesel engine emissions (based on a Tier 4 emissions standard), given in Table 5. Annual CO 2 reductions were based on liters (or gallons) of diesel fuel saved per year ( ). Table 5 gives the Tier 4 interim emissions standards set by EPA for non-road diesel engine gensets. Table 5. Tier 4 interim EPA emissions standards for non-road diesel engines NO X Particulate matter (PM) CO HC CO 2 g/kwh (lb/kwh) g/kwh (lb/kwh) g/kwh (lb/kwh) g/kwh (lb/kwh) kg/lit (lb/gal) 3.5 ( ) 0.10 ( ) 3.5 ( ) 0.40 ( ) 2.66 (22.2) (7) (8) 25

38 2.8.4 Economic Analysis The economic impact of installing an ORC power unit at a rural Alaska village power plant was evaluated based on payback period calculations. The payback period is determined when enough money has been accumulated at a given simple interest rate to offset the total initial investment cost ( ) and annual maintenance/operation cost ( ) based on annual cost savings. Here annual cost savings is the dollar amount saved on diesel fuel per year ( ) in operating the ORC power unit on recovered waste heat from a rural Alaska diesel engine power plant. Note that dollar amount saved on diesel fuel per year ( ) was calculated based on 363 power unit working days per year with two days of maintenance, as explained earlier. The total initial investment cost ( ) can be further divided into component costs ( ) and installation costs ( ). Component costs ( ) are the material and instrumentation cost incurred on building the whole heat recovery system and data acquisition system. For the present case, the component costs ( ) included the cost of purchasing the ORC power unit, steam-to-hot water heat exchanger, steam valve, hot and cold water pumps, air separator, expansion tank, pressure relief valve, pipes for hot water and cold water, flow meters, thermocouples, Gruvlok fittings, supporting structural material (e.g. struts, pipe hangers), electrical cables, other miscellaneous parts (nuts, bolts, tees, pipe couplings), and freight charges. Table 6 gives the categorized component costs incurred in building the experimental system. The component cost ( ) of the experimental setup was estimated at $191,500. Table 6. Total component cost incurred in building the experimental system Total component cost Steam loop cost = Power unit cost = Hot water loop cost = Cold water loop cost = Electrical system cost = Instrumentation cost = Structural material cost = Miscellaneous parts and other costs = Maintenance parts One-time cost= Total component cost = The installation cost ( ) may include the number of days for installation, number of personnel required for installation, cost of labor per hour per person, travel cost (if any), and other installation costs. Based on our experience with the ORC power unit experimental system, it requires five personnel and 30 days to complete installation of the hot water loop, cold water loop, electrical system, and instrumentation (assuming all the components are available for installation). Assuming a labor cost of $70/person/hour, and $5,000 for travel, the total cost of 26

39 installing ( ) the whole system is $89,000. This value was used in the payback period calculations. The total initial investment cost ( ) is estimated at $280,500, which is the sum of the component costs ( ) and the installation costs ( ). According to the power unit manufacturer and from our reliability test experience, the maintenance requirement for the ORC machine is similar to the maintenance requirement for air-conditioning and refrigeration systems, and minimal in economic concerns. The expected maintenance is mostly visual inspection and simple measurements, small changes (belts, lubricant, filters, and batteries), and simple cleaning jobs. Considering the maintenance requirements for the exhaust heat recovery system, from previous experience, it is determined that two days of maintenance per year is required. The annual maintenance cost ( ) can be further divided into three major costs: labor, travel, and parts. Labor costs for maintenance is estimated at $2,250, assuming $70/person/hour, two personnel required for two days of maintenance (eight hours per day). A travel cost of $5,000 and parts cost of $350 is estimated. Parts cost may include cost of the drive belt, lubricant, batteries, etc. Therefore, the total annual maintenance cost ( ) is estimated at $7,600. We extrapolated potential additional kw that could be generated using an ORC unit for each community in Alaska in Appendix VI. It must be noted that many of these communities do not produce adequate waste to operate a 50 kw Green Machine, which in principle required a 500 kw average diesel power electric output to operate at full rated capacity. In addition, many of these communities use rejected heat for space and water heating applications during the winter months, including Tok. This means that waste heat may only be available to operate an ORC seasonally. A full analysis of the potential annual production and cost savings for the Tok powerhouse is included in Section 3.10 on page 198 of this report, using actual measured output from the field testing in Tok. 2.9 Reliability Testing and Results Preparation The focus of preparation work was to set up the test system (the ORC system or Green machine, hot water loop, cold water loop, and electrical circuit) and the needed measurement, data acquisition, and monitory devices for safe and smooth running of the ORC system and reliable data Green Machine Setup Parameters The GM setup parameters have default factory settings that are good for normal operation conditions and do not need changing. The GM setup has six HMI screens; of which only three are important for setup and startup. The six HMI screens are Setup, Startup Parameters, Options, Machine Defaults, PLC I-O, and Veris Setup. Among these six screens, the first 27

40 three are only considered for setup and startup of the GM; they are explained in detail in the following tables. The other three screens are shown for reference. 1. Green Machine HMI screen: Setup The GM Setup parameters screen is shown in Fig. 5, and the parameters are explained in detail in Table 7. Fig. 5. GM Setup parameters screen 28

41 Table 7. GM Setup parameters table with range and default values Parameter and Description Range Default Value DELTA T SHUTDOWN: This is the hot water and cold water supply temperature difference 50 o F to 80 o F 80 o F ENTER MINIMUM HZ RANGE: This is the GM pump minimum frequency 15Hz to 30Hz 22.0Hz ENTER MAXIMUM HZ RANGE: This is the GM pump maximum frequency 30Hz to 60Hz 58.0Hz MIN KW NET OUTPUT: This is the GM minimum net power output 5kW to 25kW 10.0kW MAX EXP PRESSURE: This is the maximum expander inlet pressure 130PSI to 200PSI 185PSI EXP DIF PRESSURE MIN: This is the minimum expander differential pressure 0PSI to 70PSI 32PSI ENTER ON GRID : This is the expander RPM at which it should go on grid 1400rpm to 1800rpm 1730rpm ENTER MAXIMUM KW: This is the maximum gross power output required from the GM. 10kW to 50kW 50.0kW HW SHUTDOWN TEMP F: This is the maximum hot water supply temperature at which GM will shutdown 190F to 250F 235F ENTER POWER FACTOR: This is the power factor which is required to match the frequency of GM to frequency of 0.60 to power utility to which GM power is uploaded ENTER FULL LOAD AMPS: This is the current amperes the GM needs to generate at full load 30A to 130A 75A ENTER LINE VOLTAGE: This is the line voltage of the utility to which the GM power is being uploaded 240V to 500V 480V 2. Green Machine HMI screen: Startup Parameters The GM Startup Parameters screen is shown in Fig. 6, and the parameters are explained in detail in Table 8. 29

42 Fig. 6. GM Startup Parameters screen Table 8. GM Startup Parameters table with range and default values Parameter and Description Range Default Value HW MIN START TEMP F: Hot water minimum supply temperature to start the GM 150 o F to 230 o F 150 o F HW MAX START TEMP F: Hot water maximum supply temperature to start the GM. 185 o F to 250 o F 230 o F HW RANGE MINUTES: This is the time required for the hot water to make one complete loop. 0 to 20minutes 20.0 minutes MIN START DELTA T: This is the minimum temperature difference between hot water and cold water supply for 80 o F to 110 o F 80 o F starting the GM TIME BETWEEN STARTS: This is the time GM waits to start after shutdown due to not serious issues such as low hot 1sec to 360secs 30secs water supply temp, Delta T shutdown etc. START DELAY: This is the time GM waits after the TIME BETWEEN STARTS ends and keeps on trying to start at this 1sec to 600secs 15secs regular intervals KW STARTING SETPOINT CONFIGURATION PARAMETERS: This parameters are used to set the initial kw the GM needs to produce at the startup based on the hot water supply temperature and temperature difference between hot and cold water. N/A Factory default 30

43 3. Green Machine HMI screen: Options The GM Options parameters screen is shown in Fig. 7 and the parameters are explained in detail in Table 9. Fig. 7. GM Options parameters screen Table 9. GM Options parameters table with range and default values Parameter and Description Range Default Value CW FLOW SWITCH: Cold water flow switch installation N/A INSTALLED/NOT INSTALLED CW FLOW METER: Cold water flow meter installation N/A INSTALLED/NOT INSTALLED AIR/LIQUID COOLING: This describes the type of cooling system installed N/A LIQUID/AIR GRID PROT RELAY: Grid protection relay installation N/A INSTALLED/NOT INSTALLED REF BYPASS VALVE: Refrigerant bypass solenoid valve position i.e. OPEN when the GM is not running and N/A OPEN/CLOSE CLOSE when GM is running VFD MANUAL RUN: It is used to check the working of GM pump manually by pressing this tab N/A ON/OFF GENERATOR MANUAL RUN: It is used to check the working and direction of rotation of GM expander and N/A ON/OFF generator GEN NAMEPLATE RPM: This is generator rated RPM and is factory default to 1830RPM 1500RPM to 3700RPM 1830RPM ENTER ENCODER PPR 100ppr to 1024ppr 600PPR 31

44 4. Green Machine HMI screen: Machine Defaults, PLC I-O, and Veris Setup The following GM HMI screens are not generally used in GM setup or startup. These screens are for GM user reference purposes only. The Machine Defaults screen (Fig. 8) comes with all factory default settings. The other two screens PLC I-O (Fig. 9) and Veris Setup (Fig. 10) are the GM data acquisition screens. Fig. 8. GM Machine Defaults parameters screen 32

45 Fig. 9. GM PLC I-O parameters screen 33

46 Fig. 10. GM Veris Setup parameters screen Hot Water Loop Setup Parameters The hot water loop parameters to set up before starting the GM are as follows: 1. Check for leaks in the hot water system at thread-o-let connections, Gruvlok connections, elbows, joints, etc. 2. Check the pump rotation direction by turning on the pump for short period of time (2 seconds). 3. Check that the pumps inlet and outlet pressure are within the operating range of the system, i.e., that the pressure relief valve located at the pump inlet is at 45 psi. 4. Check the flow meter; make sure the flow meter is oriented in the correct direction, and verify that the flow meter is reading. 5. Check the pump VFD, i.e., as the VFD frequency changes, the pump responds to it by changing the flow rate. 6. Check that the required temperature is reached by opening the steam manual valve and operating the steam automatic valve using LABVIEW. 34

47 2.9.4 Cold Water Loop Setup Parameters The cold water loop parameters to set up before starting the GM are as follows: 1. Check for leaks in the cold water system at thread-o-let connections, butterfly valves, Gruvlok connections, elbows, joints, etc. 2. Check the pump rotation direction by turning on the pump for short period of time (2 seconds). 3. Check the flow meter; make sure the flow meter is oriented in the correct direction, and verify that the flow meter is reading. 4. Check the operation of bypass valves, bypass loop for the direction of cold water flow for cold water temperature control Operation Procedure The operation procedure explains the sequential steps in starting the GM; the main steps involved in this process are explained below. 1. Check that the cold water is flowing through the GM condenser. This can be done by looking at the GM HMI screen, which shows that the CW FL SW (cold water flow switch) is ON, with the green tab color indicating the cold water flow. 2. On the GM HMI screen, if MCR is shown OFF with gray tab color, then press the ENABLE button on the GM front panel (not on HMI screen) to bring the master control relay (MCR) into the ON position with green tab color. 3. With the cold water flow switch and MCR in the ON position, the GM HMI screen should show START button. 4. GM cannot be started directly at a hot water temperature of 220 F, as it will stop abruptly due to over-speed. For the reliability test, first set the hot water flow rate at 160 gpm (by adjusting the VFD frequency) and at 200 F (by operating steam valve using the LABVIEW). 5. For the reliability test, set the cold water flow rate at 160 gpm. 6. Now the GM should be ready for startup; press the START button on the GM HMI screen. 7. After running the GM for 15 minutes at 200 F of hot water supply temperature, increase the hot water temperature to 220 F for continuing with the reliability test Checklist The checklist is the sensors, gauges, joints, etc., located on the GM itself, hot water loop piping, and cold water loop piping, which need monitoring from time to time while the GM is in operation. The GM itself does not require any special inspection every day, as observed from the reliability test. Table 10 gives the list of sensors and their location, which could be monitored from time to time (on a weekly basis). 35

48 Table 10. Checklist for GM, hot water and cold water loops Sensor or Gauge Hot water and cold water flow rates Hot water supply temperature GM evaporator and condenser pressure drop on hot water and cold water loop side respectively. Hot water pump pressure boost Visual inspection of all joints Location Respective flow meters Data acquisition system (if manual thermometer is installed, than at hot water inlet to GM) Across GM Evaporator and Condenser Hot water pump inlet and outlet pressure Gruvlok couplings, elbows thread-o-lets etc., Reliability Test Results The reliability test consists of the operation of the GM at full load (50 kw) for 600 hours. During this phase of the test, important observations would include changes in GM performance for the long-term run at full load, noting GM shutdowns, reason for shutdown, and difficulty in solving this problem. Table 11 gives the GM gross and net power output during three different times of the reliability test. This table also provides details of the hot water and cold water supply conditions. Fig. 11 gives the GM HMI screen-shot at the 600 th hour of the reliability test. The reliability test was completed on the ORC power unit at full load (50 kw gross power output) for 600 hours to determine the long-term endurance and performance of the unit. Fig. 12 and Fig. 13 give the results of measured parameters from the test; they also show plots for some measured parameter results for more than 40 hours of the reliability test. The average hot water supply temperature was C and flow rate, m 3 /hr. Similarly, the average cold water conditions were 9.7 C and m 3 /hr. Table 11. Reliability test results at three different times of the test Date and time Hot water supply Cold water supply GM gross output GM net temperature and flow rate temperature and flow rate (kw) output (kw) 12/23/2011; 11:15:22 AM o F and 160gpm 51.4 o F and 160gpm /6/2012; 4:59:48 AM o F and 160gpm 59.1 o F and 160gpm /23/2012; 5:45:45 PM 218 o F and 160gpm 51.2 o F and 160gpm

49 Fig. 11. Green machine HMI screen-shot during reliability test operation 37

50 Fig. 12. Hot water and cold water supply temperatures to ORC power unit during reliability test 38

51 Fig. 13. Net power generated, power consumed by power unit pump and hot water pump during reliability test Green Machine Shutdown during the Reliability Test On January 3, 2012, during the reliability test, a GM automatic shutdown was observed. The shutdown occurred frequently (nearly 7 times), and the Alarms list of the HMI screen showed that the shutdown reason indicated was Expander high-pressure switch limit has been exceeded. When contacted, ElectraTherm (the GM manufacturer) noticed that the expander high pressure switch was malfunctioning. The expander high-pressure switch is one of the GM s many safety features that will turn off the GM when the high pressure of the working fluid exceeds the designated limit (see Fig. 5). To continue with the reliability test, we bypassed the expander high-pressure switch and started the GM again. The reliability test continued without any problem. The occurrence of Expander high-pressure switch failure is rare, and according to experts, such switches work for a lifetime without failure. However, ElectraTherm sent a new highpressure switch that was installed after all the tests at UAF were completed and the GM was ready to be shipped to the AP&T power plant in Tok for permanent installation. 39

52 2.10 Performance Testing and Results Performance Test The performance test was conducted mainly to learn the performance characteristics of the GM and its components (expander, evaporator, condenser, pump) under different operating conditions (i.e., hot water and cold water flow rates and temperatures). Another purpose for this test was to create performance characteristic charts of the GM at different heat source and heat sink conditions. The plan was to test the GM at 5 different hot water flow rates, 5 different hot water temperatures, 3 different cold water flow rates, and 2 different cold water temperatures. The list of hot water and cold water conditions at which the GM was tested is given in Table 12A. The total number of performance tests covered in the plan was 150. The procedure of performing this test (i.e., the method of changing temperatures and flow rate, data collection, etc.) is explained in the Operation Procedure section of this chapter. The preparation work and checklist for the performance test is similar to the reliability test (refer to the section Reliability Test ). Table 12A. Various hot water and cold water flow rates at which GM was tested Hot water temperatures ( o Hot water flow rate Cold water Cold water flow rate F) (gpm) temperatures ( o F) (gpm) Table 12B. Actual Input Conditions for GM Performance Testing Hot water temperatures ( o F) Hot water flow rate (gpm) Cold water temperatures ( o F) (for cold water flow rates of 120, 160, and 200 gpm) (for CW flow rate about 200 gpm) Operation Procedure For the cooling water at 50 F, the operating procedures are listed below: 1. The cold water flow rate was set at the desired value by turning the manual flow rate valve near the fire hydrant. The temperature of cold water from the fire hydrant was about 50 F year-round. 2. At this cold water flow rate, the desired hot water supply temperature to the ORC power unit was set by operating the steam flow control valve using LabVIEW software. 3. By varying the hot water pump VFD frequency (e.g. VFD frequency of 24 Hz corresponds to 27.2 m 3 /hr (120 gpm) and 55 Hz to 68.1 m 3 /hr (300 gpm) of hot water flow), the desired hot water flow rate was set. The hot water flow rate can be read in flow meter display in cubic-meter/hour. 40

53 4. After setting all four parameters (hot water and cold water flow rates and temperatures) at desired conditions, wait for approximately 30 minutes for steady-state condition for data collection. 5. Steady-state data collection was done for 30 minutes at each set of hot water and cold water temperatures and flow rates. This completes the performance test for one set of hot water and cold water flow rates and temperatures. 6. Change the hot water flow rate to the next value (e.g., 120 gpm to 160 gpm) by varying the VFD frequency, keeping the other three parameters the same. Then repeat Step 4 and Step 5. In this manner, the tests at other hot water flow rates were performed. 7. Change the hot water supply temperature using Step 2 and repeat Step 3, Step 4, and Step 5 for different hot water flow rates. 8. Step 2, Step 3, Step 4, and Step 5 were repeated iteratively for the three different cold water flow rates listed in Table 12. For the cooling water at 68 F, the cold water temperature was controlled using the bypass valves to allow a portion of warmer cooling water (exiting the GM) back into the GM. The rest of the operation procedure was similar to that for 50 F. However, a couple of difficulties in operation condition control for the case of cold water at 68 F were noticed, as described below: 1. For 68 F cases, due to the difficulty in obtaining stable cooling water flow rate to specific values (i.e., 120, 160, and 200) using the water hydrant and manually controlled bypass valves, the recorded cold water flow rate (for each hot water temperature and flow rate pair) was the one for which the flow rate and system performance were in stable condition (no fluctuation). Based on the measured ORC performance data of cold water at 50 F, the effect of flow rate on GM performance was observed to be not critically significant. For example, for cooling water at 50 F and hot water at 155 F and 120 gpm (for lowest heat energy available), the GM system efficiencies are 5.266, 5.557, and for cooling water flow rate at 120 gpm, 160 gpm, and 200 gpm, respectively. Another example, for hot water at 225 F and 300 gpm (for highest heat energy available) and cold water at 50 F, the GM system efficiencies are 7.278, 7.344, and for cooling water flow rate at 120 gpm, 160 gpm, and 200 gpm, respectively. Therefore, to use one cooling water flow rate of about between 160 gpm and 200 gpm to obtain system operation data for system performance analysis may give reasonable representative result for the system under normal operation conditions (i.e., not for extremely low or extremely high cold water flow rates). 2. Another shortcoming of the test for cooling water at 68 F is that, at the time of conducting the performance test, the maximum obtainable heating water temperature was 215 F (instead of 225 F, as it was during the test for cooling water at 50 F). This was due to the reduced pressure of the steam heat source, which in turn resulted from the reduced load requirement to the UAF (CHP) power plant while summer was approaching. Since the temperature range for this test has covered a large portion of the potential heat source temperatures for the targeted application (i.e., diesel engine waste heat), the obtained data were believed sufficient for the goal of this project. 41

54 System performance at different combinations of temperatures and flow rates may also be estimated using well-designed simulation models. In order to use the model for moderate extrapolation (i.e., for conditions not too much outside of the range covered by the operation conditions of this experiment), it is advisable to use a system simulation model based on performance models of individual components instead of the performance model of the whole system. A component model is a model obtained from physical laws of a particular component and the input and output data of the component. Due to the shortages listed in the previous paragraphs, the total number of test cases were 190 (refer to Table 12B). There were 75 cases for cooling water for 50 F and 20 cases for 68 F instead of 250 as planned (Table 12). As explained previously, the reduced number of cases may not critically affect the ability to accomplish the goal of this project Results The data listed in Tables 13A to 17B give the GM performance test results for cold water temperature of 50 F, and Tables 18A and 18B give the results for cold water temperature of 68 F. Table 13A. Performance results for HW Temp = 155 F; HW flow rate = 120 gpm to 300 gpm; CW Temp 50 F and CW flow rate = 120 gpm, 160 gpm, and 200 gpm # Hot water flow rate - Average (gpm) Hot water Supply Temp - Average (F) Hot water return Temp -Average (F) Cold water flow rate - Average (gpm) Cold water supply Temp - Average (F) Cold water return Temp - Average (F) GM (ORC) Net Power - Average (kw) GM (work fluid) Pump Power - Average (kw) Hot water pump power - Average (kw)

