A ZERO-ODP, LOW GWP WORKING FLUID FOR HIGH TEMPERATURE HEATING AND POWER GENERATION FROM LOW TEMPERATURE HEAT: DR-2

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1 JRAIA2012KOBE xxxx A ZERO-ODP, LOW GWP WORKING FLUID FOR HIGH TEMPERATURE HEATING AND POWER GENERATION FROM LOW TEMPERATURE HEAT: Konstantinos (Kostas) KONTOMARIS DuPont Fluoroproducts--Wilmington, Delaware, USA Konstantinos.Kontomaris@DuPont.com ABSTRACT This paper introduces a new developmental refrigerant,, as a potential working fluid for high temperature heat pumps and Rankine cycles. is a hydrofluoro-olefin-based fluid with an ozone depletion potential of zero and a global warming potential of less than 10. It is non-flammable and has a favorable toxicity profile based on testing to date. remained chemically stable in the presence of metals up to the maximum temperature tested of 250 o C. thermodynamic performance under cycle conditions representative of potential applications was evaluated through computational modeling. has a relatively high critical temperature, generates relatively low vapor pressures and enables high cycle energy efficiencies. It could enable more environmentally sustainable heat pump and Rankine cycle platforms for the utilization of abundantly available low temperature heat to generate mechanical or electrical power or meet heating duties at higher temperatures and with higher energy efficiencies than incumbent working fluids. Keywords: heat recovery, low GWP, global warming, Hydro-Fluoro-Olefin (HFO), refrigerants, working fluids, climate change, energy efficiency, sustainability INTRODUCTION Current trends shaping the global energy landscape suggest an expanding utilization of low temperature heat (i.e. heat at temperatures lower than about 300 o C) in the near future. Such heat may be recovered from various commercial or industrial operations, can be extracted from geothermal or hydrothermal reservoirs or can be generated through solar collectors. Elevation of the temperature of available heat through high temperature mechanical compression heat pumps (HTHPs) to meet heating requirements and conversion of the available heat to mechanical or electrical power through Organic Rankine Cycles (ORCs) are two promising approaches for the use of low temperature heat. (Cooling, heating or power generation through absorption cycles is beyond the scope of this paper.) Heat at temperatures below about 75 o C would probably be best used for heating (i.e. as input to a HTHP), since conversion to power with commonly available sinks for heat rejection would be uneconomical in most cases due to poor energy efficiency (except in special circumstances where the power generated is highly valued). Heat at temperatures higher than about 125 o C would probably be best converted to power in most cases given the high versatility in power use. Motivation for low temperature heat utilization is provided by: i) current and projected energy supply shortages (e.g. electricity in Japan and China; peak electricity and petroleum globally); ii) increasing energy prices resulting from supply/demand imbalances in conventional forms of energy or the adoption of generally more expensive alternative (e.g. renewable) forms of energy; iii) growing awareness of the environmental impacts, in general, and the threat to the earth s climate, in particular, from the use of fossil fuels and resulting intensifying efforts to reduce fossil energy use and associated greenhouse gas emissions; iv) increasing recognition of the economic risks from dependence on and the geopolitical risks in securing access to unreliable energy supply sources; v) heightened awareness by militaries around the world of the risks to national security and troop safety posed by the logistics of supplying large amounts of energy to remote and/or hostile theaters of operation; vi) proliferation of emerging technologies that could expand the easily accessible low temperature heat resource (e.g. local, stand-alone, remote, supplemental or backup distributed power generation and cogeneration of heat and power for commercial, industrial and military end use from a variety of primary energy sources including biomass, municipal and other waste, natural gas and bio-gas, bio-diesel, and solar or geothermal heat). It seems likely that the economic attractiveness of low temperature heat utilization will be enhanced by optimum integration within broader energy systems that will allow highest value use of the available heat. System optimization at the scale of individual or multiple buildings is already underway. System optimization at

