ice_e info pack 3 heat exchangers

8/13/2019 ICE_E Info Pack 3 Heat Exchangers

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The environmental benefitrelated to intrinsic negligible directeffect of carbon dioxide, when usedas a refrigerant, is sometimesstrongly mitigated by the low

efficiency of the transcriticalthermodynamic cycle.One of the ways to improve theenergy efficiency of CO2 systemsis through internal cooling of theCO2 leaving the gas-coolerthanks to heat transfer to therefrigerant at the evaporatoroutlet.Domanski et al. (1994) proposed athermodynamic approach for theevaluation of suction-line/liquid-line

heat exchange effect in thesubcritical refrigeration cycle. Theproposed considerations aregeneral and based on the analysisof the thermodynamic andthermophysical properties of therefrigerant and so they can beeasily applied also to transcriticalcycles. As a matter of fact, carbondioxide gains great benefit frominternal heat exchange, since themajor exergy loss is due to thethrottling process because of thelow critical temperature of this fluid;for this reason, vapour qualitieshigher than 0.5 can be observed atthe evaporator inlet when nointernal heat exchange is present.The consequent thermodynamicpenalization is furthermore coupledwith a heat transfer penalization,since dry-out for CO 2 is recognizedto occur at relatively low vapourqualities, typically x>0.50.6. Boeweet al. (2001) investigated

experimentally the performance of awell instrumented prototype of atranscritical mobile air-conditioningsystem using R-744 as the

refrigerant, a single stagecompressor and a heat exchanger.They found an increase on systemCOP up to 25%. Chen and Gu (2005) recognized the

internal heat exchanger efficiency tobe a key factor for the total systemefficiency: the higher the internalheat exchanger efficiency the higheris the improvement. In thisframework, the high vaporsuperheating related to the internalheat exchange can be an issue.The use of two stagecompression with inter-stagecooling performed by an externalfluid is considered as a possible

solution. In this Info Pack, the results of anexperimental investigationpresented in Cavallini et al (2006),is reported with the aim of pointingout the positive effect in energyconsumption obtained through theuse of an internal heat exchanger.

A test rig was built at the Universityof Padova for testing CO 2 equipment operating with air as thesecondary fluid at the gas-coolerand the evaporator and water at theinter-stage cooler. The effect ofinternal heat exchange on optimalgas-cooling pressure and the effectof inter-stage cooling oncompression efficiency wasexperimentally investigated.

TEST RIGThe CO 2 circuit (Figure 1) carriesout a double compression with gasintercooling between the twocompression stages and single

throttling. The compressor is a two-stage semi-hermetic reciprocatingunit running at 1450 rpm (50 Hz).The nominal swept volume of the

low pressure stage (one cylinder) is3.0 m 3h -1, while that of the secondstage is 1.74 m 3h -1 (vol. ratio 1.7).The lubricant is a PAG oil 46 ISOgrade. No oil cooler was installed.

The intercooler (IC) is a coppertube-in-tube heat exchanger withthe CO 2 flowing inside three pipes(ID 4 mm, OD 6 mm) fed in paralleland inserted inside a 20 mm ID (22mm OD) copper tube. Water flowsinside the outer tube in counter-flowto CO 2. A suction line-gas cooleroutlet line internal heat exchanger(SLHX) is installed. The highpressure fluid coming from the gascooler outlet feeds three copper

tubes (4 mm ID and 6 mm OD) inparallel. The fluid from theevaporator outlet flows in theannulus between the mentionedcopper pipes and a stainless steeltube (ID 20.9 mm). The total SLHXlength is 10 m, split in two stretches,each 5 m long, with one U bend.The throttling device used in thetests is a back-pressure valve: thisallows the operator to set and keepconstant the gas-cooler outletpressure. The R-744 circuit isequipped with an oil separator.

Air is the external fluid for both thegas-cooler (GC) and the evaporator.For the tests here reported finnedcoils are employed with roundcopper tubes (9.52 mm ID and10.82 mm OD), subdivided into twocircuits for both the gas-cooler (fourranks) and the evaporator (tworanks), with aluminum louvered fins(2.1 mm spacing). The face area forboth exchangers is 500x500 mm.

A detailed description of the test rigand of the measurement accuracyis reported in Cavallini et al. (2006).

All the measurements are real time

ICE-E NFORMATION

PACK The optimal heat rejection pressure in CO 2 transcritical systems


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Figure 1. Test rig lay-out. 1a,b two-stage compressor, 2 intercooler, 3gas-cooler, 4 evaporator, 5 back-

pressure valve, 6 nozzle, 7 auxiliaryfinned coil, 8 electrical heater, 9centrifugal fan, 10 metering valve,MF: Coriolis mass flow meter, OS:oil separator, P: pressuretransducer, SLHX: suction line heatexchanger, T: thermocouple, T:thermopile.

and elaborated: The accuracy andrepeatability of the reported testsfulfill ASHRAE S.TANDARD 116

(1995).