55 Table 13B. Induced performance results from measured readings of Table 13A # Hot water heat input (kw) Cold water heat rejected (kw) Cold water pump Power (kw) * System Operating Power (GM pump, CW pump) (kw) GM Gross Power (NO pumps) (kw) System operating efficiency (GM pump, CW pump) (%) GM net efficiency (GM pump) (%) GM gross efficiency (NO pumps) (%) Table 14A. Performance results for HW Temp = 175 F; HW flow rate = 120 gpm to 300 gpm; CW Temp 50 F and CW flow rate = 120 gpm, 160 gpm, and 200 gpm # Hot water flow rate - Average (gpm) Hot water Supply Temp - Average (F) Hot water return Temp - Average (F) Cold water flow rate - Average (gpm) Cold water supply Temp - Average (F) Cold water return Temp - Average (F) GM Net Power - Average (kw) GM Pump Power - Average (kw) Hot water pump power - Average (kw)

56 # Hot water flow rate - Average (gpm) Hot water Supply Temp - Average (F) Hot water return Temp - Average (F) Cold water flow rate - Average (gpm) Cold water supply Temp - Average (F) Cold water return Temp - Average (F) GM Net Power - Average (kw) GM Pump Power - Average (kw) Hot water pump power - Average (kw) Table 14B. Induced performance results from measured readings of Table 14A # Hot water heat input (kw) Cold water heat rejected (kw) Cold water pump Power (kw) * System Operating Power (GM pump, CW pump) (kw) GM Gross Power (NO pumps) (kw) System operating efficiency (GM pump, CW pump) (%) GM net efficiency (GM pump) (%) GM gross efficiency (NO pumps) (%) Table 15A. Performance results for HW Temp = 195 F; HW flow rate = 120 gpm to 300 gpm; CW Temp 50 F and CW flow rate = 120 gpm, 160 gpm, and 200 gpm # Hot water flow rate - Average (gpm) Hot water Supply Temp - Average (F) Hot water return Temp -Average (F) Cold water flow rate - Average (gpm) Cold water supply Temp - Average (F) Cold water return Temp - Average (F) GM Net Power - Average (kw) GM Pump Power - Average (kw) Hot water pump power - Average (kw)

57 # Hot water flow rate - Average (gpm) Hot water Supply Temp - Average (F) Hot water return Temp -Average (F) Cold water flow rate - Average (gpm) Cold water supply Temp - Average (F) Cold water return Temp - Average (F) GM Net Power - Average (kw) GM Pump Power - Average (kw) Hot water pump power - Average (kw) Table 15B. Induced performance results from measured readings of Table 15A # Hot water heat input (kw) Cold water heat rejected (kw) Cold water pump Power (kw) * System Operating Power (GM pump, CW pump) (kw) GM Gross Power (NO pumps) (kw) System operating efficiency (GM pump, CW pump) (%) GM net efficiency (GM pump) (%) GM gross efficiency (NO pumps) (%)

58 Table 16A. Performance results for HW Temp = 215 F; HW flow rate = 120 gpm to 300 gpm; CW Temp 50 F and CW flow rate = 120 gpm, 160 gpm, and 200 gpm # Hot water flow rate - Average (gpm) Hot water Supply Temp - Average (F) Hot water return Temp -Average (F) Cold water flow rate - Average (gpm) Cold water supply Temp - Average (F) Cold water return Temp - Average (F) GM Net Power - Average (kw) GM Pump Power - Average (kw) Hot water pump power - Average (kw) Table 16B. Induced performance results from measured readings of Table 16A # Hot water heat input (kw) Cold water heat rejected (kw) Cold water pump Power (kw) * System Operating Power (GM pump, CW pump) (kw) GM Gross Power (NO pumps) (kw) System operating efficiency (GM pump, CW pump) (%) GM net efficiency (GM pump) (%) GM gross efficiency (NO pumps) (%)

59 # Hot water heat input (kw) Cold water heat rejected (kw) Cold water pump Power (kw) * System Operating Power (GM pump, CW pump) (kw) GM Gross Power (NO pumps) (kw) System operating efficiency (GM pump, CW pump) (%) GM net efficiency (GM pump) (%) GM gross efficiency (NO pumps) (%) Table 17A. Performance results for HW Temp = 225 F; HW flow rate = 120 gpm to 300 gpm; CW Temp 50 F and CW flow rate = 120 gpm, 160 gpm, and 200 gpm # Hot water flow rate - Average (gpm) Hot water Supply Temp - Average (F) Hot water return Temp -Average (F) Cold water flow rate - Average (gpm) Cold water supply Temp - Average (F) Cold water return Temp - Average (F) GM Net Power - Average (kw) GM Pump Power - Average (kw) Hot water pump power - Average (kw) Table 17B. Induced performance results from measured readings of Table 17A # Hot water heat input (kw) Cold water heat rejected (kw) Cold water pump Power (kw) * System Operating Power (GM pump, CW pump) (kw) GM Gross Power (NO pumps) (kw) System operating efficiency (GM pump, CW pump) (%) GM net efficiency (GM pump) (%) GM gross efficiency (NO pumps) (%)

60 # Hot water heat input (kw) Cold water heat rejected (kw) Cold water pump Power (kw) * System Operating Power (GM pump, CW pump) (kw) GM Gross Power (NO pumps) (kw) System operating efficiency (GM pump, CW pump) (%) GM net efficiency (GM pump) (%) GM gross efficiency (NO pumps) (%) Table 18A. Performance results for HW Temp = 155 F to 220 F; HW flow rate = 120 gpm to 300 gpm; CW Temp 68 F and varying cold water flow rate # Hot water flow rate - Average (gpm) Hot water Supply Temp - Average (F) Hot water return Temp -Average (F) Cold water flow rate - Average (gpm) Cold water supply Temp - Average (F) Cold water return Temp - Average (F) GM Net Power - Average (kw) GM Pump Power - Average (kw) Hot water pump power - Average (kw)

61 Table 18B Induced performance results from measured readings of Table 18A # Hot water heat input (kw) Cold water heat rejected (kw) Cold water pump Power (kw) * System Operating Power (GM pump, CW pump) (kw) GM Gross Power (NO pumps) (kw) System operating efficiency (GM pump, CW pump) (%) GM net efficiency (GM pump) (%) GM gross efficiency (NO pumps) (%) Data Analysis, Performance Curves and Example Based on Developed Performance Curves In this section, the analysis results obtained from the experimental data are discussed. Data analysis is based on measured and deduced data from the tests and other information observed during the reliability and performance tests. The method applied to obtain the analysis results is based on procedures discussed in the section Data Reduction. The result of the analysis is performance characteristics of the ORC system and plots of performance curves. An example of applying performance curves to estimate the performance and economics of using the ORC with an Alaska rural village diesel generator, using engine jacket water as a heat source, is presented in later sections of this chapter Analysis Results from Reliability Test The reliability test was completed on the ORC power unit at rated load (50 kw gross power output) for 600 hours to know the long-term endurance and performance of the unit. Table 19 gives the results of measured and derived parameters from the test. As shown in Table 19 the 49

62 average hot water supply temperature was C and flow rate, m 3 /hr. Similarly, the average cold water conditions were 9.7 C and flow rate, m 3 /hr. Table 19 also gives the reliability test results for emissions reduction, CO 2 reductions (a greenhouse gas), and payback period for operating the ORC power unit at rated load for 363 days a year with two days for maintenance Performance from Reliability Test 1. No major problems were observed with the ORC power unit, such as drift in power output during long-term operations or power unit shutdowns, during the reliability test. 2. No major problems were observed with the steam loop, heating loop, cooling loop, and instrumentation; we were able to keep stable hot water and cold water flow rates and temperatures to the power unit. Occasionally hot water temperature was disturbed due to the surge in power plant steam supply conditions, i.e., change in steam supply pressure (note that the steam is a saturated steam). 3. During the reliability test, the average electrical power output ( ) by power unit was obtained at 47.8 kw with power unit efficiency ( ) of 7.8% and system operating efficiency ( ) of 7.5%. Here in calculating, the cold water pump power consumed was taken to be 1.76 kw, the same as hot water pump power consumption. 4. As noted in Table 19, the ORC power unit has a maximum heat input limit of around 610 kw (due to the output limitation of 50 kw applied to the expander). If excess heat is available (i.e., more than 610 kw), the extra heat should be used for other purposes such as building heating. For a varying heat source like jacket water from a rural village generator, a process based on the heat availability of village diesel power plant to optimize the benefits achievable is recommended. The data obtained from this test may provide the rural diesel plant personnel with information on how to distribute the waste heat for optimal benefit. 5. The ORC power unit, which achieved an efficiency ( ) of 8.4% at full load operation, was well within the manufacturer s claim of 8.5%. 6. A payback period of 2.1 years and 2.4 years was obtained with a 0% and 10% interest rate on capital, respectively. 7. To operate the system, an operator does not need an advanced technology background. The only thing that needs to be remembered for operating the GM is to keep the heat source temperature below 230 F to start the GM. Designing the heating and cooling loops with the capability of heating and cooling process control for startup is recommended. This will make startup just as simple as pushing the Start button. The design of a heating and cooling process control is standard practice. 8. The maintenance requirements of the system do not need an advanced technology background, and the GM maintenance process can be combined with the routine diesel generator maintenance schedule. Since virtually no defects were found during the reliability test, the maintenance requirements were expected to be as simple as following the items listed in the Maintenance Items list recommended by the manufacturer. The routine maintenance schedule includes quarterly, semi-annual, annual, and bi-annual maintenance items. Quarterly maintenance items are basically inspections (visually inspect plumbing for leakage oil/water and brackets, connections, 50

63 bolts, conduit, wire). Semi-annual (replace filter/dryer element), annual (check heat exchanger pressure drop in hot and cold water supply), and bi-annual (replace pressure relief valves). Maintenance items could be combined with the routine maintenance schedule of the diesel generator and carried out by the diesel power plant engineers. 9. The technology is feasible for rural Alaska villages (the ease in installation, operation, and maintenance requirements). The performance data obtained from the reliability test show that the ORC system can consistently generate a gross power of 50 kw (rated power of the GM) and a net output power of 47.8 kw (difference between the gross power and the required working fluid pump power of the ORC system) under the conditions of a sufficient heating source and a sufficient cooling source. Based on these results, the potential of the GM in terms of emission reduction, CO 2 reduction, fuel savings, and payback period are evaluated and listed in Table 19. Table 19. Reliability test results Parameter Value Average hot water supply temperature to power unit ( ) o C (219.7 o F) Average hot water flow rate to power unit ( ) 36.28m 3 /hr (159.8gpm) Average cold water supply temperature to power unit ( ) 9.7 o C (49.4 o F) Average cold water flow rate to power unit ( ) 37.15m 3 /hr (163.6gpm) Power unit electrical power output ( ) 47.8kW Power unit pump power consumption ( ) 3.61kW Hot water pump power consumption ( ) 1.76kW Cold water pump power consumption ( ) 1.76kW* System operating power output ( ) 46.04kW** Heat supply by hot water to power unit evaporator ( ) 610.4kW Screw expander efficiency ( ) 8.4% Power unit efficiency ( ) 7.8% System operating efficiency ( ) 7.5% Diesel fuel saved per year ( ) gal Dollar amount saved on diesel fuel per year ( ) $ year Emissions reductions*** Oxides of nitrogen (NO X) lb/year Hydrocarbons (HC) 351.4lb/year Particulate matter (PM) 87.9lb/year Carbon monoxide (CO) lb/year Carbon dioxide (CO 2) 316tons/year Payback period Payback 0% interest on capital 2.1years Payback 10% interest on capital 2.4years *Cold water was from hydrant, pump power consumption (P pump, CW ) data was not measurable for cold water flow. Cold water pump power consumption for a given flow rate was estimated using the measured hot water pump power consumption P pump, HW at the same flow rate. **P op = P Net P pump, CW. Hot water pump power is not included in this calculation. This is based on the assumption that the ORC is applied to recover diesel generator jacket waste heat and no extra power (besides the diesel engine jacket water pump power) is needed to keep jacket water flowing through the ORC. ***The potential emissions reduction is based on Tier 4 regulations. CO 2 reduction is based on a CO 2 generation rate of about 22 lb/per gallon of diesel fuel. 51

64 Analysis Results from the Performance Test Planned operating conditions for the performance test on the 50 kw ORC power unit are listed in Table 12; the actual operating conditions for testing are given in Table 12B. The reduced number of test cases did not critically affect the importance and applicability of the experiment to achieve the goals of this project; the reasons are described in the section Operation Procedure. Based on the data given in Table 13A to Table 18B, Fig. 14 to Fig. 22 were plotted. Fig. 14 to Fig. 16 give the performance test results for cold water supply temperature of 50 o F (10 o C) and varying other three parameters. Fig. 14 to Fig. 16 were plotted based on the measured average values of hot water and cold water supply temperatures and flow rates and the power unit electrical power output ( ), electrical power consumption by power unit pump ( ) and cold water pump ( ). The average values are the average obtained from 30-minute sampled data after the system reached steady-state condition. Temperatures were sampled at a frequency of 1 second, electrical power data were sampled at a frequency of 30 seconds, and flow rate data were noted manually from the flow meter display screen. Fig. 14 gives the heat input to the evaporator of the power unit for 5 hot water flow rates, 5 hot water supply temperatures, 3 cold water flow rates, and cold water supply temperature of 50 o F (10 o C). Fig. 15 and Fig. 16 are the plots for heat rejected by the working fluid to cold water in the condenser and system operating power output, respectively, for 5 hot water flow rates, 5 hot water supply temperatures, 3 cold water flow rates, and cold water supply temperature of 50 o F (10 o C). Similar types of curves (similar to Fig. 14 to Fig. 16) are obtainable for cold water supply temperature of 68 o F (20 o C) and are not given here, but the charts corresponding to deduced data that can be used for performance analysis are included in this chapter.. Fig. 17 to Fig. 22 show the performance curves of the ORC power unit deduced from measured data of the performance test. Each figure corresponds to a case of a specific cold water temperature and flow rate. The curves in each figure correspond to the operation conditions of 5 hot water temperatures and 5 hot water flow rates Discussion Based on Performance Test Results For a given hot water supply temperature and cold water flow rate, in general, heat supplied by hot water ( ) to the power unit evaporator increased with the increase of hot water flow rate, as shown in Fig. 14. For example, at hot water supply temperature of 79.4 o C (175 o F) and cold water flow rate of m 3 /hr (160 gpm), heat supplied by hot water increased from kw at m 3 /hr (120 gpm) of hot water flow rate to kw at m 3 /hr (300 gpm) of hot water flow rate. However, for some cases, an irregular trend occurred due to an occasional disturbance in hot water supply temperature that resulted from a surge in power plant steam supply condition. The corresponding irregular trend was also observed from heat rejection and power output curves in Fig. 15 and Fig. 16, which result from disturbance occurring to the hot water supply temperature. As the hot water flow rate increased for a given hot water supply temperature, the heat input to power unit reached asymptotic condition (Fig. 14); i.e., for a given hot water supply 52

65 temperature, the heat absorption by working fluid in the evaporator reached a limiting value for higher hot water flow rates. The same trends were observed for system operating power output as it reached asymptotic condition for higher hot water flow rates (Fig. 16). The reason for this asymptotic condition is that the ORC power unit evaporator reached its design capacity. There is another limitation in the ORC unit PLC software that prevents the screw expander from generating more than the rated load of 50 kw. The PLC software limitation, which limits the R- 245fa flow entering the screw expander, is one of the many safety features that protect the screw expander from over-speeding. Observe in Fig. 14 to Fig. 16 that for a given hot water flow rate and hot water supply temperature the effect of cold water flow rate on ORC system operation power output and ORC system operation efficiency is minimal within the cold water flow rate range used in the experiment. For example, for cooling water at 50 F and hot water at 155 o F and 120 gpm (for lowest heat energy available in this experiment), the GM system operation efficiencies are 5.266, 5.557, and for cooling water flow rate at 120 gpm, 160 gpm, and 200 gpm respectively. Another example, for hot water at 225 F and 300 gpm (for highest heat energy available) and cold water at 50 F, the GM net efficiencies are 7.278, 7.344, and for cooling water flow rate at 120 gpm, 160 gpm, and 200 gpm, respectively. Another example for the system operation power output for hot water flow rate of 45.4 m 3 /h (200 gpm) and supply temperature of 90.5 o C (195 o F) and cold water temperature of 50 o F, the system operating power output is kw, kw and kw for cold water flow rates of 27.2 m 3 /h (120 gpm), m 3 /h (160 gpm) and 45.4 m 3 /h (200 gpm), respectively. It was also observed that asymptotes for system operation output and efficiency versus heat source (hot water) flow rate curves exist and the values depend on the heat source temperatures. For example, for cold water of 50 o F and hot water of 195 o F, the system operation output has the asymptote of about 35 kw and system operation efficiency of about 7.4%. Fig. 17 to Fig. 22 give performance curves of the heat supplied by hot water, heat rejected to cold water, system operating power output, efficiency, payback period and reductions in CO 2 emissions versus different hot water supply temperatures, and hot water flow rates for each of the cold water supply temperatures of 50 o F (10 o C) and 68 o F (20 o C). In each of the plots (Fig. 17 to Fig. 22), the top plot is for 50 o F (10 o C) cold water temperatures and the bottom plot is for 68 o F (20 o C) cold water temperatures. All six plots are presented on the same hot water supply temperature scale and with the same color coding for ease of reading. For example, at hot water supply temperature of o C (215 o F) and hot water flow rate of 45.4 m 3 /h (200 gpm), from Fig. 17, the hot water heat input to the evaporator is kw and kw for 50 o F and 68 o F cold water temperature, respectively; from Fig. 18, heat rejection to cold water is kw and kw for 50 o F and 68 o Fcold water temperature, respectively; from Fig. 19, the system operating power output is kw and 41.8 kw for 50 o F and 68 o F cold water temperature, respectively; from Fig. 20, the system operating efficiency is 7.3% and 7.1% for 50 o F and 68 o F cold water temperature respectively; from Fig. 21, the payback period of 2.5 years and 2.7 years for 10% interest rate on capital could be achieved for 50 o F and 68 o F cold water temperature, respectively; and from Fig. 22, CO 2 reductions of short-tons/year and 53

66 288.6 short-tons/year for 50 o F and 68 o F cold water temperature, respectively, could be achieved. Payback periods and CO 2 reductions were calculated based on equations and procedures discussed in the section Data Reduction. The combined effect of heating source temperature and cooling source temperature was observed. For hot water at 215 o F, the ORC screw expander power outputs are kw and kw for cooling water of 50 o F and 68 o F, respectively and the corresponding ORC net efficiencies are 7.515% and 7.413%. For hot water temperature of 155 F, the ORC screw expander power outputs are kw and kw for cold water of 50 o F and 68 o F, respectively, and the corresponding ORC net efficiencies are 5.911% and 4.423%. The results show that cold water temperature has more effect on the ORC performance when hot water temperature is low. This finding matches the thermodynamic principle that higher temperature difference across a thermal engine normally means more efficiency. In Fig. 17 to Fig. 22, for cold water temperature of 68 o F, the results were presented only up to the maximum hot water supply temperature of o C (215 o F), because the low saturated steam pressure in the power plant prevented the hot water supply temperature from reaching the expected maximum of o C (225 o F) during the test. 54

67 Fig. 14. Heat input to power unit evaporator vs. hot water flow rates at different hot water supply temperatures and cold water flow rates 55

68 Fig. 15. Heat rejected to cold water in power unit condenser vs. hot water flow rates at different hot water supply temperatures and cold water flow rates 56

69 Fig. 16. System operating power output vs. hot water flow rates at different hot water supply temperatures and cold water flow rates 57

