Development of high-efficiency carbon dioxide

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Development of high-efficiency carbon dioxidecommercial heat pump water heaterMichael Petersen a , Chad Bowers a , Stefan Elbel a b & Pega Hrnjak a ba Creative Thermal Solutions, Inc., 2209 North Willow Road, Urbana , IL 61802 , USAb Department of Mechanical Science and Engineering , University of Illinois , Urbana-Champaign , IL , USAAccepted author version posted online: 19 Aug 2013.Published online: 25 Oct 2013.

To cite this article: Michael Petersen , Chad Bowers , Stefan Elbel & Pega Hrnjak (2013) Development of high-efficiencycarbon dioxide commercial heat pump water heater, HVAC&R Research, 19:7, 823-835

To link to this article: http://dx.doi.org/10.1080/10789669.2013.833543

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HVAC&R Research (2013) 19, 823835Copyright C 2013 ASHRAE.ISSN: 1078-9669 print / 1938-5587 onlineDOI: 10.1080/10789669.2013.833543

Development of high-efficiency carbon dioxide commercialheat pump water heater

MICHAEL PETERSEN1,, CHAD BOWERS1, STEFAN ELBEL1,2, and PEGA HRNJAK1,2

1Creative Thermal Solutions, Inc., 2209 North Willow Road, Urbana, IL 61802, USA2Department of Mechanical Science and Engineering, University of Illinois, Urbana-Champaign, IL, USA

Although heat pump water heaters are widely accepted in both Japan and Europe, where energy costs are high and governmentincentives for their use exist, acceptance of such products in the United States has been limited. This trend is slowly changing withthe introduction of heat pump water heaters into the residential market, but acceptance remains low in the commercial sector.The objective of the presented work is the development of a high-efficiency R744 heat pump water heater of approximately 35-kW (10-ton) heating capacity for commercial applications with effective utilization of the cooling capability for air conditioningand/or refrigeration. This unit will be targeted at commercial use where some cooling load is typically needed year round, suchas restaurants, hotels, nursing homes, and hospitals. The improvement process concentrated on the heat exchangers of the system.Further optimization potential was identified by investigating the gas cooler as well as the expansion device of the heat pump waterheater by using a two-phase ejector. In addition, a comparison to a commercially available baseline R134a unit of the same capacityand footprint was made where significant package size reduction potential of the R744 heat pump water heater was discovered aswell as performance improvement, especially at high water temperature lifts.

Introduction

In todays world with increasing costs of energy, heat pumpwater heaters (HPWHs) offer great potential to reduce en-ergy consumption in water heating applications. HPWHs arewidely accepted in both Japan and Europe, where energy costsare high and government incentives for their use exist. Accep-tance of such a product in the United States has been slow,with approximately 15,000 units sold in 2009 (D & R Inter-national 2010) compared to a few hundred thousand per yearin Japan with the help of generous government and utilityincentives (Heat Pump & Thermal Storage Technology Cen-ter of Japan [HPTCJ] 2012). Barriers to HPWH acceptancehave historically been performance, reliability, as well as initialcosts. The dominant styles of water heaters used today in theUnited States are still electric resistance and gas, split roughly50/50 in market share (D & R International 2010). The tech-nology for these systems is quite mature, and all have primary

Received January 31, 2013; accepted August 6, 2013Michael Petersen,MSc, is Research Engineer.ChadBowers, PhD,Associate Member ASHRAE, is Senior Research Engineer. Ste-fan Elbel, PhD, Member ASHRAE, is Adjunct Professor at De-partment of Mechanical Science and Engineering and Chief En-gineer at Creative Thermal Solutions. Pega Hrnjak, PhD, FellowASHRAE, is Research Professor in Department of Mechani-cal Science and Engineering and President of Creative ThermalSolutions.Corresponding author e-mail: [email protected]

energy efficiencies less than 1. Commercial water heaters arerated by a combination of thermal efficiency and passive lossesto ambient, called stand-by losses. Heat pumps used for waterheating are essentially refrigerationmachines and have been inuse for many years. Thermal efficiency, or heating coefficientof performance (COP), of electrically driven heat pumps usingconventional refrigerants (R22, R134a, R410A) are in the 3to 5 range compared to 0.8 to 0.95 seen in gas and electricresistance water heaters, respectively.

The use of the natural refrigerant R744 as a refrigerantfor heat pumps is relatively new, starting first in the automo-tive industry but quickly moving to residential HPWHs inthe mid-1990s in Japan with sponsorship from the Japanesegovernment. Many Japanese manufacturers now have fullycommercialized residential R744 HPWHs, collectively calledEco-Cute, which have become fully accepted as the mostpromising technology for reducing Japans dependence on oil.Many studieswere published in the last two decades investigat-ing the potential and improvement of R744 refrigeration cy-cles focusingmainly on the heat exchangers and the expansiondevice. For heat pumps, one of the most attractive features ofusingR744 as theworking fluid is the utilization of the temper-ature glide in the gas cooler of the system, which reduces lossesin the heating process as described byNeksa et al. (1998). Stene(2005) used a tripartite heat exchanger to reduce conductionlosses in the gas cooler and to utilize different temperature lev-els for hot water as well as space heating. Stene (2007) notedthat using R744 allowed for the production of hot water in thetemperature range from 60C to 85C (140F to 185F) whileobtaining the highest possible COP for an HPWH system.