EXPERIMENTAL RESULTS

Several tests were run both withand without the SLHX: each of thedata here included represents theaveraged value of 30 min steadystate acquisition.The air dew point was always foundto be lower than the saturationtemperature corresponding to theR-744 pressure at the evaporator

outlet. All the measurements reported,both with and without SLHX, wereobtained with R-744 temperature at

the gas-cooler and the IC outlettGC,out =t IC,out =33.10.1C. The CO 2 pressure at the evaporator outletwas kept at 3.7390.15 MPa. Beingfixed the evaporator outlet pressure,three different conditions wereconsidered, by changing R-744temperature and vapor quality: Test A: no SLHX; evaporator

outlet temperature 810.0 C. Test B: SLHX present; evaporator

outlet saturated or superheated(011.0 C) vapor.

Test C: SLHX present; evaporatoroutlet quality 0.750.85.

Because of the mixing chambersinstalled both at inlet and outlet ofthe SLHX, the related local pressuredrops contribute to reducing thesuction pressure. The SLHX LPside pressure drop was measuredto range between 12.0 and 17.0kPa. By considering the totalpressure drop (including mixingchambers and SLHX LP pressuredrop) the LP side SLHX outletpressure was 3.6640.016 MPa.The system performance wasinvestigated for GC outlet pressuresfrom 7.8 to 11.0 MPa. The airtemperatures and flow rates at GC

and evaporator inlets, as well as thecooling water temperatures at ICinlet, were adjusted to fulfill thementioned reference operatingconditions for R-744. Since no liquidaccumulator was installed in the

circuit, the carbon dioxide chargewas changed in each test.In Figure 2 the measured COP el values are reported as a function ofthe ratio r p of GC inlet pressure toevaporator outlet pressure, being

2 , ,CO LP out HP out ref el

el el

m h h P COP

P P

(1)

In (1) h LP,out is the R-744 specificenthalpy at evaporator outlet whenno SLHX is installed or at the SLHXoutlet, low pressure side. h HP,out isthe specific enthalpy at the gas-cooler outlet, without SLHX, or atthe SLHX high pressure side outlet.is the carbon dioxide mass flow rateand P el is the electrical power inputto the compressor. R-744 specificenthalpies were calculated with theREFPROP 7.0 code by NIST

(2002).It can be inferred that the optimalGC pressure occurs at slightly lowerrp values in the presence of theSLHX. However, the curves appearflat in correspondence to themaximum COP el. The COPelimprovement is higher than 20%with the SLHX. Furthermore, fromfigure 2, it can be seen that thecooling capacity is almost the samein the three data sets, whereas theR-744 mass flow rate is much lowerwith the SLHX because of thehigher vapor superheat at thecompressor suction.In Cavallini et al. (2005) a code wasdeveloped for the simulation ofdifferent CO2 refrigerating cycles,including the one in figure 4. Inparticular, the compressor wassimulated by considering the actualisentropic efficiency of the two-stages of compression:

, , ,,1

, ,

IC in is LP out is

IC in LP out

h h

h h (2)


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INFO PACK

, , 2

,2

, 2

GC in is suc is

GC in suc

h h

h h

(3)

where h IC,in, h GC,in are R-744 specific enthalpies at IC and GC inletrespectively; h 2 suc is the specificenthalpy in the 2 nd-stage suctionchamber.In Hubacher and Groll (2002 and2003) the overall compressorefficiency is conventionally definedas in eq. (4) and (5):

,,

comp isis TOT

el

P

P (4)

with

2, , ,comp is CO GC in is suct P m h h (5)being h suct the R-744 specificenthalpy at the compressor suction.It is worth noting that in thecompound compressor there is anon-negligible heat transfer from thehigh pressure stage to the lowpressure stage: expressions (2) and(3) assume adiabatic compression;this assumption involves an error,because the real compression workis larger than the denominator in (3)and smaller in (2). Since the ICoutlet temperature (i.e. thetemperature of R-744 entering theelectric motor) was kept constant(tIC,out =33.10.1 C), we might saythat the higher the first stagedischarge temperature, the lower isthe heat transfer from the second tothe first stage. When no SLHX isinstalled the second stagecompression process getssignificant improvement by

transferring a higher heat flowtowards the first stage. This isconfirmed by the results of theproposed compression efficiencies

(eqs. 2 and 3) approach that leadsto values for is,2 roughly 65%higher than the is,1 values whenevaluated at the same pressureratio. The effect of the SLHX is toenhance the vapor superheating at

the compressor suction and thus toincrease the first stage dischargetemperature. The heat flowtransferred from the second stage ismuch lower in this circumstance. Asa consequence, is,1 was found to be around 20 % higher than themeasured values without SLHX(same first stage pressure ratio) and

is,2 around 14% lower than thevalues determined at the samesecond stage pressure ratio without