2 the scale of whole cities is emerging on the horizon and the necessary infrastructure (monitoring, data handling, computing, electric power grid systems, etc.) is already under development by credible organizations (e.g. [1]). Heat pumps operating according to a reverse Rankine cycle and ORCs require the use of working fluids. Working fluids currently in common use for HTHPs and ORCs either are controlled as ozone depleting substances under the Montreal Protocol (e.g. CFC-114) or are coming under increasing scrutiny because of their high GWPs (e.g. ). Clearly, there is an increasing need for more environmentally sustainable working fluids for HTHPs and ORCs, especially given that environmental sustainability is a primary motivation for low temperature heat utilization. A new generation of refrigerants with low GWPs, known as Hydro-Fluoro-Olefins (HFOs), may enable the design and operation of energy efficient HTHPs and ORCs with reduced environmental impact. is a relatively low pressure working fluid, based on HFO technology, that was recently evaluated as a replacement for HCFC-123 in centrifugal chillers [2, 3] and as a working fluid for HTHPs [4, 5]. Preliminary assessments of as an ORC fluid have also been carried out [6, 7]. Table 1 summarizes key thermodynamic, safety, health and environmental characteristics of. It compares to a reference fluid with similar thermodynamic properties, HCFC-123, and two reference fluids that have been widely used for HTHPs and ORCs, CFC-114 and. HCFC-123 has not been widely used for HTHPs or ORCs, despite its attractive thermodynamic properties, apparently because of its limited thermal stability. miscibility with polyol ester (POE) lubricants at temperatures up to 85 o C, chemical stability in the presence of metals with and without lubricant at temperatures up to 175 o C, and compatibility with common plastics and elastomers at room temperature have been reported [2]. The main objective of this paper was to evaluate the potential of as a working fluid for HTHPs and ORCs. A related objective was to assess the chemical stability of in the presence of metals at temperatures up to 250 o C. METHODS The thermo-physical properties of were estimated using group contribution methods and equations of state from measurements of the liquid density and vapor pressure over a wide temperature range and of the critical properties. The thermodynamic performance of in illustrative idealized cycles representative of potential applications was computed and compared to the performance of, a familiar working fluid used in HTHPs and ORCs. The thermal stability of in the presence of aluminum, steel and copper was assessed according to the familiar sealed glass tube methodology of ASHRAE-ANSI Standard 97. Table 1: Key properties of compared to incumbent working fluids. HFC- HCFC CFC (1) 245fa (1) 114 (1) T b, o C T cr, o C P cr, MPa LFL, vol% None (2) None None None OEL, ppmv 500 (3) ,000 Safety A1 (5) B1 B1 A1 Class (4) ODP 0 (6) ALT, yrs (7) GWP 9.4 (7) 77 1,030 10,040 (1) From [8] (2) At 60 o C and 100 o C according to ASTM E (3) DuPont Allowable Exposure Limit (AEL) (4) ASHRAE Standard 34 (5) Not established, but meets criteria of A1 (6) No halogen atoms in other than fluorine (7) National Oceanic and Atmospheric Admin., 2010 RESULTS (a) Thermodynamic Properties Figure 1 compares the vapor pressure of to. generates vapor pressures at high temperatures substantially lower than. As a result, use of could allow higher condensing temperatures for HTHPs and higher evaporating temperatures for ORCs than without exceeding the maximum design working pressure of commonly available low cost equipment. For example, the maximum permissible working pressure for some commonly available large centrifugal heat pumps may be limited to values below 2.18 MPa. and would allow condensing temperatures of up to about 155 o C and o C, respectively, without exceeding a vapor pressure of 2.18 MPa. Furthermore, the higher critical temperature of (T cr =171.3 o C) would also be advantageous for HTHPs and ORCs because it would allow conventional subcritical operation at higher condensing and evaporating temperatures, respectively, than (T cr =154 o C). Pressure (MPa) Fig Temperature (C) vapor pressure compared to.