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Kim et al. (2005) described the influence of an internal heatexchanger (IHX) on an R744 HPWH system performance.They saw COP improvement while capacity decreased due totrade-offs between effectiveness and pressure drop in the IHX.Sarkar (2006) found the main COP-influencing parameters ofanR744 heat pump system for simultaneous heating and cool-ing to be compressor speed and discharge pressure as well ascoolant temperatures for evaporator and gas coolers. The op-timization process of an ejector expansion R744 heat pumpcycle was investigated by Sarkar (2008), where such ejector pa-rameters as entrainment ratio and pressure lift and their effecton COP were described. A theoretical comparison of R744and R134a in a tap water heat pump was described by Cecchi-nato et al. (2005), where competitive results can be achievedfor R744 by using its beneficial properties.

This article presents the development and improvementprocess of anR744HPWH that was performed in several stepsinvestigating the heat exchangers and the expansion device ofthe system. This HPWH was compared to a commerciallyavailable R134a HPWH of the same footprint. The retrofitof the working fluid R134a with R744 was done to improvesystem performance characteristics and, at the same time, toshow the influence of the higher volumetric capacity of R744compared to R134a.

Experimental facility

The experimental facilitywasdesignedaccording toASHRAEStandard 118.1 (ASHRAE 2003), which describes the testing

of Type IV HPWHs that can be operated without a waterstorage tank. The heat pump system was installed in an envi-ronmental chamber in order to provide steady ambient condi-tions. The unit was instrumented in such a way as to achieveenergy balances on the cooling side and heating side of the cy-cle. The accuracy of the experimental results was within 5%for the COP on the cooling and heating sides. On the coolingside of the cycle, the balance was achieved on the air streamand refrigerant stream. The cooling capacity on the air streamwas determined using a separate wind tunnel directly con-nected to the evaporator air discharge of the heat pump unit.This wind tunnel was built and instrumented according toASHRAE Standard 37 (ASHRAE 2005). The heat pump unitwas equipped with a blower to provide airflow over the evapo-rator. The power consumption of this blower as a function ofair temperature was determined and added to the compressorpower to receive the total HPWH power consumption. Afterdetermining the blower power consumption, it was removedfrom the unit because it was not strong enough to generate thenecessary pressure head to overcome the pressure drop causedby the flow nozzles used to determine airflow rate. Airflowwasprovided by the wind tunnel blower that was able to providethe pressure lift required to maintain a constant airflow rateof 1800 l/s (3800 CFM) for all tests and in all system configu-rations. The second determination of the cooling capacity ofthe system was made through measurements obtained on therefrigerant flow stream of the evaporator of the system. As anexample of the installed instrumentation in the experimentalfacility, a schematic of the R744 HPWH with IHX is shownin Figure 1.

Fig. 1. Experimental facility.

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Table 1. Instrumentation.

Description Range Accuracy

Temperature T-type thermocouple 200C350C (328F662F) 0.5 KPressure Refrigerant: Sensotec TJE 0138 bar (02000 PSIG) 0.1% full scale

Air: Setra 265 0625 Pa 1% full scale(02.5 W)

Mass flow Water/glycol: MicroMotion DS100 0455 kg/min (01000 lbs/min) 0.05% (liquid) of rateMass flow Refrigerant: MicroMotion DS25 020 kg/min (045 lbs/min)Humidity General Eastern: Model D2 35C25C Td (1%100% RH) 0.2CPower Ohio Semitronics: GW5-008DY22 060 kW (017ton) 0.2% full scale

On the heating side of the system, three different energydetermination methods were employed. Using the instrumen-tation on the refrigerant cycle described above, the heatingcapacity can be determined from the temperatures, pressures,and mass flow values at the inlet and outlet of the con-denser/gas cooler. The second heating capacity determinationwas made using temperature and mass flow measurements onthe water stream. In order to reject the heat input into thewater stream by the heat pump, a glycol chiller was used. Thethird determination of heating capacity was made on this gly-col stream. For this, the glycol temperature was measured todetermine the specific heat necessary for the capacity deter-mination. The glycol/pump facility was designed and con-structed to transfer heat between the hot water stream and thecold glycol stream through a brazed plate heat exchanger. Inaddition to the brazed plate heat exchanger, the facility wasalso equipped with mass flow meters for each fluid stream,trim heaters, and a pump to control the water flow rate. Theinstruments that were used and their accuracy are summarizedin Table 1.

Investigated systems

The investigation started with an evaluation of a commer-cially available R134a HPWHmarketed toward indoor appli-cations. This baseline unit was a packaged air-source HPWHwith a nominal heating capacity of 35 kW (10 ton) and COPof 3.9. A scroll compressor was used with a condenser ofbrazed plate design. The evaporator was a round tube platefin design, and the expansion device was a thermostatic ex-pansion valve. This baseline system was then compared to anR744 system of the same footprint. An R744 compressor thatwould provide similar capacity at the rating condition for thebaseline R134a system was chosen. This compressor was of asemi-hermetic reciprocating design. It has been demonstratedseveral times before (Bullard 2004; Elbel and Hrnjak 2008)that the performance of a transcritical R744 system can beoptimized using the high-side pressure. An electronic expan-sion device was used to vary the high-side pressure duringtesting.

