Performance Comparison of Hydronic Secondary Loop Heat Pump and Conventional Air-Source Heat Pump

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1 Purdue Universit Purdue e-pubs International Refrigeration and Air Conditioning Conference School of Mechanical Engineering 2012 Performance Comparison of Hdronic Secondar Loop Heat Pump and Conventional Air-Source Heat Pump Ian H. Bell James E. Braun llo this and additional orks at: Bell, Ian H. and Braun, James E., "Performance Comparison of Hdronic Secondar Loop Heat Pump and Conventional Air-Source Heat Pump" (2012). International Refrigeration and Air Conditioning Conference. Paper This document has been made available through Purdue e-pubs, a service of the Purdue Universit Libraries. Please contact epubs@purdue.edu for additional information. Complete proceedings ma be acquired in print and on CD-RM directl from the Ra W. Herrick Laboratories at Herrick/Events/orderlit.html

2 2597, Page 1 Ian H. BELL1*, James E. BRAUN2 Performance Comparison of Hdronic Secondar Loop Heat Pump and Conventional Air-Source Heat Pump 1 Bell Thermal Consultants ian.h.bell@gmail.com 2 Purdue Universit, Mechanical Engineering Department, Herrick Labs West Lafaette, IN, USA jbraun@purdue.edu * Corresponding Author ABSTRACT In residential heat pump sstems, the motivation for secondar loop sstems is to allo for the use of flammable or toxic refrigerants ith loer global arming potentials than the currentl emploed HFC refrigerants. The addition of radiant panels as integral building components (embedded in concrete at construction or attached to the underside of ood flooring) is becoming more common. Combining the large surface area of the radiant panel and an efficient primar loop, a hdronic secondar loop heat pump sstem can greatl outperform a conventional air-to-air heat pump. The improvement in coefficient of performance is as much as 38% over a conventional air-to-air heat pump hen the secondar loop hdronic sstem is emploed. Due to the large area of the radiant panel, the condensing temperature of the primar loop for the hdronic secondar loop heat pump can be reduced b as much as 5 C at high ambient temperatures. 1. INTRDUCTIN Radiant heating and cooling sstems are becoming increasingl popular around the orld. As of 2002, radiant heating and cooling ere emploed in about 90% of ne constructions in Korea, and 30-50% in German (lesen, 2002). Man of these sstems emplo electric ater heaters as the heat source, hich hile eas to implement, is not the most efficient solution. The conventional air-to-air heat pump is quite ubiquitous. Approximatell 1.8 million split-tpe air-source heat pumps ere produced b U.S. manufacturers in 2011, and 20 million units beteen 1992 and 2011 (AHRI, 2012). Man of the modern heat pump units operate ith the refrigerant R410A hich has a global arming potential (GWP) of 2100, versus a GWP for propane of 20 (Calm, 2007). Regulator pressure in Europe and elsehere is pushing toards sstems that do not use HFC refrigerants like R410A, and regulator bodies are reconsidering the use of flammable refrigerants like propane. Radiant heating sstems are emploed not just for their performance benefits. The can also improve occupant comfort due to a more even temperature profile in the occupied space (lesen, 2002). In addition, it has been suggested that the air temperature setpoint can be reduced due to the use of radiant heat, although that effect has not been investigated here. In principle, the hdronic heating sstem could also be coupled ith solar collectors and use the solar energ as the heat source for the heat pump. This ould enable solar collector efficiencies that are substantiall better than direct space heating sstems because of relativel lo collector temperatures and heat pump CPs that are significantl better than air-source heat pumps. Alternativel, the hdronic heating sstem could be coupled ith geothermal heat pumps, hich also have significant performance advantages as compared ith air-source heat pumps. International Refrigeration and Air Conditioning Conference at Purdue, Jul 16-19, 2012

2597, Page 2 2. SYSTEM DESCRIPTINS 2.1 Sstems Under Consideration In this stud, to different tpes of sstems are under consideration for the heating of a residential structure.