70 Fig. 17. Heat input vs. hot water supply temperature 58

71 Fig. 18. Heat rejected vs. hot water supply temperature 59

72 Fig. 19. System operating power output vs. hot water supply temperature 60

73 Fig. 20. System operating efficiency vs. hot water supply temperature 61

74 Fig. 21. Payback period vs. hot water supply temperature 62

75 Fig. 22. CO 2 reductions vs. hot water supply temperature 63

76 Example Based on the Performance Curves From the Power Cost Equalization (PCE) program data published by the Alaska Energy Authority for fiscal year 2011 and based on available diesel engine data at the location, Tok, Alaska, was selected for evaluating diesel engine waste heat recovery for power generation using the present ORC system. Based on PCE data, Tok s annual electrical load is 10,902,597 kwh, and all of this power is generated using an isolated Caterpillar 2 MW diesel engine. The specifications of the engine are given in Table 20. Table 21 gives the diesel engine power output, specific fuel consumption, exhaust temperature, heat rejected by engine to jacket water, and exhaust at different loads of the engine. Note that the heat present in exhaust is based on a lower heating value of exhaust, that is, cooling the exhaust to only o C (350 o F) to avoid acid formation in the exhaust manifold. From the annual electrical load consumption, the average electrical load on the diesel engine is 1250 kw ( hp). Considering 1.3 MW (1700 hp) as average load on a diesel engine, the average percent load on the diesel engine is 65.7%. By interpolation with between 50% and 75% for 65.7%, Table 21 also gives the diesel engine data at 65.7% load. For evaluating the ORC performance for waste heat recovery from stationary diesel engines, two cases were simulated: the jacket water heat recovery system only, and the combined jacket water and exhaust heat recovery. For both simulation cases, it was assumed that a water cooling source as heat sink is readily available at 50 o F (10 o C), which is about or above the yearround ground water temperature in Tok, with flow rate ranging from 27.2 m 3 /h (120 gpm) to 45.4 m 3 /h (200gpm). Table 20 gives the engine jacket water temperature at 99 o C (210.2 o F). Assuming that 200 gpm jacket water is bypassed to be supplied as a heat source for the ORC power unit, Table 22 gives the results for operating this ORC power system on waste heat from jacket water of the Tok generator. Observe in Table 21 that the exhaust temperature at 65.7% engine load is 402 o C, which is well above the o C (225 o F) required for the ORC to generate maximum system operating power. For the simulated case of combined jacket water and exhaust heat recovery system, if the heat recovery system is designed such that the jacket water from the engine is first passed through the exhaust heat exchanger, it is possible to achieve o C (225 o F) as hot water supply temperature for the ORC power unit evaporator. Table 22 also gives the ORC power unit performance for both jacket water, and combined jacket water and exhaust heat recovery system installed together. In Table 22, the heat input to power unit, system operating power output, efficiency, payback period, and CO 2 reductions can be obtained from Fig. 17 to Fig. 22. For combined jacket water and exhaust heat recovery, the system operating power output is 45.7 kw with a payback period of 2.4 years. Considerable reductions in emissions could be achieved, as listed in Table 22, which were calculated based on the EPA Tier 4 interim reduction standards discussed earlier. 64

77 Table 20. Diesel engine specifications Diesel Engine Caterpillar C Brake power 1.9MW (2588BHP) Number of cylinders 16 Compression ratio 16.7 Speed 1200rpm Jacket water temperature 99 o C (210.2 o F) Jacket water flow rate 120m 3 /h (528gpm) Aspiration Turbocharged (no EGR) Percent load Brake power, hp (MW) Table 21. Diesel engine specifications Diesel fuel consumption, lit/s (gpm) Heat rejection to jacket water, kw Exhaust temperature, o C ( o F) Heat present in exhaust at o C (350 o F), kw (1.9) (2.08) (784.04) (1.4) (1.61) (759.38) (1.0) (1.12) (749.48) (0.5) (0.63) (644.72) (1.3) (1.42) (755.6) Table 22. Estimated ORC performance for operating on waste heat recovery from diesel engine Parameter Jacket water heat only Jacket water + Exhaust heat* Hot water supply temperature to ORC power unit 99 o C (210.2 o F) o C (225 o F) Hot water flow rate to ORC power unit 45.4m 3 /h (200gpm) 45.4m 3 /h (200gpm) Heat input to evaporator of ORC power unit 576.5kW 617.7kW System operating power output 41.7kW (363.77MWh/year) 45.7kW (398.5MWh/year) System operating efficiency 7.2% 7.4% Diesel fuel saved gal/year gal/year Dollar amount saved on diesel fuel $ /year $ /year Payback 0% interest 2.3years 2.1years Payback 10% interest 2.7years 2.4years Reductions in CO 2 emissions 288.4short-tons/year 316short-tons/year Reductions in NO X emissions 2807lb/year 3075lb/year Reductions in HC emissions 320.8lb/year 351.4lb/year Reductions in CO emissions 2807lb/year 3075lb/year Reductions in PM emissions 80.2lb/year 88lb/year *225 F is the highest hot water temperature for testing; 250 F is the highest hot water temperatures allowed to the ORC system Further Discussion of Adopting the GM for a Rural Genset To estimate if a village diesel genset is appropriate for adopting the GM, payback period may be one of the most important parameters. Fig. 21 presents curves of payback periods versus hot water temperature and flow rate and cold water temperature. In constructing those curves, capital cost and parasitic power (for this case, cooling system power consumption) were considered constants (capital cost of $280,000 and parasitic power of 1.8 kw). Since capital cost and parasitic power consumption may depend on the condition of the individual villages, to construct a general-purpose payback period chart that is useful for all rural diesel gensets, the capital cost and cooling system power consumption should also be treated as variables. More independent variables (i.e., fuel cost) may make the payback period chart more useful. In order to do so, more sets of a similar chart could be made available for each of the different fuel 65

78 costs. In this chapter, only one set of charts for a fuel cost of $5/gallon is presented for demonstration purposes. For a different fuel cost, a similar set of charts can be produced following the same procedure. Since capital and fuel cost are case- or village-dependent, a payback period chart, for which capital and fuel cost are treated as variables, may be more useful than a chart made for specific capital and fuel cost. Fig. 23 shows a set of charts with varying capital and fuel costs for payback period calculations. To use the charts, one needs to have information on jacket water temperature and flow rate and cooling water temperature available; the charts can then be used to estimate the payback period under whatever the case of capital and fuel costs. These curves may help determine if the GM could be beneficial to a specific village, for which information on diesel generator load pattern and fuel cost is known. Fig. 23A. Pay period at 0% interest rate on capital for different Green Machine ORC power outputs, fuel prices, and capital costs 66

79 Fig. 23B. Pay period at 10% interest rate on capital for different Green Machine ORC power outputs, fuel prices, and capital costs 2.12 Discussion (Economics, Emissions, Findings) This section summarizes some of the findings. This section is divided into four parts. The first part summarizes general findings during the process of the experiment. The second part relates to the performance characteristics of the ORC system. The third part is about how to match an ORC to a rural village diesel generator set, and the fourth part recommends how the information obtained could be translated into a general policy for applying the ORC system to rural Alaska villages. A conclusion is included to address how the controlled test procedure can be improved Findings of General Information Related to the GM ORC Testing System Design and Installation: In 2010, it was found that only one ORC system, which was in the stage of mixed testing and commercialization, was available to test for a rated heat source equal to or less than 800 kw. 67

80 To select a testing site, the available space and details related to the availability of utilities/heating/cooling sources are important for testing system design and installation. Equally important is information about the local codes that must be followed, which may significantly affect the project cost. Based on data gathered during the procurement, for economic consideration of the ORC system, the installation and operation costs of the cooling system (which is not the property of the diesel generator set) may become one of the dominate factors for ORC application. The costs of the cooling system may depend on the availability of the type and capacity of the cooling source at the test site. A good preliminary theoretical model (for this case, the model included both heat transfer and thermodynamics theories) was found useful in designing a test plan and test system (heat source and heat sink temperature and flow rate ranges), selecting appropriate piping components, determining parameters needed for system characteristic analysis, and choosing appropriate measurement devices. For useful simulation results, specifications of the ORC system, heat source, and cooling source were needed too. For an isolated small city, a careful purchase plan is needed to avoid long schedule delays. For this project, the installation and instrumentation process (steam loop, hot water loop, cold water loop, electrical circuit, signal/monitoring/control circuits, data acquisition system) was not difficult. In other words, the installation of the GM, in general, does not need to be a complicated procedure if a heat source and cooling source infrastructure are readily available. Design and Operation Shortfalls of the GM During the commissioning, a couple of design issues were discussed: 1. The selected low-temperature ORC heat engine (or GM) could be converted from a 50 kw capacity into a 65 kw capacity (since almost all the components are designed for a load of 65 kw). 2. A system computer program imperfection to cause emergency shutdown of the ORC system: The expected shutdown resulted from trying to start the system under undesired heating flow condition. It may not cause any damage to the system, but is undesirable. This problem may be easily overcome by following correct starting procedures or adding control capability into the installed heating loop. Currently, there is an updated 65 kw ORC unit from the same manufacturer that does not have the design and operation shortfalls mentioned above. Some of the system limitations: 1. The GM has a maximum power output of 50 kw, so it has a limitation on the conversion of heat into power. Excessive heat is advisable to be distributed for other usage or it may simply be wasted. 68

81 2. The GM may shut down if the voltage and frequency fluctuations are high for the location where it is installed (voltage fluctuations greater than ±5 V and frequency fluctuations greater than 10%). 3. The GM starts only when the temperature difference between the hot and cold side is greater than the preset value. The preset value can be adjusted between 50 o F and 80 o F. 4. The GM initial (i.e., when we start the machine) power output is set by a predefined equation, which is a function of temperature difference between the heating side and cooling side (ΔT H-C ). If the initial temperature difference (ΔT H-C ) is greater than 170 o F, then the predicted starting power output of the GM is greater than 50 kw and the machine is expected to over-speed. The machine shuts down as a consequence. So the initial ΔT H-C should not be too high for the machine to start smoothly. 5. The GM has a hot water supply temperature limitation of 250 o F. If the temperature is above this value, the GM shuts down. 6. The GM has a minimum net power output limit of 5 kw. For output less than 5 kw, the machine shuts down. 7. The components of this ORC system are capable of 65 kw output. The system may be modified to increase its power capacity from 50 kw to 65 kw Findings in GM Performance and Comments on Applications Heating Source: The ORC net efficiency does not vary drastically across a wide range of input heat (combination of hot water temperature and flow rate). For example, for a heat source of 627 kw (the highest output for this experiment) and 225 o F and cooling source (i.e., heat sink) of 50 o F, the GM ORC operates with rated outputs of about 50 kw and the ORC net efficiency is 7.6%. For a heat source of 257 kw (the lowest output) and 155 o F and cooling source of 50 o F, the GM ORC net efficiency is 5.266%. At a given heat source temperature, the system efficiency does not vary much against heat source flow rate. For example, for a heat source at 225 o F (highest heat source temperature) and cooling source at 50 o F and 200 gpm, the ORC net efficiencies are 7.531%, 7.613%, 7.554%, 7.468%, and 7.389%, respectively, to the heating source flow rates of 120 gpm, 160 gpm, 200 gpm, 250 gpm, and 300 gpm. Another example is the heat source at 155 o F (lowest heat source temperature) and cooling source at 50 o F and 200 gpm. For this case, the net efficiencies are 5.896%, 5.911%, 5.904%, 5.949%, and 5.985% for the respective flow rates of 120 gpm, 160 gpm, 200 gpm, 250 gpm, and 300 gpm. In other words, the system can be adopted for a situation of diesel generators with relatively large heat temperature variations. The performance properties mentioned in the previous two paragraphs may make the GM ORC applicable to a wide size range of diesel engines and suitable for diesel engines that operate under a wide range of varying load and jacket water temperature. This information may also help with estimating if the heat source condition of an individual diesel generator is appropriate for applying the 50 kw GE ORC for the desired economic benefit. As observed from test data, the GM may also have asymptotes in heat absorption, heat dissipation, and GM ORC net power output versus heating fluid flow rate for a specific heat 69

82 source temperature. An asymptote with respect to the flow rate means, at a given heat source temperature, further increase of the flow rate may not effectively increase the heat absorption, heat dissipation, and GM ORC net output power accordingly. In general, asymptotes may largely depend on the heat exchanger design and power output limitation (50 kw) of the ORC screw expander set by the manufacturer (i.e., if a heat source is more than needed for 50 kw of expander output, part of the heat energy will bypass the expander and not be used for power generation). For a given village diesel generator set, the asymptotes information can be useful in determining appropriate waste heat flow distribution between a GM and other co-existing heat recovery devices (e.g., for heating, other heat for power systems) to avoid over feeding the ORC and ineffective use of the heat energy. For the GM ORC the asymptotes in heating source temperature seem nearly linearly dependent on the heat source temperature for the temperature range practiced in this experiment (155 F to 225 F). Allowable maximum and minimum heat source temperature settings by the manufacturer for the GM starting process are 150 F and 250 F, respectively. The high and low heat source temperatures of this testing seem close enough to the temperature limits of the GM ORC. It is thought that incorporating a proper system modeling scheme and the experimental data performance of the GM ORC may be predicted with reasonable reliability. Further Discussion: According to this experiment, of which the heat source (e.g., hot water, hot fluid) temperature ranges from 155 o F to 225 o F and flow rate ranges from 120 gpm to 300 gp, the temperature of hot water exiting the GM ORC ranges from about 140 o F to 205 o, of which plenty heat energy still exists in the exiting hot fluid (i.e., residual heat). Furthermore, if the heat source temperature is 250 o F, the residual heat temperature may be even higher. For these cases, the residual heat could be further used, if other coexisting heat recovery facilities are available. When applying the residual heat, caution needs to be practiced, if jacket water is involved. For example, if the return hot water is going to enter the engine jacket, the preferred minimum temperature setup by the manufacturer needs to be considered. In general, the temperature difference between the jacket water across the engine jacket (e.g., 15 F) can be found in the diesel generator specifications. If the heat source is from exhaust (most likely there is need of an exhaust to liquid heat exchanger) only (i.e., a heating loop without involving engine jacket water), all (if possible) the residual heat energy in the hot water may be applied for further heat recovery. Cooling Source: There is no requirement to the minimum cooling source temperature (besides freezing point of the cooling fluid) for the GM ORC. The maximum cooling water temperature is 110 o F, set up by the manufacturer. This paragraph may duplicate a portion of a previous section. The purpose of duplication is mainly to repeat that a limited number on cold water temperatures has been adopted for the GM test. The cold water (heat sink) was from a fire hydrant, and the temperature of the water was nearly constant (around 50 o F) for all test cases. The temperature of the cooling water entering the GM can be controlled to higher than 50 o F, but not lower. For this project, the 70

83 effect of cooling water on GM performance has been investigated for two cooling water temperatures: 50 o F and 68 o F. The three cold water flow rates used in the test were 120 gpm, 160 gpm, and 200 gpm for cooling water at 50 o F. The three flow rates used in the tests were selected based on how crucial the pump power requirements were relative to the power generated by the GM. For example, power requirements for the pump running at 120 gpm, 160 gpm, 200 gpm, 250 gpm, and 300 gpm are 1.02 kw, kw, kw, kw, and kw, respectively. Pump power consumption seems an exponential function of flow rate. Pump power consumption at 1 kw (for a flow rate of 120 gpm) may be considered not critical based on GM gross output data (from about 11 kw to about 50 kw) obtained from this test, which covers the expected gross power output range of practical GM applications to rural diesel generators. The effect of cooling water (CW) flow rate on efficiency was observed as minimal. Fox example, for heating water at 155 F and 120 gpm (the lowest heat energy available for this test) and cooling water at 50 F, the GM ORC net efficiencies are 5.266%, 5.557%, and 5.896% for CW flow at 120 gpm, 160 gpm, and 200 gpm respectively. Another example for heating water at 225 F and 300 gpm (the highest heat energy available) and cooling water at 50 F, the GM ORC net efficiencies are 7.278%, 7.344%, and 7.389% for CW flow at 120 gpm, 160 gpm, and 200 gpm, respectively. This (the effect of cooling water flow rate on efficiency was observed minimal) may also result from all selected cooling water flow rates being more than required for economically beneficial operation of the GM. More drastic decrease in efficiency, which is not desirable for practical applications, is expected at some flow rates lower than 120 gpm (obviously, GM may not work at 0 gpm cooling water flow rate). Considering the ORC operation efficiency on diesel generators, the cooling water pump power (parasitic power) needs to be considered, which may make the higher than necessary cooling water flow rate (much higher parasitic power) less attractive. Optimization in operation efficiency and economic return is recommended, if feasible, for determining desirable cooling water condition. The combined effect of heat source temperature and cooling source temperature was observed. For hot water (HW) at 215 o F, the ORC screw expander power outputs are kw and kw for cooling water of 50 o F and 68 o F, respectively, and the corresponding ORC net efficiencies are 7.515% and 7.413%. For hot water temperature of 155 F, the ORC screw expander power outputs are kw and kw for cold water of 50 o F and 68 o F, respectively, and the corresponding ORC net efficiencies are 5.911% and 4.423%. The results show that cold water temperature affects the ORC performance more when hot water temperature is low. This finding matches the fundamental thermodynamic principle. Considering applications in rural Alaska, jacket water temperatures are normally between 175 o F and 195 o F (based on log data of a few village diesel generator sets). As mentioned previously, for 120 gpm of cooling water flow, the estimated pump power (parasitic power) requirement is around 1 kw, which is low in comparison with the level of GM ORC output for jacket water heat application (between 20 kw and 40 kw) in general. Therefore, use of 1 kw (or appropriate extrapolation for parasitic power estimate) for system performance and economic outcome evaluations for cases that cold flow rate less than 120 gpm may not cause significant error. In other words, for cases of villages that can only provide a 71

84 cooling water flow rate much less than 120 gpm (e.g., 80 gpm or less) or for cases that do not need 120 gpm for cooling due to the availability of very low cold water temperature (e.g., 40 o F deep well water), data obtained from this experiment may be used through extrapolation with caution. One way to use data obtained from this experiment to get a reasonable prediction of system performance for low temperature and flow rate of cold water (much less than 50 o F and 120 gpm) is to apply a good simulation model, of which the components are modeled using reliable data (for the GM model: using data obtained from this test) and model formulations are based on correct and physically meaningful parameter relationships. However, it is always advisable to use performance data directly obtained from lower temperature and flow rate measurements, if measurements and data analyses are feasible physically and economically. This subsection discusses findings in payback period based on test data of this experiment. Based on reliability test results, for full rated output (gross output of 50 kw and net output of 47.8 kw), the estimated payback period is 2.1 years for a 0% interest rate and 2.4 years for a 10% interest rate, with the total installation cost of $280,500, fuel cost of $5/gallon, and operation cost of $7600/year. Details of the economic analysis and reductions in emissions and GHG can be found in Table 19. Performance curves were plotted (Fig. 17 to Fig. 22) for heat input to evaporator, heat rejected to cold water, system operating power output, efficiency, payback period and emissions, and CO 2 reductions with respect to hot water supply temperatures between 155 o F and 225 o F and cold water temperatures of 50 o F and 68 o F, respectively. Even for the lowest hot water supply temperature (155 o F), the payback period of 7.5 years and 10 years could be achieved for 50 o F and 68 o F cold water temperatures, respectively. Reductions in emissions and CO 2 can be found in Fig. 22. One reminder is that the ORC system operation power output and efficiency in this report is based on the assumption that the heat source comes from engine jacket water, for which the operation power output and efficiency calculations only consider power consumed by the cooling water pump and working fluid pump as parasitic power. Heating water pump power consumption is not considered parasitic power due to the assumption that the addition of the heat recovery loop may only cause insignificant change in the power consumption of the engine jacket water pump. If the heat source is from exhaust heat and the heating loop includes an exhaust-to-liquid heat exchanger and associated components, heating fluid pump power may need to be considered parasitic power for performance and economic analyses. An additional reminder is that when another type of heat source (e.g., exhaust) is used, the residual heat in the hot water exiting the GM may be further recoverable to an extensive level Match between the GM ORC and a Diesel Generator Set The determination of whether a specific village generator will gain desired economic benefit from applying the GM ORC system is discussed in this subsection. The major factors of concern include type (exhaust and/or jacket water) and condition of heat source (year-round temperatures and flow rates; availability of heat energy for GM operation due to co-existence of and potential application of other heat recovery systems), type and condition of cooling source (surface water, underground water, cooling tower, radiator, convention heat 72