In addition to possible performance benefits of using R744as the working fluid, the high volumetric capacity of R744enables the construction of the same capacity unit with a sub-stantially reduced systemvolume.Displaying this unit package

reduction potential was a secondary aim of this work besidethe COP improvement process. The volume reduction wasdemonstrated in the reduced height of the unit, which wasprimarily achieved through a more compact heat exchangersize. The reduction of the evaporator volume was achieved byusing a round tube plate fin evaporator coil with a fin densityof 16 fins per inch and a tube diameter of 3/8 in., originallydesigned for a 17-kW (4.8-ton) R134a system environmentalcontrol unit where it was used successfully, as described by El-bel andHrnjak (2010). This resulted in a 40% reduction in facearea, primarily in height, and a 55% reduction in evaporatorvolume.

The first step in the system improvement process focused onthe modification of the heat exchangers of the system, namelythe IHX, the evaporator, and the gas cooler. The IHX con-sisted of microchannel tubes assembled in a sandwich config-uration having the high-temperature stream in the center andlow-temperature streams on the outside. Two identical compo-nents in parallel set up in counter flowwere used in order to getsufficient capacity for the heat transfer from the high-pressurerefrigerant after the gas cooler to the low-pressure refrigerantafter the evaporator. An improved IHX of brazed plate designwas used in a later step of the improvement process. This com-ponent had similar capacity but lower pressure drop; it wasdesigned to provide an effectiveness of approximately 70%.

For the following improvement steps of the R744 HPWHrefrigeration system the full R134a housing volume was used.The additional height allowed the use of a larger evaporator.For this purpose, the baseline R744 evaporator volume wasenlargedby 50%by combining itwith half of a second identicalcomponent creating an evaporator of almost the same heightas the R134a evaporator. A comparison of the dimensions ofthe three evaporators that were used is shown in Figure 2.

The R744 baseline gas cooler was a commercially availablemodel with a much narrower design compared to the R134acondenser. The reduction of the gas cooler volume was ap-proximately 50%. In the following system development steps,a performance improvement potential was identified by reduc-ing conduction losses, especially at high water outlet tempera-tures, as described by Kim et al. (2004). For this, a staged gascooling process using multiple brazed plate heat exchangerswas used to reduce conduction effects compared to one com-pact gas cooler. Four brazed plate heat exchangers were usedwith 9 kW (2.5 ton) each. A comparison of the dimensions ofthe three heat exchangers is shown in Figure 3.

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Fig. 2. Evaporator dimensions.

The previously described R744 HPWH systems wereequipped with two electronic expansion valves (EEVs) thatwere set up in parallel to provide sufficient refrigerant flow.These valves allowed convenient control of the high-side pres-sure during operation. Part of the improvement process wasthe use of an electronically actuated ejector that was used to re-duce throttling losses during the expansion process. A systemschematic and the corresponding cycle plotted in the pressurespecific enthalpy diagram describe the main differences whenusing an ejector compared to an expansion valve, as shown inFigure 4.

The ejector system consists of two sides. A driven low-pressure side with the suction flow and a driving high-pressureside with the motive flow. The high-pressure side containsthe gas cooler and high-pressure side of the IHX. After theIHX, the motive flow enters the ejector. The ejector can becompared to a pump in which the high-pressure motive flowpumps the low-pressure suction flow. The motive flow entersthe ejector and is accelerated in the motive nozzle. The suctionflow is entrained by the high-velocity motive flow in the suc-tion chamber and both streams mix in the mixing chamber.

Fig. 3. Gas cooler/condenser dimensions.

In the diffuser, the kinetic energy of the refrigerant stream isconverted into pressure energy, creating an intermediate pres-sure that reduces the compressor power. After the ejector, therefrigerant flow enters the phase separator that divides the liq-uid and the vapor phases of the flow. The liquid fraction isexpanded and enters the evaporator as a low-quality suctionflow. The vapor flow enters the low-pressure side of the IHX,and the cycle starts over again.

For better control and more flexibility during testing, theejector was equipped with a stepper-motor-controlled needleto vary the area of themotive nozzle. The ejector is of modulardesign to allow themodification of the component dimensionsas described by Elbel et al. (2012). A schematic of the ejectorthat was used with its main sections and the two refrigerantstreams is shown in Figure 5.

The investigated R744 HPWH system configurations aresummarized in Table 2.

Results and discussion

The strategy of the investigation was a comparison of theR744 development stages to the R744 baseline system toconcentrate on the improvement potential of each step. Gascooler water inlet temperature and water flow rate varia-tions were investigated for the first four R744 developmentsteps (BEVAP [baseline evaporator], BEVAP+MC IHX

Table 2. Investigated R744 HPWH system configurations.