3 2597, Page 2 2. SYSTEM DESCRIPTINS 2.1 Sstems Under Consideration In this stud, to different tpes of sstems are under consideration for the heating of a residential structure. These sstems are a conventional air-to-air heat pump that operates ith an HFC-based orking fluid (here refrigerant R410A) and a secondar loop heat pump sstem that emplos radiant panels for heating of the occupied space and propane as the refrigerant. The secondar loop sstem allos the use of flammable orking fluids like propane that ould not be acceptable for direct expansion sstems. Figure 1a shos a schematic of the air-to-air direct expansion heat pump sstem. In the direct expansion sstem, refrigerant exits the evaporator (the outdoor heat exchanger) at state point 1. The compressor then compresses the refrigerant from state point 1 to state point 2. The refrigerant is heated against the ambient air up to state point 3 as it passes through the vapor line on its a to the condenser. In the condenser, heat is delivered to the heated space b the condensation of the refrigerant, and the refrigerant condenses from state point 3 to state point 4. An auxiliar heater can also be provided to add additional heating capacit beond that hich the heat pump can provide. In general, the refrigerant is subcooled at the outlet of the condenser. At the outlet of the condenser, the refrigerant passes through the liquid line from state point 4 to state point 5. The refrigerant next passes through the expansion device hich throttles the refrigerant to state point 6, the iet to the evaporator. Then the refrigerant re-enters the evaporator and the ccle continues. (a) DX R410A Sstem (b) Hdronic Loop Sstem Figure 1: Schematics of DX and Secondar Loop sstems In the secondar loop heat pump sstem considered here, there are a fe modifications to the conventional air-to-air heat pump. Figure 1b shos a schematic of the secondar loop sstem. The primar refrigerant loop (state points 1 to 4) is essentiall the same, except that the refrigerant delivers its heat into a secondar loop in the plate heat exchanger as it condenses, and there are no vapor or liquid lines since the hole compressor-expansion device-plate heat exchanger loop can be close-coupled, decreasing greatl the piping pressure losses and the refrigerant charge. In the secondar loop, a secondar orking fluid is heated from state point 5 to state point 6 in the plate heat exchanger against the condensing refrigerant. The armed secondar orking fluid then passes through an auxiliar heater to provide additional heating capacit if needed. The secondar orking fluid passes through the suppl pipe and is then delivered to the radiant panels. In the radiant panels, the secondar orking fluid is cooled and delivers its heat to the heated space. The cooled secondar orking fluid is then returned to the plate heat exchanger through the return pipe. The pump is used to force the secondar orking fluid through the loop. r the secondar loop sstem, there are a number of options for the secondar orking fluid, but ater is an excellent choice. It has the highest mass specific heat of an liquid, and one of the highest densities of an liquid. International Refrigeration and Air Conditioning Conference at Purdue, Jul 16-19, 2012

2597, Page 3 In addition, its lo viscosit results in lo pressure drops. r radiant sstems applied to heating applications, ater is the tpical secondar orking fluid.

4 2597, Page 3 In addition, its lo viscosit results in lo pressure drops. r radiant sstems applied to heating applications, ater is the tpical secondar orking fluid. r use ith a heat pump, it might be necessar to have controls and hardare for automatic draining and re-priming of the secondar loop for freeze protection in the event of other hardare failures. 2.2 Air-Conditioning Heat Pump Model (ACHP) A specialized model has been developed to analze direct-expansion and secondar loop heat pump and airconditioning sstems (ACHP) that is freel available oine, including source code. The details of the ACHP model are provided in the documentation available oine 1. In ACHP, each of the heat exchanger models are based on moving boundar formulations. Essentiall, the moving boundar heat exchanger model is based on using a numerical solver to find the locations here the refrigerant changes phase, and then solve each of the portions of the heat exchanger separatel. The addition of partiallet/partiall-dr air-side surfaces analsis is also included in the evaporator. The compressor model is based on a 10-coefficient compressor performance map ith appropriate correction for compressor iet superheat. Models are also available for the plate heat exchanger, line sets and other components. The refrigerant and humid air properties are based on a reference-qualit propert database developed in parallel ith ACHP2. Coupling all the component models together, a multi-dimensional numerical solver is used to find the evaporation and condensation saturation temperatures for the refrigerant loop and enforce a fe energ balances. Either charge or refrigerant subcooling can be imposed, though subcooling as imposed for all the ork carried out here. The expansion device is an idealized device that can achieve a given evaporator outlet superheat. In the case of the secondar loop, the solver is also used to find a secondar loop temperature. 2.3 Radiant Panel Model Figure 2: Schematic of radiant panel model The radiant panel is formed of a number of tubes ith iet and outlet manifolds. In practice, the tubes ma be bent to fit the contours of the space it is installed in, but for the purposes here, the tubes are assumed to all be straight. The radiant panel is then sized based on the area available for the panel. If the total area available for the panel is given b Apanel, then the length of the tubes can be obtained from A panel Ltube (1) tube tube N tubes 1 2 ACHP 1.3: CoolProp: Fluid properties for the masses. International Refrigeration and Air Conditioning Conference at Purdue, Jul 16-19, 2012