85 exchanger), willingness of the village to adopt the GM heat recovery system, installation and infrastructure costs for GM, operation cost of GM, and interest rate. Due to the hot water temperature limits to the GM (150 o F to 250 o F), the system seems not to be able to take full advantage of very hot water temperature (higher than 250 o F) from using diesel engine exhaust heat. Therefore, the heat source (e.g., hot water) may come from jacket water alone or combined jacket water and exhaust for a heat source temperature not over 250 o F. A proposed procedure, which may not be the best and/or the only choice, is presented below to match the performance between the GM and any individual diesel generator sets: 1. Estimate fuel prices, capital costs, interest rate, and desired payback period. 2. Use Fig. 23 to determine the minimum requirement of operation power output of the GM that matches the desired payback period. 3. Estimate the availability of conditions (flow rates and temperatures) of hot water and cold water of the diesel generator set, which will be available to the GM application. The conditions may be as detailed as hourly data or as rough as yearly average data. More details may give a more accurate prediction. 4. Estimate the parasitic power for operating the GM heat-to-power system. (Sources of parasitic power may depend on the heating and cooling sources used for operation; please refer to the previous section.) 5. Use Fig. 16 (or Table 13 to Table 18) to estimate the net output of the GM based on local hot water and cold water conditions available to the GM. More detailed data may take more time to process and are expected to generate better results. If necessary, appropriate extrapolation could be applied. 6. If estimated GM system operation output (which is the GM net output minus the parasitic power) equals or is more than the minimum required operation power output for the desired payback period, the GM is recommended for the particular village power plant. If the amount of heat is much more than the minimum requirement for the payback period, a further study of how to more effectively use the waste heat (e.g., add more GM, add a heat recovery system for heating, select a different size of heat-topower application if available.) may be needed. 7. If estimated GM system operation output is less than the minimum required operation power output for desired payback period, use modified and achievable combinations of hot water and cold water conditions (same amount of heat may exist in fluid of different temperature and flow rate conditions) and repeat Steps 4 to 6 to determine the match. Practically, a few iterations of combinations of hot and cooling water conditions may help reach the final decision. 73

86 The following is an example that demonstrates the proposed procedure: Suppose a power plant has a hot jacket water flow of 125 gpm and 185 F and a cooling water source of 120 gpm and 50 F. We would like to determine the feasibility of using the GM ORC as its waste heat recovery system. Other known conditions and desires are that the cooling water pump power requirement is 1.5 kw for a 120 gpm flow rate, the payback period requirement is 10 years, the expected capital is $300,000, the interest rate is 10%, fuel cost is $5/gallon, and the system will operate year-round. Based on the given information of hot-water temperature and flow rate and cold water temperature, the GM net output power (which can be found using Fig. 19) is 27 kw and the operating output power is 25.5 kw (27 kw 1.5 kw). Based on the expected capital cost and interest rate, fuel cost, and desired payback period, the required GM operating output power is about 24 kw (found using Fig. 23). Comparing the GM operating output power (25.5 kw) with the required power of 24 kw, it seems marginally feasible to adopt the GM system for jacket water heat recovery for this power plant. Since this calculation is based on water as heating fluid, if the heating fluid is changed to 60/40 PG/W, the GM operating output power needs to be modified accordingly. Then the GM will become marginally unfeasible at this power plant Economic (Payback Period) and Performance Estimates The test data of the GM performance (e.g., power output, efficiency, emissions and CO 2 reductions) and economics can be directly applied to the village diesel generator applications when the field diesel operation data are available. An example to evaluate the economics of applying the GM ORC system using real-life diesel engine operation data is presented for jacket water heat recovery, and combined jacket water and exhaust heat recovery systems (the best result are for hot water temperature up to 225 o F) using the developed performance curves (see section heading Example Based on Performance Curves ). The example shows that the performance data obtained from this experiment can be used to simulate and evaluate the application of this ORC system to an Alaska village generator set for power output, efficiency, payback period, emissions reductions, etc. For jacket water temperature at 99 o C (210.2 o F) (this may be optimistic; for regular application the temperature may be around 85 o C), 41.7 kw system operating power output was achievable with 7.2% efficiency and 2.7 year payback. From our observation of another example, it is possible to generate 45.7 kw system operating power output with 7.4% efficiency and 2.4 year payback using this ORC power unit working on waste heat from stationary diesel engines if the waste heat is from both jacket water and exhaust heat exchanger (for up to 225 o F hot water temperature). Considerable reductions in emissions could be achieved, as listed in Table 22. The test data can also be used to estimate the effect of applying the GM on the economy of the rural diesel power industry. In rural Alaska, more than 180 villages use diesel generator sets for their electrical power supply. The average electrical power of each of the villages is given in the Statistical report of the power cost equalization program for fiscal year 2011, published by Alaska Energy Authority in April Based on the diesel power data published, and assuming that engine efficiency is 40%, the amounts of jacket water (about 20% of fuel energy) and 74

87 exhaust (about 30% of fuel energy) are calculated and used to estimate electrical power that could potentially be generated by GM for each of the village diesel generator sets. The estimated fuel savings and power generation through applying the 50 kw GM for each of the villages in rural Alaska are listed in the Table of Appendix VI Proposed General Policy in GM Application In general, the greatest potential that waste heat recovery applications have for rural Alaska diesel generators are for heating and for power (against other applications, such as chilling, desalination, etc.). Heat recovery for heating may capture up to 85% of waste heat, depending on type of heat source, heat flow conditions (temperature and flow rate), effectiveness of infrastructure, and detailed application (e.g., space heating, city water temperature maintenance, washateria), and the application is seasonal (cold season). Considering heat recovery for power in rural Alaska, the ORC is currently considered the most efficient system. Furthermore, considering the most popular engine sizes used in rural Alaska (diesel generator output is roughly less than 1 MW), the GM ORC is one of the most effective systems available so far. Based on test data obtained from this experiment, the expected operation efficiency of the GM ORC for diesel generators is about 5% to 7.5%, depending on jacket water and cooling water conditions. For preliminary estimates of economic effects of waste heat recovery for heating and waste heat recovery for power on the rural Alaska diesel power industry, major factors may include fuel savings, costs of fuel/heat/electrical power, and costs of infrastructure and installation. Except for fuel savings, other factors that are village- and time-dependent may not be obtainable from data resulting from these tests. Cost of electricity and cost of MBtu of heat energy may be easily found for a particular village. In this section, a rough comparison in annual fuel savings between waste heat for heating and waste heat for power is presented. This way the monetary benefits in fuel savings can be easily obtained from fuel savings using electricity and heat energy cost information for individual villages. In addition, fuel savings may also be converted into potential benefit of reductions in emissions and CO 2 for individual villages. Considering heat recovery for heating, the assumptions involved in fuel saving estimates include 80% efficiency of heat recovery for heating, 7 months of average season of heating application, and 80% of boiler efficiency. Assumptions used for heat recovery for power applications (based on GM ORC data) include 6.8% of heat-to-power efficiency, year-round application, and 35% of diesel generator efficiency. Fuel heating value is assumed at 130,500 Btu/gallon. In addition, the discussion in this section is based only on jacket water heat recovery, since currently, only jacket water is preferred by the Alaska power industry. According to assumptions mentioned above, every 1 kw of jacket water heat may recover 4,380 kw-hr of heat annually for useful application, which corresponds to a potential of diesel fuel savings of 143 gallons. For heat recovery for power, every 1 kw-hr of jacket water heat may generate 596 kw-hr power annually, which corresponds to a potential of diesel fuel savings of about 45 gallons. If exhaust heat recovery is considered, the fuel saving for heat recovery for power may be improved because of the possibility of a much higher temperature of heat source, but the expected fuel saving for power generation alone may still be significantly lower than that for 75

88 heat recovery for heating. However, if the potential of further recovery of residual heat is also considered, the percentile difference may become much less. A few points worth mentioning are that electrical power is more flexible to apply than heat energy; the cost of infrastructure and installation for both applications are village-dependent and may become critical for economic benefit (for heat recovery for heating, the amount of arctic pipe needed may become the dominate cost; for heat recovery for power, the cooling system may become very important.); and cost of fuel may be different for heating application as it is for power generation (fuel cost difference may not be essential). In addition, for heat recovery using high temperature heat, the residual heat in the hot fluid (e.g., water) existing in the GM ORC may still have good quality (or temperature) for multi-stage heat recovery (heating applications and/or multi-stage power applications). Therefore, if the case is suitable, multistage and/or multi-tasks heat recovery applications may be the preferred choice. In general, the following conditions may be critical in selection of an ORC system for individual villages: Preference of the Village: The preference of the village may not have much to do with technology and/or economy. Based on previous records, some villages are reluctant to modify their diesel generators for heat recovery applications. Existence of Heat for Heating System: If a village already has heat recovery for its heating system in operation and can effectively use all the heat energy, then heat recovery for power can only be considered for summer operation after having been verified that it will benefit the village economy. However, due to low summer village electricity usage, the possibility for the existence of other types of electrical power systems (such as hydraulic) and the possibility that high cooling source temperature is available for ORC cooling, the benefit to the village may be negligible or not worth the potential problems (maintenance, management that results from applying the ORC system) caused by applying the ORC. For the case if heat energy is more than the need of the village, an ORC may be adopted and considered for year-round usage, assuming that the leftover heat is enough to operate the ORC and can bring benefit to the village. Low Availability of Heat for Heating System: For this case, economic benefit may become the greatest concern for choosing one technology over another. Economics may depend on many factors, such as location of the power plant (the distance between the power plant or heat source and the location of application, normally public buildings such as power plant space and office, hospital, school, library, washateria, waste water treatment, city water temperature maintenance), the existence of (workable) heat recovery for heating infrastructure for heating (heat exchanger, arctic piping system). Some villages may have infrastructure, but it may be in need of repair or renovation. 76

89 Detailed policy in how to use the 50 kw GM can also be derived based on information given in the previous three sections of this chapter Conclusions from Laboratory Testing This section discusses the accomplished tasks related to the project objectives and the work remaining. Comments about the GM based on the experience obtained from this testing project are included Accomplished Tasks Related to Project Objectives Both the objectives of this project and the performance results are given for comparison: 1. Objective 1: To prove that an improvement of the efficiency of the diesel power plant by about 10% (i.e., about 4% of fuel efficiency) is achievable through the use of an organic Rankine cycle (ORC) system, which uses waste heat contained in diesel engine jacket water and exhaust. Based on experimental results, the maximum net efficiency experienced at the heating fluid temperature of 195 o F (jacket water temperature) is about 7.4%. For more than 50% of fuel energy as waste heat (in jacket water and exhaust), the potential of fuel efficiency improvement is about 3.7%, which is close to the target of 4% fuel efficiency improvement. In addition, if a higher temperature heat source (obtainable from exhaust) is used and the cap of the GM output restriction is moved from 50 kw to 65 kw to improve the maximum efficiency for high temperature fluid (245 o F), the 4% improvement in fuel efficiency seems reachable. 2. Objective 2: To evaluate feasibility, operation and maintenance requirements, and payback time of applying a selected ORC system. Based on observation and operation experience, the ORC system is considered reliable (under normal operation condition, no foreseen technical problems are expected for long-term operation), and no advanced technology background is needed for operation. As for maintenance requirements, no advanced technology background is needed, and besides the consuming materials (filters, lube oil, etc.), extra cost is minimal if the GM routine maintenance schedule is incorporated into the routine diesel genset maintenance schedule (for details, please refer to the section Data Analysis ). Based on the reliability testing results (at full capacity of the GM), the estimated payback time is 2.1 years for a 0% interest rate and 2.4 years for a 10% interest rate. For real-life reliability, a longer field test is needed. 3. Objective 3: To develop guidelines for ORC system selection, operation, and maintenance; and to evaluate the potential impact of applying waste heat ORC systems on the rural Alaska economy, fuel consumption, and emissions and greenhouse gas reductions. 77

90 Details are provided in this chapter under the section heading Discussion. Operations: Most of the operations procedures for the GM are described in the manufacturer s manual. No additional suggestions are required for operation, but some of the most important ones are given below: a. When working on the GM, if the hot water loop does not have a bypass, then do not start the hot water or cold water supply to the GM (until ready to start the GM), as it may pressurize the working fluid and rotate the expander and generator. b. During the regular checklist, check the expander high pressure value and hot water inlet temperature. A hot water inlet temperature greater than 245 o F may result in over-pressurizing the system, which in turn may lift the working fluid pressure relief valves causing a loss in working fluid and possible damage to components. c. In case an emergency shutdown of the GM is necessary, adjacent to the HMI screen is an Emergency Stop Button on the front panel of the GM. Pressing this will shut down the GM immediately. Maintenance: Similar to operations, most of the maintenance procedures for GM are described in the manufacturer s manual and the manual is current. No additional suggestions are required for maintenance, but some of the most important ones are listed below: a. Check for non-condensable gases in the system (procedure given in manual). Purge the non-condensable gases from the system following the procedure given in the GM manual. b. Visually inspect all the joints and connections for oil/water leak. c. Visually check the electrical wiring frequently for any damaged connections due to excess heat or loose connections (by the tug test). Evaluation of the Potential Impact on Rural Alaska Economy, Fuel Consumption, and Emissions and Greenhouse Gas Reductions: Evaluations of fuel and monetary savings for each of the village gensets are mentioned in a previous section and details are given in Appendix VI. Emissions and CO 2 reductions can be found using fuel savings and power generated from the 50 kw GM ORC, such as the information given in Table Objective 4: Performance and economic comparison of the two ORC systems. One ORC system is a 50 kw system that uses a screw expander and is considered an emerging technology. The second ORC system is a Pratt & Whitney (P&W) 250 kw unit that uses a radial turbine and is considered a well-developed technology. Due to the shortage of measurement data for the Cordova ORC, performance comparison between the two systems cannot be performed directly using measured data. Performance comparison has been done based on many assumptions adopted from communication with the power plant engineers, and using the genset data sheet and field test data. 78

91 Preparation for Phase II Field Testing Following completion of the Phase I laboratory test, it was the intent of ACEP to transport the GM to a power plant in the Tanana Chiefs Conference (TCC) region and install it in order to conduct a field test under real-world conditions. ACEP maintains a partnership with TCC, a nonprofit consortium of forty-two Alaska Native tribal communities in Interior Alaska, for rural energy research and development. The power plant in Tok was selected as the project location. The following is a list of the major activities accomplished for relocation and preparation: 1. Procurement and layout of required heating and cooling piping systems, electrical circuit for uploading, and instrumentation and monitoring system: The emphases are on optimizing the benefit to the village and avoiding negative effect due to the added system on the performance of the diesel power plant, such as emissions and combustion efficiency and the stability and effectiveness of the local electrical grid. 2. Preparation of the ORC system for relocation and starting: This included discharge of refrigerant from the ORC system, packaging and shipping, recharge, and starting the ORC system. 79

92 3 Phase II: Field Testing 3.1 Project Preparation and Test Plan Upon completion of laboratory testing in Phase I, Phase II testing on the Green Machine (GM) took place in an Alaska village power plant. Alaska Power and Telephone (AP&T) agreed to install and test the GM in its Tok facility. This location was selected as a good test site for the following reasons: The community is on its own grid, enabling testing of the GM under the more challenging conditions of grid isolation. The Tok plant is representative in size (about 2 MW at peak production) of a plant in rural Alaska, and the community has an interest in saving diesel (consumes about 750,000 gallons of diesel annually). Tok is on the road system, thus lowering transportation costs and facilitating access to goods and services used in installing the GM. Employees at the AP&T Tok power plant have the expertise to install the GM and to document the process. 3.2 Installation at the Tok Power Plant The GM was run under normal operating conditions of the Tok power plant. The heat source was the jacket water from the diesel engines. The cooling source was water from an existing deep well at the power plant. Fig. 24 shows the heating loop and cooling loop connections to the GM. Flow meters are connected to the piping system with two isolation valves, one on each side of a flow meter, for data collection and performance monitoring. Temperature sensors and pressure gages are inserted into the piping through thread-o-lets, which are sealed when data collection is complete. See Appendix A for the schematic of the connections between the Organic Rankine Cycle (ORC) system and the power plant Heat Source The heat source for the GM is derived from the power plant s heat recovery system, which consists of heat exchangers on the engine s jacket water and a 5-HP pump that circulates a 60/40 propylene glycol/water mix through the heat exchangers to provide space heating to the power plant and adjacent buildings. The jacket water for all diesel generators in the plant are connected in parallel and join two respective pipes that transmit relatively hot water into the GM and relatively cold water out of it. During the cold winter months, the priority use for the relatively hot jacket water is for space heating. In general, if the environment temperature is above -20 F, the jacket water will first flow through the ORC system for power generation, then continue flowing through the space-heating system. When it is below -20 F, the GM is bypassed and all the heat is delivered to the space-heating system. The temperature of jacket water ranges from 172 F to 190 F, with an overall-time average of about 182 F as it enters the GM, 80

93 which is on the lower side of the allowable inlet temperature range of 150 F to 250 F. Typical hot side flow rate is around 150 gpm with the GM in the circuit. Fig. 24. Schematic line diagram of heating and cooling to the Green Machine Seven modifications were made to the heat recovery system during the installation of the GM: (1) Replacement of the heat recovery exchangers on units 3, 8, and 9. (2) Installation of the thermostatic valves on the jacket water of diesels 4, 5, and 7 (bypassing jacket water flow to reduce heat loss to the outdoor radiators). (3) Reconfiguration of plumbing to provide series flow through the exchangers in the heat recovery loop. The series circuit maximizes flow through the heat exchanger. This modification included manual valves on the exchangers enabling bypass of any units that are not operating. (4) Raising the coolant expansion tanks on units 3, 8, and 9 to accommodate the exchanger replacements. (5) Extension of the 4-inch supply and return piping over the truck bay door to the GM. (6) Installation of isolation and bypass valves for each exchanger in the heat recovery loop. (7) The addition of flow straighteners, flow meters, pressure gages, and temperature measuring equipment; installed for monitoring performance. 81

94 3.2.2 Cooling Source The water used as cooling liquid for the ORC system is drawn from an existing well (Well No. 2), from a static water level of about 60 feet deep, with a 5-HP submersible pump (the same as the Tok No. 1 well pump attachment) in a 6-inch well casing. The well is close to an outside wall, where there is a suitable outdoor area to use for the drain field. A 3-phase 480-volt power source is also in proximity, and the GM is situated out of the way on the operating floor. The well water temperature of 36 F (read from the data acquisition system) is relatively consistent year-round and was sustained without problems during field testing. The ORC condenser has an open cooling system with a drain field and a pump that runs at a constant power of 5.4 kw for 64 gpm (read from the data acquisition system). The 2-inch well plumbing is set up with valves to discharge into the plant s common water system or into the 3-inch pipe that extends to the GM, as shown in Fig. 25. The low flow rate of 64 gpm required modification of the GM flow switch to operate at reduced flow. The drain field where the cooling water was discharged consisted of a buried header pipe running south from the plant and two 90-foot-long runs of perforated 4-inch PVC pipe running west about 6 feet apart. The gravel under the perforated pipe was loosened to a depth of about 4 feet under the pipe, and the pipe was covered with Typar to keep fine material out of the leach area. Provisions were made to connect additional perforated pipe runs off the header, but there were no problems with the drain field accepting the cooling water discharge. The electrical connection was via a 50-foot run of three #2 copper thermoplastic high heatresistant nylon-coated (THHN) wires plus a ground wire to a 3-phase, 480-volt, 100-amp fused disconnect switch, which was fed from an outdoor 75 kva pole-mounted transformer bank. Electrical metering was accomplished using a utility standard Form 16S watt-hour meter. The disconnect and the meter are visible in Fig

95 Fig. 25. Well Number 2 The cold-water piping includes a flow straightener, flow meter, and temperature and pressure probes, as shown in Fig. 26 and Fig

96 Fig. 26. GM as installed in the Tok power plant 84

97 Fig. 27. GM cold- and hot-water piping 3.3 Instrumentation and Measurement/Evaluation Equipment As part of ACEP s data collection efforts, electrical meters were installed on the Number 2 well pump and the heat recovery system pump. While the Number 2 well pump was an additional parasitic load, the heat recovered from the diesel engines for use in the GM reduced the parasitic cooling loads associated with operating the radiator fans and Number 3 well pump. Since the radiator fan loads and the Number 3 well pump were not monitored, the potential offset has not been quantified. Instrumentation for data collection from the GM field test was designed to measure the following parameters: The heat supplied to the GM by the 60/40 propylene glycol/water mixture The heat rejected by the GM to the cooling water The electrical power generation of the GM The parasitic power consumptions of hot water, cold water, and GM pumps Table 23 gives the instrumentation components procured for the data collection. 85

98 Table 23. Instrumentation equipment and components for data collection from the GM 50 kw field test at Tok, Alaska Measurements and Components Components Quantity Instrumentation Provider 60/40 propylene glycol/water and cold-water flow rates for Flow meter 2 ACEP BTU meter 60/40 propylene glycol/water and cold-water inlet and Temperature outlet temperatures for BTU sensors 4 ACEP meter Green Machine net power output and pump power Electrical meters 2 ACEP consumption 60/40 propylene glycol/water and cold-water pump power Electrical meters 2 ACEP 24 VAC power supply for BTU meters Adapter 1 ACEP Data logging Computer 1 ACEP Internet connection Two Internet connections: one for GM and another for data logging computer AP&T The peripheral equipment also included cables to connect temperature sensors and flow meters to BTU meters, electrical cable for the data acquisition system (DAQ) of the electrical meters, and a 4.5- by 8-foot table upon which were placed all the DAQ equipment, the computer, and other components. Fig. 28 shows the DAQ line diagram, which illustrates how all instruments are connected for data collection. 86