MC BPBEVAP EEVAP IHX IHX MGC EEV EJECTOR

1 X X2 X X X3 X X4 X X X5 X X X X6a X X X6b X X X

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Fig. 4. R744 HPWH ejector system schematic and cycle plotted in pressure specific enthalpy diagram (color figure available online).

[BEVAP+microchannel IHX], EEVAP [enhanced evapora-tor], and EEVAP+MC IHX). The investigation of the multi-ple gas coolers and the ejector was done at selected conditionsto show their performance improvement potential. Finally acomparison of the R744 performance to the R134a systemallowed a comparison of the two fluids. The R744 HPWH de-velopment stages were investigated at the condition the R134aunit as rated by the manufacturer. The rating temperature ofthe water and air inlet of the HPWH is 26.7C (80F). Thewater flow rate was adjusted to reach a water temperature liftof 5 K, which was a water flow rate of approximately 1760 g/s(3.9 lbs/s) for all systems. All systems were tested under theseconditions as well as under a broader range of water inlettemperatures as shown in Table 3.

The combined COP was the main interest of the investiga-tion. It was calculated according to Equation 1 as the ratioof the useful output (heating and cooling capacity) dividedby the system power consumption caused by compressor andsystem fan power consumption:

COPComb = QEVAP + QGCWHPWH . (1)

For the R744 systems, the high-side pressure at each condi-tion was optimized to provide the highest heating COP when

operating in the transcritical mode. Even though this highwater flow rate was required by the standard to rate the sys-tem, this may not represent actual operation of such units,especially if higher water temperature lift conditions are re-quired. In order to understand these effects of different wa-ter temperature lifts on system performance and gain insightinto a better control strategy, the water flow rate was reducedfrom the rating value of 1760 g/s (3.9 lbs/s) down to 1000and 400 g/s (2.2 and 0.9 lbs/s). This was done for water in-let temperatures of 12C, 26.7C, and 50C (54F, 80F, and122F) while maintaining the air inlet temperature at 26.7C(80F). This ambient conditionwas chosen because thismodelis marketed as an indoor unit, meaning the heat pump will bepumping heat from the buildings indoor environment to thewater stream. Applications in which this would be most ben-eficial are situations where there is a constant need for si-multaneous hot water and cooling. Restaurant kitchens andlaundry facilities are both excellent examples of such loca-tions. The following paragraphs describe theHPWH improve-ment steps by using different IHX designs and an enhancedevaporator. These improvements were investigated for differ-ent gas cooler water inlet temperatures and water flow rates.In addition, the improvement potentials when using multi-ple gas coolers and an ejector were determined for selectedconditions.

Fig. 5. Ejector schematic with main sections and refrigerant flows (color figure available online).

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Table 3. System performance test conditions.

Air side Temperature 26.7C (80F)Flow rate 1800 l/s (3800 CFM)

Water side Temperature 12C, 26.7C, 50C (54F,80F, 122F)

Flow rate 400, 1000, and 1760 g/s (0.9,2.2, and 3.9 lbs/s)

IHX and evaporator

An IHX allows heat transfer from the high-pressure refriger-ant after the gas cooler to the low-pressure refrigerant leavingthe evaporator. The heating and cooling capacity increase byfurther subcooling the refrigerant after the gas cooler, andtherefore, lower evaporator inlet quality is offset by the in-creasing compressor power consumption caused by highersuction temperatures. Overall, net thermodynamic benefitscan be seen for R744 systems (Kim et al. 2004) because theCOP-optimizing high-side pressure is lower when an IHX isused. The IHX that was used in the first system improvementstep was a microchannel heat exchanger.

The improvement on the evaporator side was accomplishedby increasing the component size by 50%. Consequently, thecomponent volume as well as face area were increased whilethe air face velocity decreased. The combinedCOPs of the fourR744 development steps that were investigated at an ambienttemperature of 26.7C (80F) and different gas cooler waterinlet temperatures as a function of water temperature lifts areshown in Figure 6.

The trends that can be seen in Figure 6 confirm the typi-cal behavior of HPWH systems caused by varying gas coolerwater inlet temperature and water flow rate. The system per-

Fig. 6. Comparison of combined COP as a function of watertemperature lifts at different gas cooler water inlet temperaturesfor R744 HPWH development stages.