3 Figure 2 compares the computed temperature-entropy diagram of to. Similarly to, exhibits a saturated vapor curve with a small positive [dt/ds] slope. As a result, superheat at the inlet of the expander of an ORC cycle would not be required to ensure dry expansion. pressure with at o C is approaching the design working pressure limit of commonly available large centrifugal heat pumps. Table 2: Predicted thermodynamic performance of HTHP cycles: T cond = 100 o C. Temperature [ o C] Fig. 2. (b) Entropy (kj/kg.k) temperature-entropy diagram compared to. Thermodynamic Cycle Performance: HTHPs Heating at temperatures o C is often needed for various commercial and industrial applications (e.g. district heating, hydronic space heating, equipment cleaning, boiler water preheating, drying and other process heating). Heating using a HTHP may be attractive relative to fossil fuel heating, when heat is available at temperatures in the range of o C (e.g. condenser water, process water, solar or low grade geothermal heat). Table 2 compares the thermodynamic performance of and in a hypothetical heat pump application meeting a heating duty requiring a condensing temperature of 100 o C using available heat supplied to the evaporator operating at 60 o C. could enable such an application with a very attractive coefficient of performance for heating, COP h, higher than 7. The COP h with would be 2.3% higher than. The volumetric heating capacity, CAP h, with would be about 42.7% lower than, as expected given the lower vapor pressure of. Table 3 exemplifies the use of a HTHP to meet a heating duty requiring a condensing temperature of o C (e.g. steam generation). Available heat supplied to the evaporator allows an evaporating temperature of 75 o C. The resulting COP h of over 5 suggests that heat pump heating for this application could be quite attractive relative to heating with a fossil fuel heater (e.g. steam boiler). The efficiency/capacity trade-off expected between fluids of different vapor pressures is observed: has a 4.7% higher COP h and a 39.8% lower CAP h relative to. The condensing T cond T evap T superh T subc o C o C o C o C EFFC compr P evap MPa P cond MPa PR T compr_disch o C COP h Table 3: CAP h kj/m 3 3,927 2,248 T cond T evap T superh T subc Predicted thermodynamic performance of HTHP cycles: T cond = o C. o C o C o C o C EFFC compr P evap MPa P cond MPa PR T compr_disch o C COP h CAP h kj/m 3 4,833 2,909 Table 4 assesses the possibility of using a heat pump operating with to deliver a condensing temperature of 155 o C to meet some industrial process heating duty. Attractive COP h values are enabled when input heat allows evaporating temperatures higher than 80 o C. The condensing pressure at 155 o C remains within the capabilities of most commonly available large centrifugal heat pumps. However, some equipment modifications would likely be required to mitigate the effects of the relatively high compressor discharge temperatures. Moreover, multistage compression would probably be required at high temperature lifts. could not be used for this application with

4 conventional subcritical heat rejection through condensation because the required condensing temperature would exceed the critical temperature of. (c) Table 4: Predicted thermodynamic performance of HTHP cycles: T cond = 155 o C. T cond T evap T superh T subc o C o C o C o C EFFC compr P evap MPa P cond MPa PR T compr_disch o C COP h CAP h kj/m 3 2,438 4,281 7,221 Thermodynamic Cycle Performance: ORCs The performance of in ORCs at prescribed conditions was evaluated and compared to several fluids by Datla and Brasz (2012) [7]. The heat source temperature varied from 100 to 80 o C. The performance of was found to be similar to. Table 5 summarizes the performance of in ORCs driven by heat sources allowing higher evaporator temperatures than those considered by Datla and Brasz (2012) [7]. It also compares the performance of ORCs operating with and. The prescribed cycles generate power from available heat allowing evaporator operation at 135 o C or 160 o C. They reject heat under two scenarios: at high condensing temperatures (T cond =55 o C; e.g. air-cooled condensers, hot climates) or low condensing temperatures (T cond =35 o C). The estimated ideal cycle thermal efficiency at T evap =135 o C with, 10.59% (column 2), is 2.8% higher than with (column 1). would