5 2597, Page 4 here tube-tube and Ntube are the tube-to-tube centerline distance (in meters) and the number of tubes forming the panel respectivel. r instance, the average floor space of American homes in 2010 as 222 m2 [2,392 ft2]3. If 80% of the available area is used for the radiant paneling and there are 10 tubes ith m [10 inch] tube-to-tube centerline distance, then the tubes ould need to be 69.9 meters long. If the spreading heat conduction thermal resistance in the radiant panel is neglected, there are to thermal resistances that govern the heat transfer from the radiant panel to the surroundings. The overall heat transfer conductance can therefore be given b (2) UA h A air panel hater Ntubes S Di,tube Ltube ¹ hich includes the convective thermal resistance in the tubes as ell as the combined radiant and convective air-side thermal resistance hair. The overall air-side heat transfer coefficient is usuall on the order of 10 W/m2/K. With the value for UA knon, the ater outlet temperature can then be obtained from (Bergman, 2011) (3) and the heat transfer rate is then given b (4) The pressure drop in the radiant panel is governed b the internal flo pressure drop relations as described in the ACHP documentation. 2.4 Air-To-Air Heat Pump Sstem Analsis The required model parameters for the R410A air-to-air heat pump sstem are summarized in Table 1 (heat exchangers) and Table 2 (other parameters). The heat exchanger parameters ere obtained from the analsis of Shen (2006). This sstem is a nominal 3-ton cooling capacit sstem. Table 1: Heat Exchangers for R410A sstem based on Shen (2006) Tubes per bank [-] Number of bank [-] Number of circuits [-] Length of tube [m] Tube D [m] Tube ID [m] Longitudinal tube pitch [m] Transverse tube pitch [m] Fin Tpe Fins/inch [1/in] Tice fin amplitude [m] ½ period of fin aviness [m] Fin thickness [m] Fin conductivit [W/m/K] Humid air volume flo rate [m3/s] Atmospheric pressure [kpa] Relative humidit [-] Fan poer [W] 3 Evaporator Wav Lanced Condenser Wav Lanced International Refrigeration and Air Conditioning Conference at Purdue, Jul 16-19, 2012

6 2597, Page 5 Table 2: ther parameters for R410A sstem Value R410A Scroll Parameter Refrigerant Compressor Tpe Condenser outlet subcooling [K] Evaporator outlet superheat [K] Line set length [m] 2.5 R410A Heat Pump Performance Results ANSI/AHRI Standard 210/240 governs the rating of unitar heat pump units, and provides a fe rating points for hich heat pump manufacturers must provide performance data. The ASHRAE standard rating point H1 is used in this stud as the conventional rating point. The H1 rating point emplos an 8.33 C [47 F] air iet temperature to the evaporator, and a 21.1 C [70 F] air iet temperature to the condenser. Standard 210/240 provides rating points as lo as C [17 F] air iet temperature to the evaporator (rating point H3). Table 3 summarizes the results for the three rating points. Both CSP and Capacit are ver nearl linear ith evaporator air iet temperature. Table 3: Modeling Results for the R410A Heat Pump Rating Point H1 H2 H3 Condenser Air Iet Temp C [70 F] 21.1 C [70 F] 21.1 C [70 F] Evaporator Air Iet Temp C [47 F] 1.66 C [35 F] C [17 F] CSP Capacit W 8341 W 6242 W 2.6 Secondar Loop Sstem In order to provide a fair comparison beteen the to sstems, the same operating conditions have been used for both sstems. In the secondar loop sstem, the condenser is replaced ith a plate heat exchanger described b the geometr in Table 4. This plate heat exchanger as selected based on prior analsis of secondar loop heat pumps in cooling mode in order to ield ell-controlled pressure drops on both fluid sides as ell as good heat transfer. Table 4: Geometr of Plate Heat Exchanger for Secondar Loop Sstem Parameter Number of plates [-] Bp [m] Lp [m] Plate Amplitude [m] Plate Thickness [m] Plate Conductivit [W/m/K] Plate Wavelength [m] Inclination Angle [deg] Value The saturation pressure of propane at 0 C (474 kpa) is quite a bit loer than that of R410A (800 kpa). As a result, the number of circuits in the evaporator must be increased in order to have a ell-controlled pressure drop in the evaporator. The number of circuits in the evaporator as increased to 12, otherise all the parameters of the secondar loop evaporator are the same as the evaporator in Table 1. The ater flo-rate through the secondar loop sstem as set at 0.38 kg/s. This mass flo rate is ver nearl the optimal mass flo rate for the 47 F evaporator iet air temperature condition. The same length of piping connecting the outdoor and indoor units as emploed as the conventional heat pump sstem, and the pump as given an overall (pump+motor) efficienc of 0.5. International Refrigeration and Air Conditioning Conference at Purdue, Jul 16-19, 2012

