Development of a refrigerant to refrigerant heat exchanger for high efficiency CO2 refrigerant cycle

Transcription

1 Purdue Unversty Purdue e-pubs Internatonal Refrgeraton and Ar Condtonng Conference School of Mechancal Engneerng 206 Development of a refrgerant to refrgerant heat exchanger for hgh effcency CO2 refrgerant cycle Ryuhe Kaj Dakn Industres, LD, Japan, ryuuhe.kaj@dakn.co.jp Shun Yoshoka Dakn Industres, LD, Japan, shun.yoshoka@dakn.co.jp Hrokazu Fujno Dakn Industres, LD, Japan, hrokazu.fujno@dakn.co.jp Follow ths and addtonal works at: Kaj, Ryuhe; Yoshoka, Shun; and Fujno, Hrokazu, "Development of a refrgerant to refrgerant heat exchanger for hgh effcency CO2 refrgerant cycle" (206). Internatonal Refrgeraton and Ar Condtonng Conference. Paper hs document has been made avalable through Purdue e-pubs, a servce of the Purdue Unversty Lbrares. Please contact epubs@purdue.edu for addtonal nformaton. Complete proceedngs may be acqured n prnt and on CD-ROM drectly from the Ray W. Herrck Laboratores at Herrck/Events/orderlt.html

2 2202, Page Development of a refrgerant to refrgerant heat exchanger for hgh effcency CO 2 refrgerant cycle Ryuhe KAJI *, Shun YOSHIOKA, Hrokazu FUJINO Dakn Industres LD., - Nsh-Htotsuya, Settsu, Osaka, , Japan Contact Informaton (E-mal: ryuuhe.kaj@dakn.co.jp) * Correspondng Author ABSRAC he cycle performance of CO 2 refrgerant s worse than that of HFCs snce the workng pressure s hgher whch suffers larger compresson and expanson losses. o enhance the cycle performance smlar extent to HFCs, refrgerant to refrgerant heat exchangers need to be appled as one of the many components equpped nto the CO 2 cycle. However, when such a heat exchanger s used as an economzer, whch works between hgh temperature refrgerant after the gas cooler and low temperature refrgerant by reducng to the ntermedate pressure, t s confrmed from the vsualzaton that the refrgerant n low temperature sde s observed as a vapor-lqud dspersed flow. In such a condton, the evaporatng performance s low snce most lqud flows n the vapor core as droplets. In ths study, amng at weght savng and compactness for a refrgerant to refrgerant heat exchanger as an economzer, a newly desgned alumnum double-tube heat exchanger n whch outer channel s formed by multports nstead of annulus s ntroduced. Spral fns are appled n the surface of nner channel to am for attachng lqud flowng n vapor core to the nner tube wall. he local heat transfer coeffcent was expermentally measured n a sngle tube and the performance of a heat exchanger assembly was estmated. hese results show that ths newly desgned heat exchanger s effectve to enhance heat transfer performance.. INRODUCION Improvement of energy effcency and reducton of envronmental load for ar-condtoners have been requred n recent years. Especally, HFC refrgerants wdely used n ar-condtoners have hgh Global Warmng Potental (GWP): therefore, ther reducton has been targeted by the nternatonal agreements. In terms of reducng envronmental mpact, R32 s consdered the best balanced refrgerant accordng to the paper by ara and Hakawa (204) and globally spreadng recently. For refrgerant canddates to have potental further reducng GWP of ar-condtonng systems, CO 2 refrgerant, whch s wdely used n water heat pump, refrgeraton and freezer, s consdered. Whle the refrgerant propertes show envronmentally pros, cons le n cycle performance: the performance becomes lower than conventonal HFCs due to the ncrease of dscharge temperature n compresson process and throttle loss resultng from hgh pressure dfference n expanson process. In order to compensate these losses and acheve smlar effcency wth conventonal HFCs, the refrgerant cycle wth nternal heat exchangers s necessary. In ths study, newly desgned alumnum double-tube heat exchanger s appled to the one of the nternal heat exchangers, an economzer, and the effect s verfed by measurng heat transfer coeffcent n tube and estmatng the performance of heat exchanger assembly n the CO 2 refrgerant cycle. 2. PRACICAL CONDIIONS AND CONFIGURAIONS OF HEA EXCHANGER Fgure shows the condton of the economzer and refrgerant temperature dstrbuton used n the CO 2 refrgerant cycle. Refrgerant n hgh temperature sde s supercrtcal and that n low temperature sde experences two-phase