generate a substantially lower evaporator pressure than, which could be advantageous in reducing equipment costs and fluid leakage. However, is estimated to have a 40.8% lower volumetric capacity for power generation, CAP e, which would require a larger turbine impeller (or larger expander, in general) than to deliver the same work rate. Columns 3 and 4 in Table 5 summarize the predicted performance of in ORCs operated at evaporating temperatures exceeding the critical temperature of and condensing temperatures corresponding to air-cooled and water-cooled condensers, respectively. As expected, cycle energy efficiency increases as the difference between evaporating and condensing temperature increases. Even at the lowest condensing temperature considered in Table 5, (35 o C, column 4), the condenser pressure with remains slightly above atmospheric level, thus minimizing the risk of air infiltration. (d) Table 5: Predicted thermodynamic performance of ORCs: T evap = 135 o C; 160 o C HFC- 245fa T evap, C T cond, C T suph, C T subc, C EFFC expn EFFC pump P evap, MPa P cond, MPa EFF cycle, % CAP e, kj/m Thermal Stability Sealed glass tubes each containing a carbon steel, a copper and an aluminum coupon (with the copper coupon placed between the other two coupons) immersed in were prepared. The stock of used in the sealed tube tests was wt% pure (136 ppmw of impurities) and was dried and de-aerated before aging. The tubes were aged in a heated oven for two weeks at temperatures up to 250 o C. Visual inspection of the tubes after thermal aging indicated clear liquids with no discoloration, residues or other visible deterioration of the refrigerant. Moreover, there was no change in the appearance of the metal coupons indicating corrosion, insoluble residues or other degradation. Chemical analyses of the refrigerant liquids after aging are summarized in Table 6. They indicated negligible concentrations of fluoride even at the highest temperature tested. The fluoride ion concentration can be interpreted as an indicator of the degree of

5 degradation. Table 6 suggests that degradation was minimal even at the highest temperature tested (250ºC). Table 6: Measured fluoride ion concentration in samples after aging at various temperatures for two weeks. Aging Temperature [ºC] F-ion [ppmm] 175 <0.15 (*) (*) No detectable fluoride (within the method detection limit of 0.15 ppm) SUMMARY-DISCUSSION is a relatively low pressure developmental refrigerant based on HFO technology. In addition to its attractive safety, health and environmental characteristics, provides an uncommon combination of properties particularly advantageous, relative to other organic fluids currently in common use, for high temperature operation of heat pumps and Rankine cycles: high critical temperature, low vapor pressure and high thermal stability. It could enable HTHPs and ORCs with high energy efficiencies and contribute to meeting sustainability objectives (e.g. reducing demand for non-renewable primary energy sources and reducing greenhouse gas emissions) with attractive economics. It seems promising that could enable large centrifugal HTHPs, largely consisting of commonly available chiller components, delivering condensing temperatures approaching or exceeding 155 o C. The critical temperature of is sufficiently higher than 155 o C that condenser design would probably pose no insurmountable challenges. Moreover, the vapor pressure of at 155 o C is sufficiently low that it could be accommodated without major mechanical component modifications. However, the high compressor discharge temperatures could pose challenges in the compressor heat management and the selection of lubricants and polymeric materials of equipment construction. A HTHP could, in principle, be operated to simultaneously provide heating (e.g. hot water for domestic service) and cooling (e.g. chilled water for air conditioning). Such a mode of operation would impose a high required temperature lift and, therefore, it would require a large amount of compression work to lift a unit of mass of the working fluid from the thermodynamic state of the evaporator to that of the condenser. The resulting COP for cooling and COP for heating would, generally, be lower than the values expected from machines specifically operated to provide solely either cooling or heating. However, the combined COP for both