7 2597, Page 6 A propane compressor as sized for this application using a compressor map provided b the manufacturer. The compressor displacement of the propane secondar loop sstem as scaled slightl in order to ield the same capacit as the R410A sstem at the H1 rating point. The propane compressor displacement (and therefore mass flo rate and electrical poer) as decreased b 5.3% hen the scaling parameter as introduced. 2.7 Performance Comparisons Figure 4: Condensing temperature of the sstems as a function of the air iet temperature to the evaporator Figure 5: Heating capacit of the sstems as a function of the air iet temperature to the evaporator Figure 6: Compressor overall isentropic efficienc for the sstems as a function of the air iet temperature to the evaporator Figure 3: CSP of the sstems as a function of the air iet temperature to the evaporator As can be seen from Figure 3 to Figure 6, b ever important performance metric, the performance of the propane secondar-loop sstem is better than that of the R410A air-to-air heat pump. Figure 3 shos the CSP for both the secondar loop heat pump and the direct expansion heat pump. These results sho that over the entire range of evaporator air iet (ambient) temperatures, the CSP of the secondar loop sstem is better than that of the direct expansion sstem. There a number of factors contributing to the improved performance ith the secondar loop sstem. r one, in the secondar loop sstem, the fan poer required for the condenser (hich is quite significant) is removed. In addition, the condensing temperature of the secondar loop sstem is significantl loer as shon in Figure 4 because of the large heat transfer area in the space. This leads to reduced compressor electrical poer input. These to poer reductions far outeigh the additional pumping poer required for the secondar loop International Refrigeration and Air Conditioning Conference at Purdue, Jul 16-19, 2012

8 2597, Page 7 Figure 5 shos the heating capacit of the sstem as a function of the air iet temperature to the evaporator. r all temperatures investigated here, the capacit of the secondar loop sstem is better than the direct expansion sstem. Although the compressor for the secondar loop sstem as not specificall selected in order to match the loer condensing temperatures, the compressor map predicts a higher overall isentropic efficienc, as seen in Figure 6. Figure 7: CSP of secondar loop sstems as a function of ater mass flo rate Figure 7 shos the CSP of the secondar loop sstems for a range of ater mass flo rates and evaporator air iet temperatures. The sstem CSP is not ver sensitive to the mass flo rate of ater beond about 0.20 kg/s; as a result, a fixed mass flo rate of ater can be used over the full operating envelope ith ver little decrease in sstem efficienc. All other parameters ere unchanged from the comparisons above. Further optimization of the mass flo rate of the secondar loop can be achieved b determining the optimal mass flo rate for a selection of evaporator air iet temperatures (shon ith the circular markers in Figure 7) and then conducting a linear fit of the optimal mass flo rate versus the evaporator air iet temperature. The optimal mass flo rate for this particular sstem as a function of evaporator air iet temperature can therefore be given b (5) here Ti,evap is in C and is in kg/s. Figure 8 shos that as long as the ater flo rate is high enough, there is ver little impact on the heat pump capacit. The use of the optimal ater flo rate from Equation (5) ould ield capacities that are near the maximum and ould allo for a simple control strateg if the secondar loop pump speed ere controllable. International Refrigeration and Air Conditioning Conference at Purdue, Jul 16-19, 2012