99 Fig. 28. Data acquisition system for Green Machine field test in Tok Several trips to Tok from Fairbanks were required for installing the instrumentation system, details of which are given below. The first trip took place on August 5, During this trip, the following tasks were accomplished: (i) (ii) (iii) Instrumentation components (listed in Table 23) as well as other accessories (such as cables, table, etc.) were delivered to the Tok power plant. A location near the GM was selected for placing the table, DAQ computer, components, and other peripherals for data collection, with consultation from power plant personnel. Required cable lengths (based on the DAQ computer location) were determined. A list of additional small items such as tools, USB cables, and LAN cables required was also compiled; these items were brought to Tok during the second trip. The second trip to Tok occurred on August 13 14, During this trip, the following tasks were accomplished: (i) (ii) The DAQ components (temperature sensors, pressure gages, BTU meters, electrical meters, etc.) were installed; no major problems were encountered. The BTU meters and electrical meters (Fig. 28) were connected to the computer for data collection and monitoring. 87

100 (iii) (iv) The readings from the components were monitored and checked to make sure all of the data collection functions were being performed as desired. The DAQ computer was connected for remote monitoring from UAF over the Internet. The remote-monitoring process was also checked during this trip. The Phase II field test collected data for estimating parameters such as efficiency, reduction in fuel consumption, reduction in emissions, and payback period. 3.4 Test Schedule and Test Plan Testing began as soon as the GM was commissioned in the Tok power plant and running smoothly, on October 2, The goal of the test plan was to obtain enough operational information for system performance analysis, which then could be used for economic and emissions analyses and development of ORC application guidelines. The test plan included the selection of parameters to be measured for needed performance analysis and the assurance that measured data could reliably represent the operation conditions of ORC technology in general, and the GM system specifically. The reliability of measured data depends heavily on the accuracy of the measurement equipment used and on the measurement schedule (e.g., time step sizes for individual parameters). Selected parameters used for measurement are found in Table 23, and details of measurement devices selected and the measurement schedule are found in Table 24, and discussed under the operations and maintenance section of this report. 3.5 Procurement of Required Equipment and Supplies See Appendix C for complete materials list. 3.6 Transport and Installation of the Green Machine To transport the GM, removal of the working fluid (R-245fa) into the GM storage tank for safe transport is required by the Federal Motor Carrier Safety Administration guidelines. (The refrigerant R-245fa has a transportation UN number, 3163, and is a Class 2 substance [pressurized gas].) The GM was transported from the University of Alaska Fairbanks (UAF) to Tok by land on July 2, using the commercial shipping company Lynden Transport. Included with the shipment was a box of additional parts, equipment, and documentation, as well as six empty refrigerant bottles, in a covered van that had air-ride suspension for protection against shock and vibration. A forklift was used to load the GM into the van, as well as to unload the GM at the Tok powerhouse. There, a 10-ton bridge crane moved the GM to its final location. No physical damage, fluid leakage, or change in pressure readings were observed during or after transport. Although the manufacturer approved this shipping procedure for this one situation, their typical recommended procedure for transporting the GM is to have the working fluid removed and shipped in separate pressurized tanks (six 100-pound shipping tanks were provided). In that case, an HVAC technician would be needed to test and recharge the system ( pounds of R-245fa) at the destination location. Depending on location, the total cost of travel and labor 88

101 for the technician would be in excess of $2,000. Although AP&T engineers started the GM with ElectraTherm s phone and guidance, a typical startup would have ElectraTherm s technicians on-site for commissioning. The GM s approximate weight is 5,000 pounds, so heavy-lifting equipment is necessary to move it. Since the heat requirement for operation of the GM is significant, installation of such a unit is likely to be at a facility sufficiently large to have on-hand equipment appropriate for moving the machine. The current model of the GM offers a skid frame to facilitate moving it. 3.7 Operation and Maintenance Requirements ElectraTherm s GM manual describes the routine and preventative maintenance items and schedule (see the attached Green Machine Operating Manual, Part 2, pages 10 12). Prior to 4,400 running hours, only completion of visual inspection of the GM is specified. The GM in Tok had 1,000 hours of service when it was shipped, thus only visual maintenance was performed during Phase II testing. The Tok unit suffered an expander failure after 2,800 hours of operation. ElectraTherm has stated that they are aware of this issue for their Block 1 machines and has implemented lubrication changes to subsequent models to avoid similar failures. ElectraTherm indicated to AP&T that, with a replacement expander and new type of oil, the expected life of the expander is 15,000 hours. If AP&T decides to rebuild and continue operation of its Block 1 unit, it would be prudent for them to design and incorporate a ventilation system to mitigate potential future refrigerant leaks. After the first seal failure, the only option available to vent the Tok powerhouse was to open the doors for an extended period. Since the refrigerant does not vaporize below 60 F, the cold incoming air limited vaporization and hampered the ventilation process. Fortunately the outdoor temperature was above 0 F, but had a leak occurred at even colder temperatures (Tok regularly experiences temperatures around -50 F), the risk of freezing equipment would be greater, and ventilation of the refrigerant even more difficult. 3.8 Data Collection As shown in Table 24 (and Fig. 28), the measurements of flow rates and temperatures for both 60/40 propylene glycol/water and cold-water loops were performed using a Kamstrup Ultraflow ultrasonic flow meter and a Kamstrup Pt500 temperature sensor, respectively. The hourly average data for flow rate and temperatures were stored in the Multical-601 calculator (marked as BTU meter in Fig. 28). LogView, custom software provided by the BTU meter manufacturer, was used to download this hourly average flow rate and temperature data in the Excel format. For electrical power measurements, as shown in Table 24, EKM-353EDM-N electrical meters were used to measure electrical power generated by the power unit, as well as power consumption by the power unit pump, the hot-water pump, and the cold-water pump. The electrical meter manufacturer had its own custom software (EKM Metering) which was used for 89

102 reading real-time electrical power measurements to the computer. This real-time data were stored in text format at 30-second intervals for future data reduction. Table 24. Parameters measured and instrumentation used during Phase II testing Parameter 60/40 propylene glycol/water supply temperature to Green Machine 60/40 propylene glycol/water return temperature from Green Machine 60/40 propylene glycol/water flow rate Cold-water supply temperature to Green Machine Cold-water return temperature from Green Machine Cold-water flow rate State point on Error! Reference source not found. T h,in T h,out V h T c,in T c,out V c Instrumentation Kamstrup Pt500 temperature sensor Kamstrup Pt500 temperature sensor Kamstrup Ultraflow ultrasonic flow meter Kamstrup Pt500 temperature sensor Kamstrup Pt500 temperature sensor Kamstrup Ultraflow ultrasonic flow meter Units o C o C m 3 /h C C m 3 /h Data collection frequency 1 hour (hourly average) 1 hour (hourly average) 1 hour (hourly average) 1 hour (hourly average) 1 hour (hourly average) 1 hour (hourly average) ORC power unit net power generated W net EKM-353EDM-N kw 30 seconds ORC pump power consumption W ORC,p EKM-353EDM-N kw 30 seconds 60/40 propylene glycol/water pump power consumption W h,p EKM-353EDM-N kw 30 seconds Cold-water pump power consumption W c,p EKM-353EDM-N kw 30 seconds 3.9 Results from Field Testing The GM ran continuously between October 2, 2013, 11:00 A.M. and November 19, 2013, 7:00 A.M., for a total of hours. Data was continuously collected between October 2, 2013, 12:00 P.M. and November 9, 2013, 1:00 P.M. Between November 9 and November 19, data were not taken due to an unintentional trip-off of the switch connecting the data acquisition computer and the sensors. After November 19, the GM was intentionally bypassed from the jacket flow piping system to prioritize space heating. The collected amount of data is prohibitively large for imbedding here; detailed measured ORC performance data of thermal parameters (hourly) and electrical parameters (every 30 seconds) will be provided upon request. Fig. 29 and Fig. 30 show variations of some of the collected and deduced daily data. Fig. 29 presents the 60/40 propylene glycol/water supply and return temperatures of the heating flow 90

103 into and from the GM. Since the heating fluid flow rate is almost constant, the heat delivered to the ORC may be determined solely from the delta between the entering and returning temperatures of the heating fluid. Fig. 30 presents the ORC power output (involving the cooling water power consumption), GM pump (for refrigerant working fluid) power consumption, and cold-water pump power consumption data. Fig. 30 shows that the power consumption of both pumps was quite steady; this continued to be the case throughout the field-testing process. The reduced form of the field test data is listed in Table 25. Due to the constancy of the cooling water temperature (36 F) at a flow rate of 64 gpm, and the 60/40 propylene glycol/water flow rate entering the ORC system at gpm, the reduced data are organized according to the only variable, which is the temperature of propylene glycol/water (PG/W) entering into the ORC system. In this report, only data taken between October 2 and November 19 are used for performance, emissions, and economic analysis. According to AccuWeather ( the environmental air temperatures during the period October 2 to November 19 range from about 50 F to about -15 F, which covers the useful operational range of the GM operation in Tok. Below -15 F, the heat is used for space heating and the GM is turned off. 91

104 Fig /40 propylene glycol/water supply and return temperatures to Green Machine 92

105 Fig. 30. Green Machine operating power output, Green Machine pump power, and cold-water pump power consumption 93

106 Table 25. Reduced form of the recorded Tok field test data and generated Green Machine ORC performance data during the field test period 60/40 propylene glycol/water supply temperature to Green Machine F ( C) Heat supplied by 60/40 propylene glycol/water to Green Machine (kw) Heat rejected to cold water by Green Machine (kw) Green Machine net power output (kw) Green Machine pump power consumption (kw) System operating power output (kw) System operating efficiency (%) 172 (78) (79) (80) (81) (82) (83) (84) (85) (86) (87) (88) After the GM was turned off on November 19 because of low environmental temperature and the unavailability of sufficient heat energy after space heating, the machine was turned on and off sporadically. On December 9, the GM was shut down due to a failed seal and refrigerant leakage from the screw expander. ElectraTherm was contacted once the leak was found. More details about the defect, which has been corrected in the current model of the GM, can be found in Section Field Data Performance and Economic Analysis The average GM performance, including average net electrical power generated and net efficiency of the GM (not including the effect of cooling water pump power), has been analyzed using collected data. This analysis corresponds to an assumed case, that the cooling pump power offsets the power consumed by the engine radiators and the Number 1 well pump. The analysis results are summarized in Table 26. For the Tok field test, cooling water for the GM came from an existing groundwater well and pump system. During the entire period of testing, the pump power consumption was almost constant at 5.4 kw for a cooling water flow rate of 64 gpm (from the DAQ). Table 26 also gives estimated system operating performance and emissions reductions analysis results, including the effect of the cooling water pump power, based on data from the hours of field testing. Based on a Bell & Gossett manual, a more effective and commercially available 2.24 kw pump can serve the same purpose. Using a smaller power pump would consume less parasitic power and give better performance efficiency to the ORC system. The analysis extends to two 94

107 other assumed cases, results of which are given in Table 26. One case assumes using a 2.24 kw cooling water pump, and the other assumes no water pump, to examine the effect of cooling water pump power consumption on the system operating performance and emissions. Table 26. ORC performance during continuous operation period from 10/02/2013, 11:00 A.M., to 11/19/2013, 7:00 A.M. ( hours) Parameter Value Average 60/40 PG/W supply temperature to power unit (T h,in ) 83 C (182 F) Average 60/40 PG/W flow rate to power unit (V ḣ ) 28.5m 3 /h (125 gpm) Average cold-water supply temperature to power unit (T c,in ) 2 C (36 F) Average cold-water flow rate to power unit (V c) 14.5 m 3 /h (64 gpm) Total power unit electrical power output for the operation period kwh Total ORC unit pump power consumption for the operation period 1558 kwh Total cold-water pump power consumption for the operation period 6176 kwh Average net power unit electrical power output (P Net ) 22.0 kw Average power unit pump power consumption (P Pump,P ) 1.4 kw Heat supply by 60/40 PG/W to power unit evaporator (Q h) 379 kw Screw expander efficiency (η SE ) 6.2% Power unit efficiency (η Net ) 5.8% Estimated results based on cold water No cold water Results based on Tok s pump selected using pump power cold water pump power Bell & Gossett considered software* Cold-water pump power consumption (P Pump,CW ) 5.4 kw 2.24 kw --- Total system operating power output for the operation period kwh kwh kwh System operating power output (P OP ) 16.6 kw 19.8 kw 22.0 kw System operating efficiency (η OP ) 4.4% 5.2% 5.8% Diesel fuel saved (F S )** lit (1350 gal) 6080 lit (1610 gal) 6770 lit (1790 gal) Dollar amount saved on diesel fuel (F $ )*** $6740 $8030 $8940 Emissions reductions Oxides of nitrogen (NO X ) 66 kg (15 lb) 79 kg (170 lb) 88 kg (193 lb) Hydrocarbons (HC) 8 kg (17 lb) 9 kg (20 lb) 10 kg (22 lb) Particulate matter (PM) 2 kg (4 lb) 2 kg (5 lb) 3 kg (5.5 lb) Carbon monoxide (CO) 66 kg (146 lb) 79 kg (174 lb) 88 kg (193 lb) Carbon dioxide (CO 2 ) kg (15 short-tons) * Bell & Gossett EPS Plus pump selection software. ** Based on diesel engine specific fuel consumption of 14 kwh/gal. *** Diesel fuel price = $5/gal kg (18 shorttons) kg (20 shorttons) As shown in the upper part of Table 26, during the testing period between October 2 and November 19, the average (gross) power generated from the screw expander was 23.4 kw, the ORC working fluid pump power consumption was 1.4 kw, the net power generated from the 95

108 ORC system was 22.0 kw (i.e., screw expander power working fluid pump power), the heat supply to the system was 379 kw, and the ORC system net efficiency was 5.8% (net output power/heat supply). During the period of field testing, the operation of the ORC system generated 18,863.6 kwh (i.e., system operating output = system net output cooling water pump energy consumption) of electrical energy, saved 1347 gallons of diesel fuel (based on a specific fuel efficiency of 14 kw/gallon), and operated at an operating efficiency of 4.4%. Considering emissions reduction during this period, the reductions in regulated emissions were 66 kg in NO x, 7.5 kg in HC, 1.9 kg in particulate matter, and 66 kg in CO; the reduction in CO 2 was 15 short tons. The estimated reduction in emissions is based on the Tier 4 EPA standard, and the reduction in CO 2 is based on 22.2 pounds of CO 2 per gallon of diesel fuel. Based on calculated system operating performance (total operating power output, diesel fuel saved) and emissions reductions during the test period, a payback period can be calculated. The estimated performance and emissions reduction calculations for the system are for year-round operation versus 7.5 months of operation, when heating is not required (Table 27). Estimated payback periods are given for a 0% interest rate and a 10% interest rate. Table 27. Estimated ORC performance in Tok, Alaska, for a full year of operation (8,760 hours) Cold-water pump power consumption (P Pump,CW ), kw System operating power output (P OP ), kw Total system annual power output, kwh System operating efficiency (η OP ) Diesel fuel saved per year (F S/Y )**, liters (gallons) Dollar amount saved on diesel fuel per year (F $/Y )*** Power production based on Tok s actual pump power Full year of operation Operation from Apr. 1 Nov. 15 (5472 hours) Estimated power production based on cold water pump selected using Bell & Gossett software* Full year of operation Operation from Apr. 1 Nov. 15 (5472 hours) Power production without cold water pump power considered Full year of operation Operation from Apr. 1 Nov. 15 (5472 hours) % 4.4% 5.2% 5.2% 5.8% 5.8% (10380) (6480) (12370) (7730) (13770) (8600) $51,880 $32,400 $61,850 $38,640 $68,860 $43,010 Emissions reductions Oxides of nitrogen (NO X ), (1120) 600 (1340) 670 (1490) kg/year (lb/year) (700) (830) (930) Hydrocarbons (HC), (130) 69 (150) kg/year (lb/year) (80) (95) (170) (106) Particulate matter (PM), kg/year (lb/year) 14.5 (32) 9.1 (20) 17.3 (38) 10.8 (23) 19.3 (43) 12.0 (27) Carbon monoxide (CO), (1120) 606 (1340) 675 (1490) kg/year (lb/year) (700) (830) (930) Carbon dioxide (CO 2 ),

109 Power production based on Tok s actual pump power Full year of operation Operation from Apr. 1 Nov. 15 (5472 hours) Estimated power production based on cold water pump selected using Bell & Gossett software* Full year of operation Operation from Apr. 1 Nov. 15 (5472 hours) Power production without cold water pump power considered Full year of operation Operation from Apr. 1 Nov. 15 (5472 hours) kg/year (short-tons/year) (115) (72) (137) (86) (153) (96) Payback period Payback 0% interest on capital, years Payback 10% interest on capital, years 23.2 N/A 13.5 N/A 10.7 N/A * Bell & Gossett EPS Plus pump selection software. ** Based on diesel engine specific fuel consumption of 14 kwh/gal. *** Diesel fuel price = $5/gal. Capital cost estimated at $394,000 (including GM cost and installation in power plant) and Maintenance cost = $7600/year Discussion The average GM net power generation in Tok was 22 kw. Assuming a Tok summer diesel electrical power generation of 1,116 kw and specific fuel consumption (SFC) of kwh/gallon, the SFC of the combined diesel generator and GM system becomes kwh/gallon, which gives an increase of 0.28 kw/gallon in SFC or a 2% of SFC improvement. As shown in Table 27, for year-round operation of the ORC system under field-test operation conditions (capital, installation, and maintenance costs are given as a footnote of Table 27), the payback times are 8.9 years and 23.2 years for 0% and 10% interest rates, respectively. If the heat from the jacket water is used for space heating 4.5 months annually and the GM is utilized for 7.5 months of the year, then the payback time is 15.8 years for a 0% interest rate. Based on an expected 20-year lifetime of the ORC system and the current GM operating environment in Tok (i.e., heat and cooling sources), the installation of an ORC system may give a marginal benefit only if the interest rate is around or less than 7%, and the system operates year-round. To have a marginal benefit for 7.5 months per year of operation, the allowable interest rate needs to be much lower. However, if the cooling water pump power is not considered parasitic power (i.e., an offset of the power consumed by the radiators and No. 2 pump), then the payback period will become much shorter. As shown in Table 27, for year-round operation of the GM and no parasitic power, the payback times are 6.4 years and 11.1 years for 0% and 10% interest rates, respectively. For 7.5 months of operation per year, the payback time is 10.7 years for a 0% interest rate. Based on review of the analysis results, topics relating to power generation and parasitic power are discussed in the following sections: Effect of jacket water temperature on system efficiency During this field test, net power generated by the ORC system was much lower than the rated (nameplate) capacity of 50 kw, which mostly is due to the low heating fluid temperature 97

110 (average of F) from the jacket water. Based on the system output versus inlet flow temperature data obtained from the previous laboratory test with hot water as the heat source, the system output could be increased by about 4.5 kw with a temperature increase from F to F, which may be an attainable jacket water temperature under different diesel engine operation situations. For example, in a plant with only one generator, a higher temperature could be obtained by using the jacket water directly instead of the heat recovery loop, although this would not be practical in plants with multiple generators. Fig. 31 presents the GM net power outputs versus heat source supply temperature for water (data from the UAF laboratory test) and for 60/40 propylene glycol/water (data from the Tok field test) as the heating fluid. For a 4.5 kw increase, the system s net output would be 26.5 kw, and operating efficiency would be increased from 4.4% to about 5.6%, assuming a 5.4 kw cooling parasitic power; or if there is no cooling parasitic load, then efficiency would increase from 5.8% to 7.0%. Fig. 31 shows that the net power output using water as heating fluid yields about 10% more power from the system than when 60/40 propylene glycol/water is used as heating fluid. Fig. 31. Comparison of Green Machine UAF lab test results and Tok field results for net power output for same heat source supply temperatures 98