cooling and heating could still be attractive. For example, a heat pump with operated at an evaporator temperature of 5 o C and a condenser temperature of 65 o C (with 10 o C of vapor superheat at the evaporator exit, 5 o C of liquid sub-cooling at the condenser exit, and a compressor efficiency of 0.80) would achieve an ideal cycle COP for cooling of and a COP for heating of 3.854, for a total COP for both cooling and heating of For some heat sources, maximum net power generation (achieved by maximization of the product of the fraction of the available heat extracted and the fraction of the extracted heat converted to power) could require heat extraction at pressures above the critical pressure of the working fluid. The thermal stability of at supercritical conditions would allow trans-critical operation of ORC cycles when required to maximize net power generation from a given heat source. Work is continuing to assess the chemical stability of at temperatures higher than 250 o C and in the presence of controlled concentrations of air and moisture that could occasionally infiltrate ORC equipment. The design of an accelerated thermal stability test that accounts for the intermittent nature of fluid exposure to high temperatures in an ORC would be of value for the rapid screening of candidate fluids. Finally, evaluation of candidate working fluids for heat recovery applications, in general, and ORCs, in particular, often requires estimation of fluid thermodynamic properties in the vicinity of the critical point. It could be benefited by practical methodologies for the accurate estimation of near-critical thermodynamic properties of HFO-based fluids. REFERENCES [1] Saito, Yutaka: Building Livable Cities, Hitachi Technology Report, 2012: icsfiles/af ieldfile/2012/08/08/r2012_technology02.pdf [2] Kontomaris, K.: "A Low GWP Replacement for HCFC-123 in Centrifugal Chillers: ", ASHRAE & UNEP Conference Road to Climate Friendly Chillers: Moving Beyond CFCs and HCFCs, Cairo, Egypt, Sep. 30-Oct 1, 2010 [3] Kontomaris, K.: "Global Warming Impact of Low GWP Refrigerants" 23rd IIR International Congress of Refrigeration, Prague, Czech Republic, Aug , 2011a [4] Kontomaris, K.: A Low GWP Working Fluid for High Temperature Heat Pumps:, Chillventa, European Heat Pump Summit, Nuremberg, Germany, Sep , 2011b [5] Kontomaris, K. and T. J. Leck: Low GWP Working

6 Fluids for Domestic, Commercial and Industrial Heat Pumps: DR-5 and, The IEA Heat Pump Center Newsletter, Vol. 29, No 4, pg , 2011 [6] Kontomaris, K.: Low GWP Working Fluids for Low Temperature Heat Utilization, Low Carbon Earth Summit, Dalian, China, Oct , 2011c [7] Datla, Bala V. and Joost J. Brasz: Organic Rankine Cycle System Analysis for Low GWP Working Fluids, International Refrigeration and Air Conditioning Conference at Purdue, July 16-19, 2012 [8] Calm, J.M. and G.C. Hourahan: Refrigerant data update, Heating/Piping/Air Conditioning Engineering, Vol. 79(1), 50 64, 2007 NOMENCLATURE ALT: Atmospheric Life Time CAP e : Volumetric capacity for power generation; ORC expander power divided by the fluid volumetric flow rate at the expander outlet CAP h : Volumetric heating capacity; the amount of heat delivered at the condenser (including the compressed vapor superheat and liquid sub-cooling) per unit volume of the working fluid entering the compressor COP h : Coefficient of Performance for Heating; ratio of the heat delivered at the condenser (including the compressed vapor superheat and the liquid sub-cooling) and the work of compression EFFC compr : Compressor efficiency EFFC expn : Expander efficiency EFFC pump : Liquid pump efficiency EFF cycle : ORC cycle thermal efficiency (% of heat input converted to mechanical energy) GWP: Global Warming Potential (one hundred year integrated time horizon) HFO: Hydro-Fluoro-Olefin HTHP: High Temperature Heat Pump (mechanical compression) LFL: Lower Flammability Limit OEL: Occupational Exposure Limit ORC: Organic Rankine Cycle P cond : Condenser pressure P cr : Critical pressure P evap : Evaporator pressure PR: Pressure Ratio (condenser pressure over evaporator pressure) T b : Boiling point at 1 atm T compr_disch : Compressor discharge temperature T cond : Condenser temperature T cr : Critical temperature T evap : Evaporator temperature T subc : Liquid sub-cooling at the condenser exit T superh : Vapor superheat at the evaporator exit