formance decreases with increasing gas cooler water inlet tem-peratures. The refrigerant outlet temperature is determined bythe water inlet temperature because of no pinch temperatureduring single-phase heat transfer in the gas cooler. Therefore,a close approach between refrigerant outlet temperature andwater inlet temperature is very important for an optimum incombined COP. Pettersen et al. (1998) described a 5% coolingCOP increase per degree of approach temperature reductionin transcritical operation in the gas cooler. Higher water inlettemperatures with high compressor discharge pressure lev-els showed lower combined COP values because more inputpowerwas needed compared to lowerwater inlet temperatures.The combined COP was more dependent on water tempera-ture lifts at low water inlet temperatures. At 50C (122F) wa-ter inlet temperature, the results were almost steady, whereasat 12C (54F), the COP dropped significantly when goingto larger water temperature lifts. This larger COP gradient atlow water inlet temperatures can be explained with the largetemperature differences in the gas cooler and the resulting po-tential losses. At high water inlet temperatures, the refrigerantgas cooler outlet temperature varied less at increasing watertemperature lifts compared to lower water inlet temperatures.Therefore, a much stronger decreasing effect on the COP wasseen at low water inlet temperatures compared to high waterinlet temperatures. In other words, the COP level at the 50C(122F) condition was already relatively low compared to the12C (54F) condition, which was much more sensitive to adecreasing water flow rate that created larger temperature lifts.Overall, the EEVAP+MC IHX system is considered the bestconfiguration for awide range of gas coolerwater inlet temper-atures and water flow rates. Even though the EEVAP systemshows good performance for different conditions, its perfor-mance gets worse at increasing water inlet temperatures. TheIHX system suffers from the low-side pressure drop, especiallyat low water inlet temperatures that cause a small compressorpressure ratio. However, at high water inlet temperatures withlarge pressure ratios, the relative pressure drop effect is lesssevere, which helps to compensate the performance decreaseat high water inlet temperatures.

In order to better understand the influence of the heatexchanger modifications (IHX, evaporator) on the sys-tem performance, the cycles were compared in an R744pressure-specific enthalpy diagram (Figure 7).

The systems without an IHX (BEVAP, EEVAP) operatedat a higher high-side pressure than the systems with an IHX(BEVAP+MC IHX, EEVAP+MC IHX). For the systemswithout an IHX, a constant superheat at the evaporator outletof 5 K was used. This was done to prevent liquid refrigerantfrom entering the compressor, which could lead to damage.The IHX systems operated at a lower optimized high-sidepressure resulting from more subcooling and therefore lowerevaporator inlet qualities. The evaporator outlet conditionwaskept constant at a quality of 0.95 by adjusting system chargein order to receive moderate superheat on the compressorsuction side. The superheat of the systems with IHX was ap-proximately 16 K. This larger amount of superheat caused alower suction density at the compressor inlet, which led tosmaller refrigerant mass flow rate.

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Fig. 7.Comparison of R744HPWH cycles in pressure versus spe-cific enthalpy diagram at rating condition (color figure availableonline).

The full potential of cycle improvements when using theIHX was not realized due to a relatively high pressure dropon the low-pressure side (Figure 8). The pressure drop on thesuction side was examined for all four systems in order tocompare the influence of the IHX. For the systems withoutan IHX, the low-side pressure drop was measured across theinlet and outlet of the evaporator. For the cycle with IHX, thelow-side pressure drop was measured between evaporator andcompressor inlet.

Two important effects of the IHX on the system can beseen in Figure 8. As described earlier, the systems sufferedfrom a lower suction density, which decreased the refrigerantmass flow rate. The second effect was a large pressure dropprior to the suction of the compressor. This pressure drop hadto be compensated during the compression process, increasingthe compressor power consumption.Consequently, the overallcombined COP increases were offset. A minimization of these

Fig. 8.R744HPWH suction pressure drop at different refrigerantmass flow rates at rating water and air temperatures.

Fig. 9. Pressure ratio versus water temperature lift at rating waterand air temperatures.

pressure losses would have further improved the COP valuessignificantly.

Another characterizing parameter used to analyze the in-fluence of the IHX was the pressure ratio of the system, whichwas calculated by the ratio of compressor discharge and suc-tion pressure according to Equation 2:

= pCp,OutpCp,In

. (2)

The pressure ratio at different water temperature lifts at anair and water inlet temperature of 26.7C (80F) is shown inFigure 9.

The pressure ratios of the systems with enhanced evapo-rator (EVAP and EVAP+IHX) were less dependent on thewater temperature lift. The increased low-side pressure forthese systems kept the pressure ratio on a lower level, whichwas beneficial for the compressor power consumption. Thesystem without an IHX showed the best COP of the investi-gated systems at this condition due to the offsetting effect ofthe low-side pressure drop in the IHX.

The evaporator approach temperature difference as definedin Equation 3 describes the temperature difference between airoutlet and refrigerant inlet saturation temperature at evapo-rator inlet pressure. The saturation temperature was used toget a good comparison for all four systems because differentevaporator outlet conditions were used as described earlier:

TEVAP,Approach = TEVAP,A,Out TEVAP,R,In,Sat. (3)

The smaller the approach temperature difference betweenthe two fluid streams was, the more effective was the heat ex-changer. An approach temperature difference of 0 represents atheoretical heat exchanger with infinite heat transfer area. Theresults for the evaporator approach temperature difference areshown in Figure 10. The dominant positive effect on the evap-orator approach temperature difference is the increase of the

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Fig. 10. Comparison of evaporator approach temperature differ-ence at different water temperature lifts for R744HPWHat ratingwater and air temperatures.

evaporator size from the BEVAP to the EEVAP systems. Thesystems with an IHX reach even lower approach temperaturedifferences mainly due to the improved heat transfer charac-teristics caused by lower evaporator inlet qualities.