111 Effects of differing pump power parasitic loads Working Fluid Pump Power The working fluid pump s power consumption of 1.4 kw is comparable to the pump power consumption obtained from the laboratory test for similar system power-output cases. Cooling Water Pump Power The cooling water pump used at the Tok power plant consumes 5.4 kw, which is 3.16 kw more than the typical power usage (2.24 kw) of a commercially existing pump for similar cooling water operating conditions. The best and ideal case would be if no cooling water pump power were needed; for example, if a sufficient local surface water resource were near the power plant. The bottom part of Table 26 compares system operating efficiency, fuel savings, and reduction in regulated emissions and greenhouse gases in three different cases of cooling pump power requirements (5.4 kw, 2.24 kw, and 0 kw). As shown in the bottom part of Table 26, for a parasitic power of 5.4 kw, system operating power output is 16.6 kw, operating efficiency is 4.4%, and estimated fuel saved during the test period (Oct 2 Dec 9) is 1,347 gallons. For the case of parasitic power of 2.24 kw, system operating power output is 19.8 kw, system operating efficiency is 5.2%, and fuel saved is 1,607 gallons. For the case of no parasitic power, system operating power output is 22.0 kw, operating efficiency is 5.8%, and diesel fuel saved is 1,789 gallons. Similarly, the effect of pump parasitic power on emissions reduction can be found in Table 26. From the above-mentioned data, the effect of parasitic power of the cooling pump on system performance is significant. Table 27 shows the effects of parasitic power, yearly operation hours, and interest rate on system performance and payback period. Three different cases of parasitic power (5.4 kw, 2.24 kw, and 0 kw), two different cases of yearly operation hours (full year operation/8,760 hours and 7.5 months operation/5,472 hours), and two different cases of interest rates (0% and 10%) are displayed in Table 27. Finding system operating performance and payback period for different combinations of the three parameters is demonstrated using the following example: Based on Table 27, for the case of year-round system operation, 0% interest rate, and 5.4 kw parasitic power, the diesel fuel saved per year (applying the GM) is 10,375 gallons, and the payback period is 8.9 years. For a similar case but with a 2.24 kw parasitic power, the fuel saved is 12,370 gallons, and the payback period is 7.3 years. For another similar case with no parasitic power, the fuel saved is 13,770 gallons, and the payback period is 6.4 years. Reductions in emissions for the corresponding cases can also be found in Table 27. From these examples, it can be seen that the effect of cooling pump power on yearly system performance and payback period is also significant. For the case where a more effective cooling water pump (2.24 kw) and higher jacket water temperature (4.5 kw gain) are both available, the SFC may increase from kwh/gallon to kwh/gallon, which is an improvement of 2.5% in SFC (instead of 2.0% as is the current case in Tok). 99

112 3.11 Issues Encountered All the aforementioned analysis is based on the assumption of smooth field operation of the GM ORC for 1,137 hours and a total operation (laboratory and field) of about 3,000 hours. Based on its performance up to October 2, the system was considered by all parties to be reliable, and maintenance requirements were considered minimal. On December 5, ElectraTherm contacted AP&T to inform it that ElectraTherm had deduced an incompatibility between the lubricant used in the GM working fluid (refrigerant) and the screw expander s seal. ElectraTherm anticipated a possible refrigerant leak, and recommended that AP&T upgrade the expander and change to a different lubricant in the working fluid. Four days later, on December 9, the GM leaked working fluid, which led to a bearing failure at 2,800 total hours of operation. The operation and maintenance manuals for the GM included a material safety data sheet (MSDS) (Green Machine Manual Part 2, pages 13 30) for the working fluid that described the fluid as benign. However, employees exposed to the refrigerant experienced dizziness, headache, and nausea and were sent home for the day. According to the power plant personnel s communication with ElectraTherm, the problem is fixable. ElectraTherm recommends changing the screw expander, upgrading the lubricant, and replacing other related accessories. As of this writing, AP&T has a quotation from McKinley Service & Equipment Inc., ElectraTherm s Alaska vendor, to replace the expander and change the working fluid; the quotation totals about $30,000 ($16,000 is for parts from ElectraTherm; the balance is for labor and travel from McKinley). Despite its low operating hours, the GM is no longer within the time period of the warranty. It is worth nothing that after completion of the laboratory testing, and prior to moving and installing the GM in the Tok power plant, ElectraTherm encouraged the parties to replace the unit with ElectraTherm s current market model, offering it at a reduced cost. ElectraTherm reiterated that if the offer were declined and TCC and AP&T opted to install the current GM, ElectraTherm would not warranty it, as the time period of the original warranty had expired. 4 Lessons Learned This chapter summarizes some of the findings from the entire testing process, including findings from both laboratory testing and field testing. The findings may be useful for predicting the operating performance of the GM ORC system under different circumstances related to energy resources and the economic requirements of individual diesel power plants. The ORC system should be designed and installed with the following points in mind: 1. Local codes and ordinances may affect the selection of installation location and the associated evaluation plan and facility needed. 2. The installation and operating costs of the cooling system, which heavily depends on the capacity of the available cooling source, and the technology employed to utilize it, may 100

113 become one of the dominant factors affecting whether the ORC is economically effective. 3. The installation of the GM, in general, does not require specialized training beyond the knowledge base that is typically available at a diesel power plant, provided a heat source and cooling source infrastructure are readily available. 4. Based on the performance of the GM and the similarity of the GM to refrigeration systems, the expected maintenance is minimal. However, if newly developed technology is used in the system, reliability and maintenance may depend on the performance of the new technology. For this project, the screw expander is a new technology. Problems encountered during operation of the GM: 1. During commissioning, a system computer program imperfection led to an emergency shutdown of the ORC system. The shutdown resulted from trying to start the system under undesirable heat flow conditions. This shutdown may not cause any damage to the system, but it is a nuisance and is undesirable. This situation may be easily overcome by following the correct starting procedure or adding control capability to the installed heating loop. 2. A leakage of the refrigerant/lubricant mix occurred to the expander after 2,800 hours of operation. This problem appears to stem from a design flaw; the lubricating fluid was incompatible with the seals used in the screw expander. Currently, the manufacturer is producing an updated 65 kw ORC unit and no longer manufactures the 50 kw model used in this test. As of this writing, the manufacturer reports that it has addressed this issue and that 27 commercial units are in operation, with two machines at over 17,000 hours run time and an additional six at over 10,000 hours. However, this study team cannot make any comment about the performance or long-term reliability of any ORC system other than the unit tested for this report. Heating Source: At a given heat source temperature, the system efficiency does not vary significantly with heat source flow rate. Because of the GM s ability to work well within a range of flow rates, the GM ORC technology may be applicable to a broad range of diesel engine sizes and a wide range of varying load and jacket water temperatures. Test data gathered for this report may help in estimating whether the heat source condition of an individual diesel generator is appropriate for applying the GM for good results. As observed from test data, the GM may also have asymptotes in performance (e.g., GM net power output) versus heating fluid flow rate for specific heat source temperatures. This information may help in determining if it may be economical to install other heat recovery devices at the same time to avoid over-feeding the ORC and wasting heat energy. The high and low heat source temperatures used for this testing (225 F to 155 F) seem close enough to the heat source temperature limits of the GM. 101

114 According to this experiment, the residual heat could be further used if the heating fluid exiting the GM has relatively high temperatures. As observed from test data, there is a point where the power output of the GM no longer materially improves beyond a certain heating fluid temperature. Knowing the temperature beyond which no material benefit to power production is achieved would be useful for determining whether it is economical to install other heat recovery devices in addition to the ORC in order to avoid over-feeding the ORC with heat energy. The high and low heat source temperatures used for this testing (225 F to 155 F) seem close enough to the heat source temperature limits of the GM. Cooling Source: The greater the temperature differential between the heating fluid and cooling fluid, the more efficient the system operates. For the unit tested, there is no requirement for minimum cooling source temperature for the GM (besides freezing point of the cooling fluid). In general, the lower the temperature of the cooling fluid, the better the system performs. Pump power consumption is an exponential function of flow rate. A pump power consumption of 1 kw (for flow rate of 120 gpm) may not be a critical parameter based on GM gross output data (from about 11 kw to about 50 kw) obtained from the laboratory test. The effect of cooling water flow rate on efficiency was observed to be minimal for a large range of flow rates. Extremely low flow rates, which may drastically decrease system efficiency or cause damage, should not occur in normal operation and are not considered in this report. Considering the ORC s relatively low contribution to overall system efficiency for generating electricity, higher than necessary cooling water flow rate should be avoided because of the associated increase in parasitic power requirements. The information obtained from this test may be used as a reference for selecting an appropriate cooling water flow rate. Combined Effect of Heat Source Temperature and Cooling Source Temperature: According to the analysis results of laboratory test data, cold-water temperature has a greater impact on ORC performance when the hot-water temperature is low. This finding matches fundamental thermodynamic principles. Considering applications in rural Alaska, jacket water temperatures typically are between 175 F and 195 F (based on log data of a few village diesel generator sets); the effect of cooling fluid temperature may become a moderately influential factor on performance. As mentioned previously, for 120 gpm of cooling water flow, the estimated pump power (parasitic power) requirement is about 1 kw, which is low in comparison with the level of GM ORC output for jacket water heat application (between 11 kw and 40 kw), in general. Therefore, the use of 1 kw (or an appropriate extrapolation for a parasitic power estimate) for system performance and economic outcome evaluations for cases where cold flow rate is less than 120 gpm may not cause significant error. For cases where villages can only provide a cooling water flow rate at much less than 120 gpm (e.g., 80 gpm or less) or for cases that do not require 120 gpm for cooling due to the availability of very low-temperature cold water (e.g., 102

115 40 F well water), data obtained from this experiment may be used through cautious extrapolation. It is always advisable to use measured performance data directly obtained from cases involving low temperatures and flow rates, if possible. Match between the GM ORC and a Diesel Generator Set: This section suggests one approach for determining whether a specific village generator will benefit economically from applying the GM Block 1 ORC system: 1. Estimate fuel costs, capital costs, interest rate, and desired payback period. 2. Use Fig. 32 or Fig. 33 to determine the minimum power output requirement to obtain the desired payback period. 3. Estimate the available heating and cooling source conditions. 4. Estimate the parasitic power to operate the GM. 5. Use Fig. 34 to estimate the net output of the GM based on local hot-water and coldwater conditions. 6. If the estimated GM system operation net output (which is the GM output minus the parasitic power) equals or is greater than the minimum required operation power output for the desired payback period, the GM is recommended for the particular village power plant. If the amount of heat is much more than the minimum requirement for the payback period, a further study about how to use the waste heat more effectively may be needed. For example, a heat recovery system may be added for heating, or a different size of heat-to-power application may be considered, if it is available. 7. If the estimated GM power output is less than, but close to, the minimum required operation power output for the desired payback period, different combinations of hotwater and cold-water conditions may be modified to determine a good setup to achieve the desired payback. For example, the same amount of heat may exist in fluids of different temperature and flow rate conditions. Then Steps 4 through 6 must be repeated until an optimal solution is found. A few iterations of combinations of hot and cooling water conditions may help with reaching the final decision. 103

116 Fig. 32. Payback period at 0% interest rate on capital for different Green Machine ORC power outputs, fuel prices, and capital costs (for full year of operation) 104

117 Fig. 33. Payback period at 10% interest rate on capital for different Green Machine ORC power outputs, fuel prices, and capital costs (for full year of operation) 105

118 Fig. 34. Green machine net power output and net efficiency versus hot-water supply temperature An example that demonstrates the proposed procedure is as follows: Suppose a power plant has a hot jacket water flow of 125 gpm and 185 F and a cooling water source of 120 gpm and 50 F. We would like to determine the feasibility of using the GM ORC as its waste heat recovery system. Other known conditions and desires are that the cooling water pump power requirement is 1.5 kw for a 120 gpm flow rate, the payback period requirement is 10 years, the expected capital is $300,000, the interest rate is 10%, fuel cost is $5/gallon, and the system will operate year-round. Based on the given information of hot-water temperature and flow rate and cold-water temperature, the GM net output power (which can be found using Fig. 34) is 27 kw and the operating output power is 25.5 kw (27 kw 1.5 kw). Based on the expected capital cost and interest rate, fuel cost, and desired payback period, the required GM operating output power is about 24 kw (found from Fig. 33). Comparing the GM operating output power (25.5 kw) with the required power of 24 kw, it seems marginally feasible to adopt the GM system for jacket water heat recovery for this power plant. Since this calculation is based on water as heating fluid, if the heating fluid is changed to 60/40 PG/W, the GM operating output power needs to be modified accordingly. Then the GM will become marginally unfeasible at this power plant. 106

119 5 Summary After completing 50 hours of laboratory testing under different prescribed controlled conditions, and 1,000 hours of testing under full load, the Green Machine was successfully installed in the Alaska Power and Telephone power plant in Tok, Alaska. There, it ran an additional 1,137 hours before a failure occurred, taking the Green Machine out of service. The concept of a small-scale organic Rankine cycle (ORC) seems technically and logistically sound for Alaska, as shown by the performance of the Green Machine both in Tok and at the University of Alaska Fairbanks power plant. Under ideal conditions, the Green Machine generated close to nameplate capacity of 50 kw, and under field conditions in Tok, it produced a time average of 22 kw. Moreover, improvements have been made in the Green Machine design since testing. No statements can be made at this time about the long-term reliability of an ORC system as installed in Alaska, however, because the Block 1 Green Machine has only seen 2,800 hours of service. 107

120 6 References [1] Capture of Heat Energy from Diesel Engine Exhaust, DE-FC26-01NT41099, Submitted by Arctic Energy Technology Development Laboratory, Institute of Northern Engineering, University of Alaska Fairbanks [2] Alaska Electric Power Statistics Prepared by the Institute of Social and Economic Research, University of Alaska Anchorage, for the Alaska Energy Authority. [3] Leibowitz, H., Smith, I.K., and Stosic, N., Cost Effective Small Scale ORC Systems for Power Recovery from Low Grade Heat Sources, IMECE [4] Quoilin, S., Experimental Study and Modeling of a Low Temperature Rankine Cycle for Small Scale Cogeneration, University of Leige, [5] Kuppan T., Heat Exchanger Design Handbook, Marcel Dekker, [6] Muley, A., and Manglik, R.M., Experimental Study of Turbulent Flow Heat Transfer and Pressure Drop in a Plate Heat Exchanger with Chevron Angle, ASME Journal of Heat Transfer, February 1999, Vol. 121, pp [7] Ayub, Z.H., Plate Heat Exchanger Literature Survey and New Heat Transfer and Pressure Drop Correlations for Refrigerant Evaporators, Journal of Heat Transfer Engineering, 2003, Vol. 25(5), pp [8] Selvam, M.A.J., Senthil, K.P., and Muthuraman, S., The Characteristics of Brazed Plate Heat Exchangers with Different Chevron Angles, ARPN Journal of Engineering and Applied Sciences, December 2009, Vol. 4(10), pp

121 Appendix A. Heating and Cooling System Design The heating and cooling systems, designed by Alaska Power and Telephone, are shown in the schematic on the following page. 109

122 110

123 Appendix B. Green Machine Startup Parameters, 27 Aug 2013 Startup data w/no. 4 and No. 9 diesels operating; all others bypassed Machine hours at startup 1146 HOBS meter HMI screen GM KW 27 GM net KW 24.5 HW flow GM off 176 gpm HW flow GM on 148 gpm HW psi into GM 11 HW psi out of GM 2 HW temp into GM F HW temp out of GM 170? Cold water flow 65 gpm Cold water psi into GM 0 Cold water psi out of GM 0 Cold water temp in 34.8 F Cold water temp out 70.2 F Unit 4 Load 733 kw Engine in F Engine out F Heat rec in 179 Heat rec out 185 Rad fan 8 hz Unit 9 Load 560 kw Engine in F Engine out F Heat rec in 179 Heat rec out 185 Exch water cooling 50 %? HR system after ORC 164 F HR pump psi 25 psi 111

124 Tok heat recovery circulating pump pressures at 60 hz 5 hp Scott 57 pump 4" x 3" 6.88" impeller x indicates closed bypass valve o indicates open bypass valve No. 3 x x x x x o No. 4 x o o o o o No. 5 x x o o o o No. 9 x x x o o o No. 8 x x x x o o No. 7 o o o o o o ORC o o o o o o Pump psi Setup screen Password 1248 Operating parameters Delta T shutdown = 50 F Min Hz pump = 22 Max Hz pump = 60 Min kw net = 5 kw Max exp pressure = 175 psi Exp differential min = 32 psi Ongrid rpm = 1730 Max kw = 50 Hot shutdown = 245 F Pwr factor =.95 FL amps = 75 Line volts = 500 Start parameters Time btw starts = 30 sec Start delay = 1 min Hot min temp = 170 F Hot max temp = 240 F Hot water delay = 20 min Min start delta T = 80 F KW start parameters Y = mx+b Offset = 31.9 kw Slope = Options screen Colt switch = installed Cold flow = not installed Cooling type = liquid Grid protection relay = installed 112

125 Gen nameplate rpm = 1835 Encoder pulse/min = 600 Defaults No change Veris setup (send to enter changes) System type = 31 CT ratio = 3 Ph loss = 10% Ph imbalance =

126 Appendix C. Materials Used in Green Machine Installation GM lifting equipment 1 10,000 lb telescoping forklift 1 Power plant overhead bridge crane (10 ton) 1 Lot Misc. nylon slings and rigging Tools and equipment 1 Lot Misc. hand tools, wrenches, screwdrivers, hammers, punches, etc. 1 Lot Tape measures, laser level, squares, string line, plumb bob, etc. 2 MIG wire feed welders, PPE and welding supplies 1 Oxyacetylene torch kit with oxygen and acetylene bottles 2 Welding tables, pipe stands, and pipe clamps 1 Pipe threading equipment up to 2" 1 Horizontal 7" x 12" wet band saw w/extra blades 1 4" portable electric band saw w/extra blades 1 Drill press and drills 1 Bench 8" bench grinder 2 4" and 9" angle head grinders 1 1/2" portable electric drill and bits 1 Electric hammer drill and bits 1 6" electric core drill and concrete core saw 1 Hole saw and electrical KO set 1 Coolant transfer system with pump, reservoir, hoses, etc. 1 Coolant storage tote tanks or barrels Drain field 1 Backhoe 1 Dump truck 200 ft Typar geofabric 180 ft 4" perforated pvc drainpipe 1 Lot Misc. 4" pvc pipe, elbows, tees 1 Lot 4" pipe insulation Heat exchangers 3 4 pack (SWEP B200Tx160) plate exchangers from ElectraTherm 2 Young 810-TR shell and tube exchangers (existing) 1 Polaris S41-IS16-82-TK plate exchanger (rebuilt existing) Thermostatic valves 2 AMOT thermostatic valve 4" flanged 1 FPE thermostatic valve 5" flanged Heat recovery isolation valves 11 4" flanged ball valves w/manual operators 4 3" threaded ball valves w/manual operators 10 1/4" automatic vent valves 5 Pressure relief valves 114

127 Pipe 105 feet 3" S10 steel pipe 231 feet 4" S10 steel pipe 100 feet 5" x.188 wall steel pipe Pipe fittings 2" and smaller, valves, nipples, elbows, couplers, unions 1 Lot Thread-o-lets 6 3 x 12 threaded nipple 11 3 x 3 threaded nipple 8 3" Thread-o-let 18 3" slip on weld flange 2 3" threaded flange 14 3" 150 lb flange bolt and gasket kits 34 3" 90 weld elbow 6 3" 45 weld elbow 8 3" threaded plug 14 4" x 4 threaded nipple 2 4 x 6 concentric weld reducer 10 4' weld Tee 48 4" slip on weld flange 38 4" 150 lb flange bolt and gasket kits 56 4" 90 weld elbow 9 4" 45 weld elbow 10 4" x 3" concentric weld reducer 4 4" 150 lb blind flange 4 4" 150 lb threaded flange 16 4" weld cap 4 5" weld tee 16 5" slip on weld flange 8 5" flange bolt and gasket kit 10 5" 90 weld elbow 6 5" 45 weld elbow 7 5" x 4" concentric weld reducer 4 5" x 3" concentric weld reducer Expansion fittings 2 5x6 flanged spherical flex 8 4x6 flanged spherical flex 2 3x6 flanged spherical flex 4 4" Aeroquip Flexmaster coupling 2 3" Aeroquip Flexmaster coupling 2 2" Aeroquip Flexmaster coupling Structural support materials 200 ft Strut 1-5/8 x 1-5/8 115

128 1 Lot Strut pipe clamps and fittings 21 ft 8" sch 40 pipe 84 ft 3" sch 40 pipe Miscellaneous hardware 1 Lot Bolts, screws, anchors 1 Lot Pipe dope, saw blades, drills, paint, etc. Electrical materials 1 3 phase 7 jw meter socket 1 3 phase 100 amp 600 volt fused disconnect switch amp 600 volt time delay fuse 20 ft 1 1/2" rigid conduit 50 ft 1 1/2" LT flex conduit 2 1 1/2" LT flex connectors 3 1 1/2" LB with cover 280 ft #2 THHN copper wire 1 Lot Lugs, straps, anchors, bushings, lockrings 1 Lot DSL equipment, Cat V cable, connectors, power supply 1 Lot PLC software for programing and remote access Heat recovery fluid 385 gal Premix 60/40 propylene glycol 220 gal Premix 50/50 long life diesel antifreeze Working fluid equipment lb freon storage bottles 1 Lot Tools to test, transfer, evacuate, and fill freon (not used in Tok) Documentation ET install, O&M manuals 116

129 Appendix IA Survey of Low Temperature Heat Engine Companies (2008) Introduction Organic Rankine Cycle. 90 ORC Companies Kalina Cycle Kalina Cycle Companies Summary of Survey Results. 96 References