SLHX. Furthermore the increasedvalue of the R-744 specific volume,due to the higher superheat,contributes to lowering theintermediate pressure (that is thefirst stage discharge pressure) for afixed total pressure ratio (secondstage discharge pressure/first stagesuction pressure). Moreover, sincethe tests were run with a fixed ICoutlet temperature (33 C), the heatflow rate exchanged at the IC isexpected to have a markedinfluence on the compressionefficiency. In figure 3, the IC heatflow is reported for tests A, and B.The main result is that the overallcompressor efficiency,conventionally defined as in eq.(4), gets improved by more than20% in the investigated pressureratios (r p=2.1 2.8). Figure 2indicates that COP el and P ref for testC are almost the same as for test B.This is consistent with the

theoretical analysis proposed byChen and Gu (2005) under thehypothesis of unitary effectivenessof the SLHX. This parameter can be

defined with reference to figure 4 asfollows:

'' 31 31

max max

-- h hh h

h h

(6)

with

'

'

1 33

max

3 11

, - ;min

, -

h t p hh

h t p h

(7)

The measured effectiveness of theSLHX was always higher that 0.95.In this experimental analysis anincrease in the evaporatoreffectiveness was found for test C,as compared to test B. In fact,

partial evaporation in the SLHXpromotes lower vapor quality at theevaporator inlet. In this way, themain part of the surface area of thefinned coil evaporator is interestedby nucleate and convective boiling,whereas the post-dryout processoccurs mainly inside the SLHX. Thelow heat transfer coefficientoccurring in the post-dryout regionwould penalize the evaporatoreffectiveness, but does not affectmarkedly the SLHX, since this isdesigned to operate with singlephase superheated vapor. As amatter of fact, the measuredevaporator effectiveness for test B(with 4C superheating) was 0.57while for test C with 0.8 outletquality was 0.59. This increase ofthe evaporator effectiveness looksmore significant if we consider thatthe larger thermal resistance is airside, and it can be consideredalmost constant for all the reported

tests.

Figure 4. Reference transcritical cycle.

CONCLUSIONS

The system Coefficient ofPerformance was found toincrease up to 20% when an

internal heat exchanger isinstalled in a transcritical CO2

refrigerating circuit operatingwith a two-stage compoundcompressor fitted with anintercooler. The significant

increase in COP is far beyondthat expected when using simple

compression. The reason lies onthe effect of the vapor superheatingat compressor suction; highertemperature enables larger heat


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INFO PACK

flow to be exchanged in IC and,whence valuable decrease incompression work result. In additionto it, another advantage occursduring real operations, i. e. heattransfer coefficient increases at the

evaporator because of delayed dry-out conditions; this cannot be seenin the tests, since the evaporationtemperature was kept constant.

REFERENCES

ASHRAE Standard 116, 1995,Methods of testing for seasonalefficiency of unitary air-conditionersand heat pumps.

Domanski, P.A., Didion D.A. DoyleJ. P. 1994, Evaluation of suction-

line/liquid-line heat exchange in therefrigeration cycle. Int. J.Refrigeration, 17(7): 487-493.

Boewe, D.E., Bullard, C.W. Yin,J.M., Hrnjak, P.S. 2001,Contribution of internal heatexchanger to transcritical R-744

cycle performance, HVAC and RResearch, 7(2), 155-168

Cavallini A., Cecchinato L., CorradiM., Fornasieri E., Zilio C., 2006.Experimental investigation on theeffects of internal heat transfer in atwo-stage transcritical carbondioxide cycle. In: 7th IIR GustavLorentzen Conference on NaturalWorking Fluids. Trondheim,Norway.

Cavallini A., Cecchinato L., CorradiM., Fornasieri E., Zilio C. 2005,Two-stage transcritical carbondioxide cycle optimisation: Atheoretical and experimentalanalysis, Int. J. Refrigeration, 28(8):1274-1283.

Chen Y., Gu J. 2005, The optimumhigh pressure for CO 2 transcriticalrefrigeration systems with internalheat exchangers, Int. J.Refrigeration, 28(8): 1238-1249.

Hubacher B., Groll E. A., HoffingerC. 2002, Performance

Measurements Of A Semi-HermeticCarbon Dioxide Compressor. Proc.of Ninth International Refrigerationand Air Conditioning Conference atPurdue, West Lafayette, USA.

Hubacher B, Groll EA. 2003,Performance measurement of ahermetic, two-stage carbon dioxidecompressor. Proc. XXI IIR Int.Congr. of Refrigeration,Washington, USA.

NIST, 2002, National Institute ofStandard and Technology, RefpropVersion 7.0, Boulder Colorado.

NOMENCLATURE

SLHX suction line heat

exchangerGC gas-cooler

IC intercooler

For more information, please contact: Claudio Zilio ([email protected])

ice_e info pack 3 heat exchangers

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