Another characteristic factor that had to be consideredfor the modification of the evaporator was the overall heattransfer coefficient U , which was calculated for the air sideby using the air-side heat transfer area and the logarithmicmean temperature difference (LMTD) of the evaporator, asdescribed in Equation 4:

UEVAP,air = QEVAPAEVAP,A LMTD , (4)

where

LMTD

= (TEVAP,A,Out TEVAP,R,Out,Sat) (TEVAP,A,In TEVAP,R,In,Sat)ln (

TEVAP,A,OutTEVAP,R,Out,Sat)(TEVAP,A,InTEVAP,R,In,Sat)

.

(5)

The airflow rate was kept constant during the tests. That iswhy the results for the overall heat transfer coefficient U areshown as a function of refrigerant mass flow rate in Figure 11.

The evaporator was modified to utilize the volume that wasprovided by the R134a HPWH housing. For this, the evapo-rator volume was enlarged by 50%. This larger heat exchangervolume led to an increase in heat transfer area, which helped indecreasing the approach temperature difference between theair and refrigerant side. The low-side pressure increased,whichreduced the necessary compressor work to provide the high-side pressure. The trends for evaporator approach temperaturedifference and overall heat transfer coefficient confirmed thebehavior that was expected. The increase of the heat transferarea reduced the approach temperature difference and there-

Fig. 11. Evaporator overall heat transfer coefficient versus refrig-erant mass flow rate for R744 HPWH at rating water and airtemperatures.

fore increased the evaporation pressure. At the same time, theoverall heat transfer coefficient was reduced.

The experimental test results were used to make predic-tions about HPWH performance at large water temperaturelifts, creating high gas cooler water outlet temperatures. Forthis, the trends that were obtained from the test results of theR134a system were extrapolated to make predictions aboutthe systems energy savings potential of R744 over R134a as afunction of water outlet temperature (Figure 12). The R134asystem cannot reach the high water temperatures requiredin sanitary applications without additional supplement heat-ing due to a limit of the compressor discharge temperatureof 107C (225F) specified by the manufacturer. That is whyan additional electric resistance heater was used to provide

Fig. 12. Energy savings of R744 HPWH systems compared toR134a at various water outlet temperatures at rating water andair inlet temperatures.

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heating capacity for water outlet temperatures above 82C(180F). This meant that solely HPWH operation was doneup to a water outlet temperature of 82C (180F). Above thattemperature, a hybrid operation of combined HPWH usageand additional supplemental resistance heating with a COP of1 was used to make up the remaining temperature difference.

For the R744 system with internal heat exchange, up to a22% savings in energy consumption were achieved in compar-ison to the R134a system. The baseline R744 system achievedan energy savings potential of 9%, which can be seen above90C (194F). The assumption of a 100% efficient electric re-sistance heater does not represent the situation in real systemsand is therefore a conservative choice. For the situation with agas condensing water heater with an efficiency of 0.85 for theadditional heating capacity above 82C (180F), even largerimprovement potential can be seen. For the BEVAP system,an improved performance of 14%, and for the BEVAP+MCIHX system, 27% improvement at 95C (203F) water out-let temperature over R134a can be reached. The location ofthe crossover points in performance between R134a and R744were interesting for several reasons. The first and primary rea-son was that the temperature region in which R744 began toexcel over R134a was noticeably higher than what would beneeded in residential water heating applications but was ex-actly where the water needed to be heated to in commercialapplications. According to theU.S.Department ofHealth andHuman Services (Food Code 2005), fresh hot sanitizing wa-ter for ware-washing equipment should be no less than 74C(165F) for stationary, single-temperature machines and noless than 82C (180F) for all other machines. If a storagedevice were to be implemented, water would likely need to beproduced at an outlet temperature at or above 85C (185F) toaccount for losses encounteredduring storage.Currently, thesehigher temperatures are achieved through secondary electricresistance heating, starting somewhere between 43Cand60C(109F and 140F). In the case of R744, no secondary heatingsource would be needed; however, many installations wouldlikely still have this as a backup method of heating. In mostcases, even the primary heating source is electric or gas withan energy factor of 1 or less. In these cases, the use of any heatpump technology would result in significant energy savings,especially if the free cooling is used to augment existingbuilding cooling.

Gas cooler

The transcritical operation on the high-pressure side of theR744 refrigeration system is a gas cooling process with a glid-ing temperature instead of the constant temperature duringcondensation of the R134a system. This gliding gas coolertemperature profile is useful for heating applications wherethe hot refrigerant rejects heat to the cooling fluid. How-ever, there is also a larger potential for losses in the R744gas cooler compared to the R134a condenser caused by thevery large temperature differences between hot gas comingfrom the compressor and the cooling fluid (air, water). Theselosses are often caused by conduction from the hot refrigerantinlet to the cool refrigerant outlet. This heat bypass makes the

heat transfer process differ from the ideal counter-flow con-figuration and, therefore, increases the approach temperaturedifference between refrigerant and water side. In order to re-duce these effects, the heat path must be interrupted. This canbe done for refrigerant to air gas coolers by preventing con-duction through cutting fins, as described by Park and Hrnjak(2007) or for water heating applications by a stepwise heattransfer process using multiple heat exchangers. This stepwiseprocess not only reduces losses but also makes it possible touse the different water temperature levels for domestic hotwater or hydronic heating, as described by Stene (2005).