130 Introduction In today s energy crisis people are looking at almost any means in hopes to lower their energy bill. Some of these technologies are already in the market as cost effective options and others still need more development before they make economic sense. Waste heat recovery at an industrial level is a proven technology with many instances of it being successfully used to power plants operations more efficient. Recovering the waste heat from diesel engines, on the other hand, is not a mature field. There are companies out there looking to address this problem but they are not quite there when it comes to a cost effective solution. The technology is there but the main problem is in developing a product that produces a positive net energy. Because the waste heat is a low energy source it is easy for parasitic power from the recovery unit to be higher than what is being produced. For purposes of this report three main technologies will be examined, viz the Organic Rankine Cycle, the Kalina Cycle and the Stirling Engine. The basics of these technologies will be examined in brief as well as a more detailed look at some of the companies involved in each and what the status of their product. A survey summary is attached to the end of this report. 118

131 Organic Rankine Cycle Diesel engines lose about 30-40% of the input energy in the exhaust. Organic Rankine Cycles (ORC) look to take this low-temperature heat source and improve the overall efficiency of the engine or produce electricity directly. Some estimate the efficiency of an ORC unit to be about 10-20%, which would improve the overall engine efficiency by about 3-5%. An ORC unit works very much like that of a traditional steam Rankine Cycle: a working fluid is pumped through a heat exchanger where it is vaporized and passed through a turbine and then re-condensed, and then the process is repeated. The big difference is in that of the working fluid. The traditional Rankine Cycle uses H 2 O as its working fluid with has a high boiling point. Since the waste heat from the engine is a low-temperature source H 2 O will not work. So an ORC unit uses some organic fluid as its working fluid. Choice of the working fluid will depend on the heat source and some companies have their own special blend that allows them to capture heat at low temperatures. ORC units have been installed in various places. Ormat, a leader in the ORC technology, has units in North America, Europe and Asia 1. These units, however, are being installed on an industrial scale, ranging from 200 kw to 22 MW. For use on a diesel generator one would be looking for a unit less than 100 kw for the most part. Not many units have been built at this size, although there are companies with them in the works. In order to get a better grasp of the state of this technology it is will be helpful to look at these companies individually. Editorial comments will be left out as much as possible with the hope that the facts gleaned from the companies websites and through personal contact will be enough to show how close we are to seeing an ORC unit installed with a diesel generator as a way to capture waste heat. ORC Companies After a thorough web search, seven companies were found that use ORC technology in their products. Some of these companies specifically targeted the waste heat from diesel engines while others were broader in their application. The companies looked at were Global Energy, Barber-Nichols, ElectraTherm, TransPacific Energy, Deluge Inc., Ormat, Turboden, UTC Power and GMK. Each company will be looked at in turn. Global Energy Global Energy, a Madison, WI company under the leadership of Greg Giese, has developed the Infinity Turbine. This is an ORC turbine built for waste heat and geothermal applications. While there are potentially numerous uses for this turbine one that is specifically being targeted is the diesel engine exhaust. According to the website, the Infinity Turbine consists of a single skid-mounted assemble that fits in the standard 20 or 40 foot ISO standard shipping container. All the equipment required for the power skid to be operated (i.e. heat exchangers, piping, working fluid feed pump, turbine, electric generator, control and switch-gear) fit into the container. In early mid-july 2008 a 30 kw unit and an 80 kw unit were being built in Toronto. By July 29, 2008 the 30kw unit had been sold to a geothermal application in Casper, 119

132 WY. It is hoped that in a month or so data will be ready for analysis. While the website does list some performance specifications this are only theoretical calculations. Hopefully this test site in WY will show that the theoretical calculations where right. A price of $60,000 was quoted over the phone for the 30 kw unit although it is not know how much it sold for to the geothermal plant in WY. Also a delivery time of 11 weeks was quoted, presumably from day of purchase. Barber-Nichols Barber-Nichols is a leader in the field of turbomachinery. Thermodynamic Cycle Systems including the Organic Rankine Cycle and the Steam Rankine Cycle are part of their core competency. They have already built waste heat applications but on an industrial scale. These units are much too large to be used with diesel engines but they demonstrate that Barber- Nichols is competent in the technology. They are currently working with some Canadian companies, however, to develop smaller units that could be used with diesel engines. Since these are custom designed units based on the size of the generators an interested buyer would need to send in specifications regarding their application. During a phone call with a company representative a price range from few hundred-thousand dollars up to $1M was put forward as an estimate of the total cost, including initial research, NRE, and unit itself. ElectraTherm Nevada based ElectraTherm has recently come out with their own ORC unit that captures waste on a smaller scale. They have plans for units ranging from kw. It is not clear how many units they have indeed sold and that have been field tested, but one unit at SMU in Texas is known to be running and undergoing field tests, and it is most likely that this is their only one. This unit was demoed at the Geothermal Conference at SMU in June of Dr. Dennis Witmer from the Alaska Center for Energy and Power at the University of Alaska, Fairbanks traveled there to get a first hand view of the unit. Overall, the demonstration was not very impressive according to Dr. Witmer. The unit was very loud and there seemed to be some grinding sound, as well some of the parts looked like they were used parts. More importantly, though, it is unclear whether the unit is making any net power when parasitic power is taken into account. On a follow up phone call with Michael Paul from SMU in July 2008 he mentioned that an availability test had not been done since it was not a waste heat application. So it is unclear how long these units can be up and running before maintenance needs to be done. Currently, SMU is not planning on buying the unit after what they saw in the tests they did. The latest price that was available was $2400-$2700 per kw. TransPacific Energy This Nevada bases company is developing a heat recovery/energy conversion system using ORC technology. They claim on their website that their system can use heat sources at temperatures as low as 80 o F and up to 900 o F. This is a larger range than most other ORC units which usually limited to a low temperature of around 200 o F. They can do this, they say, because of their own specially designed refrigerant. They believe that this refrigerant is their key advantage over competitors. In an response from Jim Olsen, a company representative, it was revealed that they do not actually have any units installed yet as they are a new company. This most 120

133 likely means that any specifications they are stating are simply theoretical calculations and not actual measured results. The company website says this about the sizing of their units: Systems are sized from 20 kw up to 20 MW+ modules, containing all the equipment required for the units to be operated (i.e. heat exchangers, piping, working fluid feed pump, turbine, electric generator, controls and switch-gear). Larger units are composed of multiple modules, pre-assembled at the factory. The price for a 115 kw unit was quoted as running up to $250,000 for the complete unit. Delivery time is expected to be 6-9 months depending on their fabricator, Concepts NREC, out of Massachusetts. Deluge The Deluge Natural Energy Engine is not an ORC unit but rather a thermo hydraulic engine that produces mechanical energy by heating a fluid so that it expands and moves a piston. The heat source can be solar, geothermal, or waste heat. The main components of the engine are the piston/cylinder and the heat transfer system. The cylinder contains the piston and the working fluid, usually C0 2. The heat transfer system is made up of heat exchangers and a system to circulate the heat exchange fluid, usually H 2 O. A three step process creates a back and forth movement of the piston, in turn generating mechanical energy. The technology has been independently verified by university and government studies. In an from July 2008 they reported they finished building their first 250 kw unit and are shipping it to a jobsite in Hawaii for installation. They claim that 250 kw can be produced from a 150 gpm flow of water at 190 o F and that temperature drop through their engine would eliminate the need for the radiator cooling. These flow rate and efficiency numbers are based on calibrated tests in the shop. The price for a 250 kw engine/generator set is $400,000 when buying one at a time. The delivery timeframe is 90 days from purchase order. Ormat Ormat is the world leader when it comes to ORC technology. They have successfully installed ORC units all around the world. They specialize in Geothermal Power, Recovered Energy Generation, and Remote Power Units. Their units range from 200 kw to 22 MW for the Recovered Energy Generations units that cover waste heat recovery. Their remote power units range in size from watts. When given the specifications for a 125 kw diesel generator the company said that the application was too small for their Recovered Energy Generation units. Currently, Ormat is looking into developing a smaller unit that could be used with diesel generators, but as of now none are available. They will let ACEP know if they decide to pursue a smaller unit. Turboden Turboden is an Italian company that specializes in ORC technology. They have Combined Heat and Power systems in set sizes ranging from 200 kw to 2000 kw. They also have Heat Recovery systems that come in set sizes ranging from 500 kw to 1500 kw. They are also able to build custom sizes but currently do not manufacture any under 500 kw for applications requiring a single unit. They have installed many units, mostly in Europe and in the biomass industry. They company did not respond to s so there is little details as to cost and delivery time. 121

134 UTC Power UTC Power is a division of United Technologies Corporation based in Connecticut. They provide environmentally responsible power solutions and recently have developed an ORC unit. The PureCycle Power System is an electric power generating system and runs off of any hot water resource at temperatures as low as 195 o F. The hot water can come from a geothermal source or some other waste heat source. Currently this ORC unit is sized at 280 kw (gross) of electrical power. One of these is commercially running at Chena Hot Springs Resort in Alaska and they have sold 75 units to date. The average price per kw is $1250 with a delivery time of 8 weeks. They are currently working on a 1 MW unit to be completed in the early part of There are also plans underway for a possible smaller unit with the size to be determined. GMK Gesellschaft für Motoren und Kraftanlagen (GMK) is a leading ORC-module developer and producer in Europe. They have three different products bases on application. Their INDUCAL is an ORC used in power plants to recover waste heat. The electrical output of these units varies from MW. The two other products are GEOCAL and ECOCAL which use geothermal energy and biomass respectively to produce electrical power in similar amounts as that of the INDUCAL. They have various installations of all three products that are currently running in Germany. The company seems to be focusing on larger, industrial power plants rather than on units that would be useful in heat recovery from small diesel generators. The Kalina Cycle The Kalina Cycle was developed by the Russian engineer Aleksandr Kalina in the early 1980s. It is a thermodynamic cycle that allows for the converting of thermal energy to mechanical power which can then be used to create electrical power. It is similar to the ORC in how it works with one major difference. The Kalina Cycle uses a binary working fluid. Usually the fluid is a mixture of ammonia and water. This allows for a broader range of boiling points because ammonia has a much lower boiling point than water. Studies have shown that the Kalina Cycle performs better than the Organic Rankine Cycle at moderate pressures. Some believe that it is in general more efficient than the Organic Rankine Cycle, but more testing needs to be done to support this claim. Commercially tested Kalina Cycle units are not common with only a few ever built. One is used in a geothermal power plant in Iceland. There are two plants in Japan, one a waste heat plant and the other a waste-to-energy demonstration plant where the Kalina Cycle was used. The first ever demonstration of the technology was in California, which proved to be very successful. Currently the geothermal plant in Iceland and the steel mill waste heat power plant in Japan are still running. 122

135 Kalina Cycle Companies The companies involved in this technology have gone through various mergers and acquisitions so that today there is currently only one company with technological rights to the Kalina Cycle. Global Geothermal was created in 2007 and now owns all those rights as well as the over 200 international patents associated with the proprietary technology. They are a licensing company and sell the rights to the technology to companies wishing to use it in power plants or other applications. The following is a more detailed look at the companies that made up Global Geothermal and some of the current companies who are licensing the technology. Original Company: Exergy In 1992, Exergy, the company founded by the inventor of the Kalina Cycle, Aleksandr Kalina, saw the first Kalina Combined Cycle power plant start running at Canoga Park, CA. The 6.5 MW Canoga Park power plant was built as a demonstration plant to show that the technology is commercially viable. The plant ran up until 1997 and was deemed a success, as much useful data was gathered showing that the Kalina Cycle is an efficient way of generating power. Run as both a waste heat power plant and as a combined cycle power plant, it was designed to be tested at extreme conditions. In total it logged about 9,000 hours of operation over its five year life span, which corresponds to about 21% availability. [2] In 1997 Exergy teamed up with Ebara Corporation, a Japanese company that specialized in advanced industrial systems, to build another demonstration plant in Fukuoka, Japan. [3] This demonstration was a 4.5 MW waste-to-energy plant that ran from and was again seen as a success. In 1998 another project in Japan was begun at the Sumitomo Metals Kashima Steelworks in Kashima, Japan as a waste heat recovery application. This was the first commercially installed Kalina Cycle power plant and it currently is still running. The 3.4 MW plant has had great availability and runs off 208 o F (98 o C) hot water. Combining of Licensees It is not very clear how all the mergers and acquisitions worked out and who all the parent companies are that own the companies with the rights to the Kalina Cycle, but an attempt will be made to shed a little light on where this industry has come from and where it is now. Recurrent Engineering Recurrent Engineering became a global licensee in They owned by Wasabi Energy. Not much is available as too what they did with the technology from 2002 to 2007, but in 2007 Global Geothermal was created to buy up Recurrent Engineering and Exergy and to consolidate other companies licensed to use the Kalina Cycle worldwide. Wasabi had been in some disagreements with their co-venturer, AMP Capital Partners LLC as to how to handle the Kalina Cycle technology and the solution was to incorporate Global Geothermal, which is owned 70% by Wasabi and 30% by AMP 123

136 Global Geothermal Due to these changes in the industry, Global Energy now owns all the rights to the Kalina Cycle as well as the over 200 international patents associated with it. The company s primary business model is to license the technology for up-front fees as well as follow-on fees bases on size and use of the licensee s power plant. They also provide engineering services, equipment procurement, and project management services through their subsidiary Recurrent Engineering Current Licensed Companies A number of companies have received licenses from Global Geothermal or the previous companies with licensing power, and are working with Global Geothermal to make the best use of the Kalina Cycle technology. Exorka A geothermal energy company in Iceland, Exorka, specializes in low-temperature geothermal sources. In 1999 they built the first geothermal power plant using the Kalina Cycle in Husavik, northern Iceland. According to their website the plant produces 2MW of power from a flow of 90 kg/s at 248 o F (120 o C) geothermal brine. Exorka acquired the rights to the Kalina Cycle in 1999, with these rights extending to Iceland and most of Western Europe. They are currently in the late development and finance stages of five projects in Germany. Each project is about a 5 MW binary geothermal power plant. It is expected that these projects will come online in Geodymanics Geodynamics is an Australian geothermal energy company that focuses on hot fractured rock geothermal energy. In 2004 they acquired the rights to use the Kalina Cycle technology in Australia and New Zealand. As of now they do not have a plant running that uses the technology. In 2007 they merged with Exorka to form Exorka International Limited. Raser Technologies A publically traded technology licensing company, Raser, was formed in 2003 with a goal to improve the efficiency of rotating electromagnetic and heat transfer applications. They have also got into the geothermal energy industry as well as into waste heat recovery with the goal to make it as efficient a process as they can using their skill and technological knowledge in heat transfer and motors. To this end they become a Kalina Cycle technology licensee in Raser controls large amounts of geothermal land in Nevada and Utah where they are developing a 20 MW binary geothermal plant. They also are working on a 10 MW geothermal plant in New Mexico. It is hoped that these plants will go online in Raser not only has Kalina license rights to geothermal applications but also to industrial waste heat applications. They are working closely with Recurrent Engineering to use Kalina technology in waste heat projects in both the US and Europe, primarily in the cement industry. Siemens A well known Germany engineering firm, Siemens purchases Kalina Cycle license rights in 2000 and is able to use them for geothermal power projects in Germany, up to a total of 10 MW in 124

137 size. Even before it gained the rights to the technology, Siemens has been evaluating and testing it. They have been working with Recurrent Engineering to complete the construction of the first Kalina power plant in Europe. The 3 MW plant was completed in December 2007 as a binary geothermal project built as a turn-key project for the City of Unterhaching, Germany. Siemens will provide long term O&M for the plant. The plant was expected to go through commissioning and acceptance testing during the first quarter of 2008 and come online in mid- June It is expected to have a lifetime of 30-plus years. Siemens is looking to add more plants in the next decade and currently has two more contracts for turn-key power plants in Germany. Other Companies Related to Kalina Cycle There are a few companies that do not fit nicely into the above breakdown but that are either using Kalina Cycle Technology or something very similar. They are discussed below. Energy Concepts Energy Concepts is a Maryland based engineering company that focuses on heat-activated absorption systems and the associated fluid contact equipment. The Absorption Cycle was invented by Ferdinand Carre in The idea is to take heat out of a system by running a cycle that uses a heat input. When ammonia is absorbed in water the vapor pressure decreases, and so according to the laws of thermodynamics, the temperature will drop also, ceterus paribus. The absorption cycle has the benefits of requiring little electric input and using natural substances, ammonia and water, instead of halocarbons. Although this in not a true Kalina Cycle, it takes advantage of the properties of the binary fluid made of ammonia and water. Energy Concepts has a way of also using this cycle to convert exhaust heat from prime movers to electric power. The Heat Activated Dual Function Absorption Cycle is capable of taking a heat source ranging from 250 F to 750 F to and converting it to electric power, refrigeration and/or air conditioning. According to the company website when using the Absorption Cycle a 1 MWe gas turbine with an exhaust temperature of 750 F can produce 400 kw. They claim this cycle works well with distributed power generators sized from 1 to 15 MWe. Rexorce An Ohio based thermal energy company, Rexorce is looking to find ways to better harness existing thermal resources and also to recover waste heat and use it in an efficient manner. They have developed a thermal engine, Thermafficient, which is designed to recover thermal energy from a range of sources and convert it into electric power, cooling and heating. Rexorce s engine uses supercritical CO 2 and other working fluids for their power generating cycle. This binary fluid cycle has many of the same benefits as that of the Kalina Cycle. According to information from a company contact, they are working on a 250 kw generator and claim 25-30% efficiency when working in diesel exhaust at 500 o F. As of July 2008 they are still working on the expansion device and hope to have it completed in a few months. The price for the unit is expected to cost less than $1500 per kw. With a higher volume of orders that price could drop significantly. 125

138 Survey Summary Company Unit Price $$/kw Delivery Contact Info Current Status/Notes ORC Units Global Energy Barber- Nicholes ElectraTh erm TransPaci fic Energy Delgue Inc. 30 kw $60,000 $2, weeks Greg Giese; ine.com 80 kw No price quote custom $200K - $1M kw Ormat > 200 kw Turboden >200 kw GMK MW UTC Power Depends on size 30 kw unit sold and should be up and running within a month. Check back in early September 2008 N/A 11 weeks As of July 14, 2008 being built in Toronto N/A N/A David K.; [email protected]; $2,400- $2,700 N/A 115 kw $250,000 $2, weeks Bill Olsen; bolson@electrat herm.com Jim Olsen; jolson@transpac energy.com 250 kw $400,000 $1, weeks Brian Hageman; bhageman@delu geinc.com; No price quote No price quote N/A N/A Colin Duncan; cduncan@ormat. com Would require a custom design that would take time. It is unclear how long this would take and when a unit could be delivered Much to prove before it becomes commerically viable No units installed anywhere yet. Only designs so far. Not ORC but thermal hydraulic engine. Built first 250 kw unit and is being installed in HI now. As of now recovering heat from a diesel generator is not an application they have units for. They are looking into it though for the future. N/A N/A [email protected] Have not been able to get in contact with company so details are scarce. No price quote 280 kw 350,000 $1,250 8 weeks [email protected] om N/A N/A [email protected] Deal with industrial power plants Commercial unit of Chena Hot Springs unit. 75 units sold to date. Working on 1 MW unit to be complete in early '09 and plan on possible smaller unit. 126

139 Kalina Cycle Units Energy Concepts Rexorce 400 kw absorpt ion cycle 250 kw Therma l Engine No price quote No price quote N/A N/A Don Erickson Not clear if a specific unit is available or if their product would be custom designed per application $1,500 N/A Michael Gurin Modified Kalina Cycle. Expect to be completed in a few months 127

140 References 1. Ormat [cited 2008 July 11]; ORCs for geothermal and waste recovery.]. Available from: 2. Kalina History [cited; Canoga Park power plant info]. Available from: =True&shareprice=Hide. 3. "Ebara and Exergy team to build first Kalina Cycle Power Plant; The Largest Japanese Environmental Engineering Company and California Energy Technology Company Begin Construction on First Kalina Cycle Power Plant in Japan". Business Wire. Feb 4, FindArticles.com. 13 Aug