9 2597, Page 8 Figure 8: Capacit as a function of secondar loop ater flo rate and air iet temperature to evaporator 3. SYSTEM CMPARISNS It is useful to consider the energ flos and other sstem parameters for the rating point H1 in order to clarif the differences beteen the secondar loop options. Table 5: Summar of sstem parameters for operation at rating point H1 Air-Source DX Secondar Loop Radiant panel area [m2] Compressor Poer [W] Evaporator Fan Poer [W] Condenser Fan Poer [W] Pump Poer [W] Net Poer [W] Sstem Capacit [W] Compressor pumped vol. rate [m3/s]4 Evaporator air-side area [m2] Condenser air-side area [m2] Plate Heat Exchanger area [m2] With the use of the secondar loop sstem, the relativel high condenser fan poer of the DX sstem is traded for a much loer secondar loop pump poer, hich reduces the energ consumption b 411 W. The decreased condensing temperature of the secondar loop sstem results in a further 491 W reduction in compressor input poer. The compressor pumped volumetric rate for the secondar loop sstem is 68% larger than that of the R410A compressor, due to the fact that the densit of the propane is loer than that of R410A. Thus, at the same evaporation temperature, and for the same heating capacit, a larger volumetric flo rate is required for the propane compressor. The cause of the loer condensing temperature for the secondar loop sstem can be clearl seen from a comparison of the condenser air-side area of the DX sstem and the radiant panel dimensions of the secondar loop 4 Based on the compressor map; pumped mass flo rate divided b densit at compressor iet. Includes the volumetric efficienc implicitl since the volumetric efficienc is built into the compressor map International Refrigeration and Air Conditioning Conference at Purdue, Jul 16-19, 2012

10 2597, Page 9 sstem. The available area for heat transfer of the radiant panel is 3.6 times greater than that of the condenser installed in the ductork. Hoever, the radiant panel is part of the building construction (i.e., floor) and the tubing can be lo-pressure plastic. Therefore, the cost per unit area of materials for the radiant panel heat exchanger should be significantl loer than that for the conventional condenser. 3.1 ther considerations The use of a secondar loop sstem is a natural first step to a multi-zoned heating sstem. Robust zone temperature control can therefore be achieved through the use of variable ater flo rates to the zones, each zone getting a radiant panel. Since ater is used as the fluid floing to the zones (rather than a refrigerant), the pressure drop can be ell-controlled for long line sets The secondar loop sstem could also be used in cooling mode in dr climates or if a separate moisture control sstem is emploed, though that is not considered here. The performance of secondar loop sstems in cooling mode has been shon to be competitive ith, if not better than, the performance of direct expansion R410A sstems. CNCLUSINS The benefits to sstem efficienc through the use of a secondar-loop hdronic heat pump are quite significant. The increase in coefficient sstem of performance can be as high as 38%. This suggests that it should be straightforard to design a hdronic heat pump that can easil achieve the same seasonal performance as a conventional HFC-based air-source heat pump. Furthermore, this technolog ould be easil adapted for cooling mode operation and/or multi-zone sstems. Units m2 J/kg/K m Description Area of panel Mass specific heat Inner diameter of tube hair W/m2/K Air mean heat transfer coefficient 2 hater W/m /K Water mean heat transfer coefficient Ltube m kg/s Length of panel Water mass flo rate kg/s ptimal ater mass flo rate K K K K K m W/K Number of tubes Evaporator air iet temperature Condensing temperature Water iet temperature Water outlet temperature Ambient temperature Tube-tube centerline distance verall heat transfer conductance Ntubes Ti,evap Tcond T,i T,o T tube-tube UA Parameter Apanel cp Di,tube NMENCLATURE REFERENCES AHRI, 2012: Central Air Conditioners and Air-Source Heat Pumps Historical Data, Bergman, T.; Lavine, A.; Incropera, F.; Deitt, D.; Fundamentals of Heat and Mass Transfer 7 th Edition, Wile Calm, J. M. & Hourahan, G. C. 2007, Refrigerant Data Update, HPAC Engineering v. 79, International Refrigeration and Air Conditioning Conference at Purdue, Jul 16-19, 2012

11 2597, Page 10 Gong, X.; Claridge, D.; 2002, Impact of the Position of the Radiators on Energ Consumption and Thermal Comfort in a Mixed Radiant and Convective Heating Sstem, ASHRAE Transactions. lesen, B., 2002, Radiant Floor Heating In Theor and Practice, ASHRAE Journal, Jul 2002 Shen, B., 2006, Improvement and Validation of Unitar Air Conditioner and Heat Pump Simulation Models at ffdesign Conditions ASHRAE Final Report 1173-RP, International Refrigeration and Air Conditioning Conference at Purdue, Jul 16-19, 2012 Poered b TCPDF (.tcpdf.org)

Performance Comparison of Hydronic Secondary Loop Heat Pump and Conventional Air-Source Heat Pump ABSTRACT 1. INTRODUCTION

2597, Page 1 Performance Comparison of Hydronic Secondary Loop Heat Pump and Conventional Air-Source Heat Pump Ian H. BELL 1 *, James E. BRAUN 2 1 Bell Thermal Consultants ian.h.bell@gmail.com 2 Purdue