The staged heating processwas realized in theR744HPWHwith four brazed plate heat exchangers that had the same com-bined capacity as the single component. These heat exchangershad a smaller hydraulic diameter that created a larger pressuredrop compared to the single heat exchanger design. That iswhy only lower water flow rates were investigated. However,at the lower water flow rates, large temperature lifts were re-alized, and also the conduction effects that led to increasinglosses were most severe. Therefore, the investigation at theseconditions showed the greatest effects. One indicator for theheat exchanger performance is a small approach temperaturedifference between the refrigerant outlet and the water inlet, asdescribed in Equation 2; the importance of a close approachbetween refrigerant and cooling fluid for an optimum in COPwas described many times before (Pettersen et al. 1998; Kimet al. 2004):

TGC,Approach = TGC,R,Out TGC,W,In. (6)

A comparison between the R744 HPWH systems witha single gas cooler and multiple gas coolers is shown inFigure 13.

The approach temperature difference as a function of watertemperature lift for the single and multiple gas cooler systemsshowed the influence of conduction on the heat exchanger

Fig. 13. Comparison of gas cooler approach temperature dif-ference at ambient air and gas cooler water inlet temperatureof 26.7C (80F) and different water temperature lifts for R744HPWH development stages.

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832 HVAC&R Research

Fig. 14. Comparison of combined COP at gas cooler water inlettemperature of 26.7C (80F) and different water temperaturelifts for R744 HPWH development stages.

performance. The multiple gas cooler system showed muchlower approach temperature differences compared to the sin-gle gas cooler systems. These lower values represented bet-ter component effectiveness during the heat transfer process.Since only small water mass flow rates were investigated thatcreated large water temperature lifts, a direct comparison wasonly possible with the BEVAP and EEVAP systems. However,the trends that were gained from the other systems at lowerwater temperature lifts showed the larger approach tempera-ture differences compared to the multiple gas cooler system.The effect of the better heat transfer characteristics of themultiple gas cooler system on the system COP can be seen inFigure 14.

Performance improvement was seen for both combinedCOP and heating capacity when using multiple gas coolers.For large water temperature lifts, where conduction effectsare most severe, the multiple gas cooler system (EEVAP+MCIHX+MGC [multiple gas cooler]) showed the best combinedCOP values of the investigated R744 HPWH systems. A com-parison between the EEVAP+MC IHX+MGC system andthe BEVAP system at a water temperature lift of 63 K showsa 35.6% increase in combined COP for the multiple gas coolersystem. However, these good results were only achieved atlarge water temperature lifts, where the benefits through animproved heat transfer process outweigh the larger pressuredrop of the multiple gas cooler assembly. The situation wouldchange at smaller water temperature lifts, where the conduc-tion effect would be less severe and the large pressure dropwould become more dominant, which would make the valuesfor combined COP less competitive compared to the single gascooler systems.

Ejector

The performance of the ejector can be characterized by severalparameters, as described by Elbel (2007). The mass entrain-

Fig. 15. Suction pressure drop of MC IHX and BP IHX at differ-ent refrigerant mass flow rates at rating temperatures and waterflow rate.

ment ratio is the ratio of ejector suction mass flow and themotive mass flow:

= mSuctionmMotive

. (7)

The pressure lift describes the pressure difference between ejec-tor diffuser outlet pressure and evaporator outlet pressure.This pressure difference elevates the suction pressure com-pared to the evaporation pressure to an intermediate level andtherefore reduces the necessary compressor power:

pLift = pDiff ,Out pEVAP,Out. (8)

Both mass entrainment ratio and pressure lift are coupled.That means that optimizing these factors is a compromise be-tween good results for either of them. For example, by openingthe liquid valve after the phase separator,more refrigerant suc-tion mass flow is provided, which increases the entrainmentratio. However, at the same time, the evaporation pressureincreases, which reduces the ejector pressure lift.

The pressure lift and its effect on the system performanceimprovement show the importance of reducing the suctionpressure drop. That is why an enhanced IHX in brazed platedesign was chosen for the comparison between direct expan-sion (DX) and ejector system to realize the full potential ofthe ejector. A comparison of microchannel IHX and brazedplate IHX pressure drops as a function of refrigerant massflow rate showed the improved component performance (Fig-ure 15). The brazed plate design (EEVAP+BP IHX [brazedplate IHX]) created in average a 37% lower suction pressuredrop than the microchannel design.

The ejector efficiency is the main parameter to characterizethe performance of the component. It is defined as the ratio ofrecovered work rate to the maximum possible expansion work

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rate potential, as described by Elbel (2007):

Ejector = WRecWRec,Max

. (9)

In order to determine the effect of the ejector on the systemperformance, the ejector test conditions were matched closelywith the DX system. The ejector condition was matched forthe heating capacity and the evaporator outlet temperatureand pressure simultaneously. In a second run, the heating ca-pacity and the gas cooler outlet temperature and pressure werematched. This approach of the investigation was done to coverboth influences of high- and low-pressure sides on the systemperformance (Figure 16).