141 Appendix IB Survey of Low Temperature Heat Engine Companies (2010) Intro: After a complete web search, 16 different companies were found to either produce ORC waste heat recovery units, or to have a plan to produce them in the near future. While the products themselves mostly focused on recovering waste heat some sort of power generation source, a lot of the companies only build units on a very large scale designed with multiple waste heat sources in mind. The companies reviewed were: Global Energy, Barber-Nichols, ElectraTherm, TransPacific Energy, Deluge Inc., Ormat, Turboden, GMK, UTC Power, BEP, KGRA Energy, BIOS Bioenergie, Calnetix Power Solutions, Cryostar, Energy Recovery Systems, and Energy Concepts. A general overview of the companies followed by a technical summary of their products follows: Barber-Nichols Barber-Nichols is one of the leaders in the ORC market for industrial sized ORC s. When communication was pursued, it was revealed that they are only interested in industrial sized ORC s at the moment and have no plans to produce anything small (10-100kw). They mention on their website that they have three major units in use right now, a 500kW unit, a 700kW unit, and a 2MW unit using waste heat and geothermal resources. BEP Energy BEP (Burke E. Porter) Energy is a subsidiary of BEP Corp. based out of Belgium. BEP Corp. produces a large variety of automotive and industrial testing equipment. BEP was contracted by ElectraTherm to produce the Green Machine, a 50kW ORC unit that recovers waste heat through thermal oil or jacket water. BIOS Bioenergie BIOS Bioenergie is an Austrian company that specializes in producing energy from biomass and energy efficiency upgrades. They seem to produce a variety of different products, but on a custom built basis. No specific information on what the sizes of their products are/could be was given when asked. Calnetix Power Solutions Calnetixis a company based out of California that specializes in motors and generators, but has a subsidiary Calnetix Power Solutions that produces the Clean Cycle ORC unit. The unit is in use in three different locations around the world. Unfortunately, a large generator is necessary to produce enough power to run the Clean Cycle ORC (approximately 2.5MW). There is no such generator in Alaska, which could present a problem for use. 129

142 Cryostar Cryostaris a company that specializes in industrial gas, clean energy, and hydrocarbon applications and is based out of France. On their website they advertise about an ORC unit that could run off of geothermal energy. However, there are no links to any specifications, pictures, or further applications. After further investigation, it was revealed that the company does not actually have a product, but they do have plans to produce one. There was no size estimate as of now. Deluge Inc. Deluge is a company that makes one model of ORC unit and a lot of information about how their product is better than any other ORC unit. However, no specifications or pictures could be found. When asked if a product had been produced yet it was revealed that they have an idea for an ORC but no actual ORC in production. There has been no change in products or technology since the last survey was completed in ElectraTherm ElectraTherm is a company that has continued to grow and expand over the years and seem to know what they re doing with ORC technology. They patented the Green Machine and have since installed it at a couple different sites. They claim that the reason their product is superior to other ORC s is that they use a Twin Screw Expander which is more efficient than other expanders used in industry. Global Energy Global Energy is a company that specializes in building Organic Rankine Cycle (ORC) units ranging from a power output of 10kw to 250kw. The contact at the company Greg Geise mentioned in some contact that he could potentially build an ORC unit anywhere in between his smallest unit (IT10-10kw unit) to the largest unit (IT kw unit) He currently has a working model of the IT10 and has some video online displaying the power output. Global energy is unique in the way that they create their own turbines for their ORC s called the Infinity Turbine. This turbine is a modular turbine that is easily assembled at the site of the ORC unit. GMK GMK (GesellschaftfürMotoren und Kraftanlagen) is a German company that specializes in ORC units in waste heat, geothermal, and biomass applications. They claim that they can produce a system that will use the waste heat from anywhere from a 500kw to 5MW Generator and have case studies that prove they have a product that is in use. Unfortunately, as of yet, no units have been installed or shipped to the United States. KGRA Energy KGRA Is a more unique company than most of the companies that produce ORC units. KGRA is willing to install the unit for you; however, you are not the sole owner of the unit. KGRA will continue to operate the unit and sell you power for a discounted price. Instead of free power 130

143 you will receive discounted power. The specifics of the units they could install were vague but it seemed like the mission statement was directed at an industrial source. Ormat Ormat is the industry leader in the field of ORC units, however, they have discovered that larger products are more marketable and efficient and directed all their efforts towards large scale industrial applications. Currently they have no plans to design or look into a smaller ORC unit that could be applicable in rural Alaska. TransPacific Energy TransPacific Energy is a company based out of California that specializes in ORC units in waste heat recovery and geothermal applications. Their claims are that their working fluid is the reason their product is so much better than the competition. They provide plenty of specifications on their working fluid. However, there are no specifications on their actual product. Turboden Turboden is a company that is based out of Italy and sells a number of ORC units. They are known throughout Germany, Austria, and Italy as the world leaders in ORC technology. At the current moment they have nearly 150 units installed worldwide including a unit in Canada. Unfortunately the scale of the units is designed for a large industrial application. However, plans to reduce the size of the units are in place. UTC Power UTC (United Technologies Corporation) is the mother company of Turboden and Pratt & Whitney who produce the pure cycle ORC unit. The pure cycle unit is an ORC built for an industrial application and does not scale down very well. As Pratt & Whitney are a turbine company, another attempt at a smaller ORC is unlikely. 131

144 Appendix IIA Preliminary Line Diagram of the Testing System 132

145 Appendix IIB Preliminary Components Selected for the Testing System Component Size Reason for selection Green machine (GM) Max power output: 50kW Min power output: 5kW Hot water supply temperature range: 180 o F to 250 o F. Cold water supply temperature range: 40 o F to 110 o F Hot water boiler Cooling tower VFD pump for hot water loop BTU meters Cold water pump 3-way butterfly valves Check valve Boiler was rated for 2400MBTU/hr output with hot water temperature variation from 155 o F to 235 o F. Delta cooling tower model number T-150I, Induced draft cooling with cooling capacity of 150ton (1800MBTU/hr) i. Bell & Gossett 20hp pump (1750rpm) rated for VFD operation. ii. Rated for 250gpm and head of 116feet of water Only commercial unit available in market at that time which can recover heat to power from stationary diesel engine jacket water temperatures (200 o F) i. Based on the heat load requirement for GM. At this stage the GM was supposed to be 50kW. At 7.0% efficiency of GM the GM would need 2400MBTU/hr of heat input. ii. Ability to change hot water temperature supply to GM from 155 o F to 235 o F i. Based on the cooling load requirement for GM of 1800MBTU/hr i. Pump was selected based on the pressure drop in the hot water piping calculated using pressure drop across various components (e.g. boiler, 4 pipe, GM heat exchanger etc.) ii. VFD pump is used to test the GM for various flow rates of hot water. Detailed description of size and reason for selection is given in Table-2 133

146 Appendix IIIA Methodology Proposed for Stage 2 and Stage 3 Modeling Stage Two: The second stage is to use measured data from experiment to fit the values of system parameters of the first stage model. Data to be used for fitting include inputs and outputs of the testing system. Inputs include, at least, inlet temperatures and flow rates of the heat source and heat sink heat exchangers. Outputs include, at least, outlet temperatures of ORC heat exchangers and electrical power generated from the ORC system, of which the effect of mechanical to electrical power conversion efficiency is included. If other measured data become available (i.e. obtained from meters come with the ORC system or meters which are allowed to installed to the ORC system), they will also be used to fit the values of the parameters. The fitted parameters values will then be used for model simulation of the test system under extra operation conditions of heat source and heat sink. The simulation results for the extra operation conditions (using fitted parameter values) will be compared with experimental data under the respective testing operation conditions to investigate the discrepancies and to qualitatively determine the feasibility of modeling process. If the model is qualitatively feasible, but results predicted need to be improved, a more detailed model will then be adopted as the third stage of the simulation process. Stage Three: A more detailed model of an ORC system may need more than just four major subcomponents (i.e. expander, condenser, evaporator, and pump) as used for stage one modeling. Pipes and control elements (e.g. valves, fittings, and expansion tank) may affect heat transfer and pressure distribution and distribution of liquid and vapor in evaporator, condenser, and expansion tank. Therefore, pipes and control elements may affect performance of working fluid in the system and need to be included in system modeling. In addition, each of the sub components may be further divided into detailed elements to get better simulation results. An expander may be further divided into its function elements, which include pressure drop element as nozzle effect, vapor cooling down element for working fluid before entering into the screw channel, isentropic expansion element accounting for the working fluid expansion in the channel between the screws, further expansion element due to ambient pressure, leakage element for leakage of working fluid from high pressure region to low pressure region, and existing cooling down element. In addition to the six elements, mechanical loss due to friction also needs to be considered for obtaining net expander work. Inside the condenser, working fluid may experience three different conditions: vapor phase, mixed phase, and liquid phase. To better represent the performance of the condenser using simple mathematical model, a multi-section model may be needed with each section having its own constant heat transfer properties to approximate the heat transfer behavior of the particular zone. The spaces occupied by liquid, mixture, and vapor may be estimated using the amount of heat transferred. Pressure drop element is also needed for the condenser simulation. Similar to condenser, evaporator can be modeled using the same modeling method. Pumps used in the testing system can be modeled using their performance characteristics given in the respective manuals and experimental data. Pipes and pipe control units can be modeled using 134

147 standard method recommended in handbooks for pressure drops. Heat transfer effect of the pipe may be less critical, if all the pipes are short and insulated properly as for the expected experimental setup. If the third stage is conducted, the model parameters will be fitted and extra simulations will be performed and results will be used to verify the model. 135

148 Appendix-IIIB Expressions for Single Phase and Two-Phase Heat Transfer Coefficient of Fluids in Plate Heat Exchangers In calculating the heat transfer coefficients of fluids we need to know the physical parameters of the plate heat exchanger. Table-IIA gives the physical parameters of a typical plate heat exchanger taken from open literature [1]for a 50 kw ORC system. Table IIIB-1:Plate Heat Exchanger Physical Parameters considered in the present simulation Plate Width (w in m) 0.5 Plate Length (L p in m) 1.1 Channel Spacing (m) 3.50E-03 Thickness of Plate (t in m) 6.00E-04 Cheveron angle (β in degrees) 45 Enlargement Factor (Φ) 1.29 Corrugation Pitch 7.00E-03 Equivalent diameter (D e in m) 7.00E-03 Projected area per plate (A p in m 2 ) 0.55 Flow area on one fluid side (A f in m 2 ) 4.20E-02 Surface area on one fluid side (m 2 ) 26.4 Hydraulic Diameter (D h in m) Pressure drop in heat exchanger (MPa) 0.05 Plate Thermal Conductivity (K p in W/m-K) 13.8 Plate Thermal Resistance (m 2 -K/W) 4.35E-05 Area of Evaporator (m 2 ) Area of Condenser (m 2 ) 28.3 Area of Pre Heater (m 2 ) 15 Crltical Pressure of R-245fa (MPa) Crltical Temperature of R-245fa ( o C) Heat Transfer Coefficient for Single Phase Fluids The expression for convective heat transfer coefficient for single phase fluids in a plate heat exchanger is given by Muley and Manglik [1]and is read as h f = (Nu K f )/D e (1) 5 [ sin[( πβ / 45) + 3.7]] 1/ = [ β β ] Re Pr ( µ / w ) (2) Nu µ In the above equation K f is thermal conductivity of fluid and the Reynolds number (Re) is based on equivalent diameter of the plate heat exchanger and is calculated as, 136

149 Re= (G D e )/µ (3) Where G is the mass velocity of the fluid, D e is the equivalent diameter and µ is the dynamic viscosity of the fluid. The mass velocity (G) is the ratio of mass flow rate of the fluid to flow area of the heat exchanger. Heat Transfer Coefficient for Evaporating Fluids The expression for convective heat transfer coefficient for evaporating fluids in a plate heat exchanger is given by Ayub [2] and is read as h Eva = ( K f / De )[Re h fg / L p ] ( p / pcr ) (65/ β ) [BTU/hr-ft 2 - o F] (4) In the above equation K f is the conductivity of the fluid, h fg is the enthalpy difference between outlet and inlet of heat exchanger, L p is length of the plate, p is pressure and p cr is critical pressure. Reynolds number (Re) is based on equivalent diameter and its calculation is similar to the above equation for single phase fluids (Eq. (3)). Heat Transfer Coefficient for Condensing Fluids The expression for convective heat transfer coefficient for condensing fluids in a plate heat exchanger is given by Selvam et al [3] and is read as h Cond = (Nu Cond K f )/D h (5) 2 1/ 3 Nu 1 Re Ge Cond = Ge Eq Pr (6) pco π Ge = β (7) Dh pco π Ge 2 = 0.35 β (8) Dh 2 GEq Dh Re Eq = (9) µ f G Eq 0.5 = G [1 x + x( ρ / ρ ) ] (10) c f g 137

150 G c m = (11) A f Where K f is the thermal conductivity of the fluid, p co is plate corrugation pitch, D h is hydraulic diameter, x is quality of fluid entering the condenser, ρ f and ρ g are the density of saturated liquid and vapor at condenser inlet condition. Here we need to observe that the Reynolds number (Re) is based on hydraulic diameter. REFERENCE: [1] Muley, A., Manglik, R. M., Experimental Study of Turbulent Flow Heat Transfer and Pressure Drop in a Plate Heat Exchanger with Chevron Angle, ASME Journal of Heat Transfer, February 1999, Vol.121, pp [2] Ayub, Z. H., Plate Heat Exchanger Literature Survey and New Heat Transfer and Pressure Drop Correlations for Refrigerant Evaporators, Journal of Heat Transfer Engineering, 2003, Vol.25 (5), pp [3] Selvam, M. A. J., Senthil, K. P., Muthuraman, S., The Characteristics of Brazed Plate Heat Exchangers with Different Chevron Angles, ARPN Journal of Engineering and Applied Sciences, December 2009, Vol. 4 (10), pp

151 Appendix IVA Available Floor Space for Test System Installation 139

152 Appendix IVB Electrical Circuit Diagram 140

153 Appendix V Pictures of the Integrated System after Installation Steam flow control valve connected to steam-to-hot water heat exchanger 141

154 Steam trap connected to steam-to-hot water heat exchanger 142

155 Hot water piping (with Gruvlok fittings) connected to steam-to-hot water heat exchanger 143

156 Hot water supply and return pipes to/from GM 144

157 Hot water pump (VFD pump) in hot water piping loop 145

158 Expansion tanks in hot water piping loop 146

159 Air separator and pressure relief valve in hot water piping loop 147

160 More hot water supply and return piping to/from GM 148

161 BTU meter in hot water piping loop 149

162 Hot water supply and return piping connections on GM side 150

163 Cold water (heat sink fluid) supply and return piping to/from GM 151

164 BTU meter on cold water supply piping to GM 152

165 Cold water bypass valves for temperature and flow control 153

166 Cold water pump 154

167 Check valve on cold water supply piping to GM to prevent back flow of water into fire hydrant 155

168 National instruments DAQ system being installed for data collection and performance analysis 156

169 Appendix VI Estimated additional kw that could be generated using an ORC unit for all Alaskan communities Table VI1. Estimated additional kw that could be generated for each community in Alaska, if an appropriately sized ORC were installed to use rejected heat. Cases were considered for jacket water (JW) only, and jacket water plus exhaust heat recovery. Note that not all of these communities have adequate heat available to operate a 50 kw Green Machine, and many communities would only be able to divert excess heat to an ORC seasonally, as heat recovery for space heating is common in Alaska. Note the table below is organized by region. Community Annual Diesel Electric Power (MWh) Average Diesel Power (kw) (40% of Fuel Energy) JW Heat (kw) (20% of fuel Energy) Exhaust Heat (kw) (30% of Fuel Energy) JW Heat recovery (80%) (kw) Exhaust Heat Recovery (50%) (kw) Heat Recovery of JW + Exhaust Heat (kw) 6.9% Efficiency (JW) 8% Efficiency (JW + Exhaust) Ambler 1, Brevi g Mission Elim 1, Gambell 1, Kiana 1, Kivalina 1, Koyuk 1, Noatak 1, Noorvik 1, Savoonga 1, Selawik 2, Shaktoolik Shishmaref 1, Shungnak 1, St. Micheal 1, Stebbins 1, Wales Deadhorse 25, Buckland Diomede Golovin Deering Kobuk Valley Kotzebue 20, Snake River 27, Anaktuvuk Pas s 3, Atqasuk 3, Kaktovik 1, Nuiqsut 4,

170 Community Annual Diesel Electric Power (MWh) Average Diesel Power (kw) (40% of Fuel Energy) JW Heat (kw) (20% of fuel Energy) Exhaust Heat (kw) (30% of Fuel Energy) JW Heat recovery (80%) (kw) Exhaust Heat Recovery (50%) (kw) Heat Recovery of JW + Exhaust Heat (kw) 6.9% Efficiency (JW) 8% Efficiency (JW + Exhaust) Point Hope 2, Point Lay 3, Wainwright 2, Teller Unalakleet 4, White Mountain Akhiok Old Harbor Karluk Anchorage Chenega Bay Chitina Seward Glenallen 16, Valdez Orca 20, Seldovia Kodiak 41, Larsen Bay Ouzinkie Tatitlek Auke Bay Lemon Creek Coffman Cove Craig, Klawock Haines Hollis Hydaburg 1, Nautaki

171 Community Annual Diesel Electric Power (MWh) Average Diesel Power (kw) (40% of Fuel Energy) JW Heat (kw) (20% of fuel Energy) Exhaust Heat (kw) (30% of Fuel Energy) JW Heat recovery (80%) (kw) Exhaust Heat Recovery (50%) (kw) Heat Recovery of JW + Exhaust Heat (kw) 6.9% Efficiency (JW) 8% Efficiency (JW + Exhaust) Skagway Whale Pass Elfin Cove Gustavus 1, S.W. Bailey 7, Centinneal Pelican Peters burg 1, Indian River Tenakee Springs Thorne Bay Chilkat Valley Hoonah 6, Ka ke 4, Kasaan Angoon 1, Wrangell Yakutat 8, Akiachak 1, Akiak Akutan Eek Goodnews Bay Lower Kalskag 1, Mekoryuk Nightmute Nunapitchuk 2, NW Stuyahok 1, Quinhagak 1,

172 Community Annual Diesel Electric Power (MWh) Average Diesel Power (kw) (40% of Fuel Energy) JW Heat (kw) (20% of fuel Energy) Exhaust Heat (kw) (30% of Fuel Energy) JW Heat recovery (80%) (kw) Exhaust Heat Recovery (50%) (kw) Heat Recovery of JW + Exhaust Heat (kw) 6.9% Efficiency (JW) 8% Efficiency (JW + Exhaust) Togiak 2, Toksook Bay 1, Tununak Atka Aniak 2, Atmautluak Bethel 41, Chignik Chignik Lagoon Chignik Lake Egegik False Pass Cold Bay 3, Igiugig I-N-N ELEC (Iliamna) 2, King Cove 3, Kipnuk 1, Kokhanok Koliganek Kwethluk 1, Kwigilligok Levelock Lime Village Manokotak McGra th 2, Chauthbaluk Crooked Creek Red Devil Sleetmute

173 Community Annual Diesel Electric Power (MWh) Average Diesel Power (kw) (40% of Fuel Energy) JW Heat (kw) (20% of fuel Energy) Exhaust Heat (kw) (30% of Fuel Energy) JW Heat recovery (80%) (kw) Exhaust Heat Recovery (50%) (kw) Heat Recovery of JW + Exhaust Heat (kw) 6.9% Efficiency (JW) 8% Efficiency (JW + Exhaust) Stony River Naknek 20, Napakiak Napaskiak Chefornak Nelson Lagoon Nikolai Dillingham 17, Pedro Bay Perryville Pilot Point Platinum Port Heiden Kongiganak Sand Point 3, St. George 1, St. Paul 3, Takotna Port Alsworth Tuluksak Tuntutuliak Twin Hills Village Council Nikolski Unalaska 29, Newtok Allakaket Bettles Chistochina Eagle

174 Community Annual Diesel Electric Power (MWh) Average Diesel Power (kw) (40% of Fuel Energy) JW Heat (kw) (20% of fuel Energy) Exhaust Heat (kw) (30% of Fuel Energy) JW Heat recovery (80%) (kw) Exhaust Heat Recovery (50%) (kw) Heat Recovery of JW + Exhaust Heat (kw) 6.9% Efficiency (JW) 8% Efficiency (JW + Exhaust) Healy Lake Mentasta Lake Northway 1, Tetlin Tok 11, Alakanuk 1, Anvi k Cheva k 1, Emmonak 2, Grayling Holy Cross Hooper Bay 2, Huslia Kaltag Marshall 1, Mi nto Mt.Village 2, Nulato 1, Pilot Station 1, Russian MIS Scammon Bay 1, Shageluk St. Mary's (Pitkas Point) 2, Beaver Birch Creek Centra l Chalkyitsik Circle Galena 9, Community Annual Diesel Electric Power (MWh) Average Diesel Power (kw) (40% of Fuel Energy) JW Heat (kw) (20% of fuel Energy) Exhaust Heat (kw) (30% of Fuel Energy) JW Heat recovery (80%) (kw) Exhaust Heat Recovery (50%) (kw) Heat Recovery of JW + Exhaust Heat (kw) 6.9% Efficiency (JW) 8% Efficiency (JW + Exhaust) E Ft. Yukon 2, Hughes Kotlik 1, Koyukuk Manley Sheldon Point Denali 2, Ruby Tanana 1, Venetie l (k ) 162

175