It must be pointed out that the ejector system performancewas not ideal due to low separation efficiencies in the phaseseparator. This caused very low phase separator outlet quali-ties on the vapor port. Ideally, the vapor and the liquid portare saturated, having 100% and 0% refrigerant qualities, re-spectively. However, the phase separator that was used onlyprovided lower separation efficiencies. That is why liquid re-frigerant left the separator through the vapor port and vaporrefrigerant left the separator through the liquid port. A phaseseparator with better separation efficiency would improve theejector and system performance. Lawrence and Elbel (2012)described the influence of separation efficiency on the systemCOP and pointed out that the COP improvement gained byusing an ejector can easily be canceled out by the effect of anineffective phase separator.

The test for the ejector and DX system were done at ratingwater mass flow rate at an ambient temperature of 26.7C(80F) and gas cooler water inlet temperatures of 37.8C and50C(100Fand122F).The characteristic ejector parametersthat were recorded are shown in Figure 17.

The values for efficiency with 21% and 24%, respectively,as well as the entrainment ratio with 59% and 67% confirmedthe good performance of the ejector. These were typical per-

Fig. 16. Ejector system suction and gas cooler exit conditionmatch with DX system (color figure available online).

Fig. 17. Characteristic ejector performance parameters.

formance values for ejector efficiency and entrainment ratio.Comparable results were reported by Banasiak et al. (2012),where a two-phase ejector in an R744 heat pump was used.The pressure lift resulted from the optimized high-side pres-sure that was adjusted by the stepper-motor-controlled nee-dle. The effect of the ejector on the system performance ispresented through the combined COPs in Figure 18.

The ejector system showed better combined COPs at thetwo investigated water inlet temperatures; at 37.8C (100F),an improvement of 5%, and at 50C (122F) water inlet tem-perature, an improved COP of 9% were seen.

Fig. 18. Ejector and DX system at matched gas cooler conditionsand heating capacities.

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834 HVAC&R Research

Conclusions

The development and improvement process of an R744HPWH for commercial applications was described. The in-vestigation aimed for a twofold objective. In the first step, thesystem package reduction potential when using R744 com-pared toR134awas shown by realizing amore compact designenabled by the beneficial fluid properties of R744. The compo-nent volume reduction was achieved mainly in the evaporatorof the system with a 40% smaller height and a 55% smallervolume compared to R134a. In the second step, the IHX andthe evaporator of the R744 HPWH were improved, creat-ing a total of four R744 HPWH systems. It was shown thatthe combined COP increased during the improvement processcompared to theR744 baseline system. Trends were developedbased on the experimental data to predict the performance ofR744 compared to R134a. An energy savings potential of 22%at a water outlet temperature of 95C was discovered for theR744 BEVAP+MC IHX system over R134a. This was mainlycaused by the necessary supplemental heating capacity thatthe R134a system requires above water outlet temperaturesof 82C due to compressor discharge temperature limitations.This hybrid operation of HPWH and electric resistance heat-ing is not necessary in the R744 HWPH that can provide largetemperature lifts in once through operation.

Further investigations were done to show the potential ofreducing conduction losses in the gas cooler by using multipleheat exchangers. Themultiple gas cooler system outperformedthe single gas cooler system with an enhanced evaporator aswell as the same systemwith an IHX for the combinedCOPby35% and 22%, respectively. This improved performance wasachieved at large water temperature lifts when the conductioneffects have the strongest impact, and therefore, beneficial heattransfer characteristics were more dominant than the largerpressure drop in the component.

Finally, the effect of an ejector was investigated that wasused to reduce throttling losses during the expansion process.The ejector test conditions were closely matched with a DXsystem. At 37.8C and 50C (100F and 122F) gas coolerwater inlet temperatures, the ejector system showed 5% and9% better combined COPs, respectively.

Nomenclature

BEVAP = baseline evaporatorBP = brazed plate heat exchangerCOP = coefficient of performanceDX = direct expansionEEV = electronic expansion valveEEVAP = enhanced evaporator (50% volume increase)EJECTOR = two-phase ejectorHPWH = heat pump water heaterIHX = internal heat exchangerLMTD = logarithmic mean temperature difference, CMC = microchannelMGC = multiple gas coolerp = pressure, kPaT = temperature, C

U = overall heat transfer coefficient, W/(m2KW = power, kW = efficiency, = entrainment ratio,

Subscripts

A = airApproach = approachBP = brazed plateComb = combinedd = dew pointDiff = diffuserEjector = ejectorEVAP = evaporatorGC = gas coolerIn = inletLift = liftMax = maximumMC = microchannelMotive = motive flowOut = outletR = refrigerantRatio = ratioRec = recoveredSat = saturationSuction = suction flowW = waterX = expansion valve

Acknowledgment

The authors gratefully acknowledge the support given by theU.S. Department of Energy. Project funding was made avail-able through ARRA 2009 under contract DE-EE0003981.The authors would also like to thank A.O. Smith Corporationfor their continued support in this project.

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