3 2202, Page 2 to superheated vapor. As seeng temperature dfference, temperature gradually decreases n hgh temperature sde snce the refrgerant s supercrtcal, whle temperature stays constant and then ncreases n low temperature sde snce the nflow refrgerant s two-phase and outflow s vapor. When hot refrgerant flows n outer channel and cold refrgerant flows n nner channel, flow pattern of cold refrgerant vares dependng on the nlet pressure as shown n Fgure 2. Whle the lqud flm flows along the tube wall and vapor and droplets flow n the tube core n the low saturaton temperature of the evaporator condton, thnner flm flows and more droplets are dspersed n the tube core s observed n the relatvely hgh saturaton temperature of the economzer condton. hs s manly caused by the effect of the densty rato between vapor and lqud. he actual ratos are 7: n evaporator (Evaporaton temperature, e = 7 o C) and 3: n the economzer condton ( e = 25 o C). hese ratos are much lower than the conventonal HFCs. For example, the ratos of R4A are 6: and 29:, respectvely. Such dspersed flow s sgnfcantly neffectve to heat transfer snce workng lqud s absent on the tube wall. In order to enhance the heat transfer, nner surface geometry havng two-type of fns was ntroduced. he geometry of newly desgned alumnum double-tube heat exchanger s shown n Fgure 3. Large fns were desgned to catch the droplets flow n the core and small fns were to keep the lqud flm attached to the tube wall. Detal specfcatons of the heat transfer tube are tabulated n able. he tube s twsted to settng helx angle, fve degree, after extrudng the alumnum to be specfed cross sectonal shape. he thcknesses of the tube wall a, b, c were decded by mplementng elasto-plastc analyss to keep hgh pressure resstng strength Hgh temperature sde Pressure [MPa] Low temperature sde Enthalpy [kj/kg] Fgure : Condton and temperature dstrbuton of an economzer n supercrtcal CO 2 cycle (a) (b) Fgure 2: Flow vsualzaton n low temperature sde ((a) P = 4.8MPa( e = 7ºC), G = 00kg/m 2 s, x = 0.5, d t = 4.6mm, (b) P = 6.43MPa( e = 25ºC), G = 00kg/m 2 s, x = 0.5, d t = 4.6mm)

4 2202, Page 3 Fgure 3: Geometry of newly desgned alumnum double-tube heat exchanger able : Specfcaton of newly desgned alumnum double-tube heat exchanger Outer dameter d o [mm] 8.0 Number of outer hole N o 2 Inner base dameter d [mm] 5.8 Equvalent outer dameter per 2.5 hole d e [mm] Fn heght (hgh) H fh [mm] 0.45 Outer thckness a [mm].7 Fn heght (low) H fl [mm] 0.5 Beam thckenss b [mm].3 Number of fn N f 40 Bottom thckness c [mm] 2. Helx angle γ [ º] 5 Surface area enlargement rato (nner sde) [%] 3 3. EXPERIMENAL SEUP Fgure 4 shows the schematc dagram of the expermental setup. he refrgerant supply system conssts of a compressor, a corols mass flow meter, heat exchangers, an expanson valve to control the nlet refrgerant condton to a heat transfer tube such as pressure, temperature, mass flow rate and qualty etc. In addton, a separated ol crculaton lne s located to enable provdng lubrcant ol wthn refrgerant flow. PAG (Poly alkyl glycol) was used n ths study. he ol crculaton rate (OCR) was measured by capturng refrgerant and ol mxture nto a samplng tank set n the refrgerant crcut and weghed refrgerant and ol masses, respectvely, accordng to ASHRAE standard (996). Water flows nto the heat exchanger counter-currently to refrgerant flow n order to supply heat source and crculates through a tank by usng a pump. Water tank exposed to atmosphere s equpped wth heatng and chllng unt to keep a constant temperature. Platnum resstance thermometers were used to measure the refrgerant and water temperature. A pressure transducer and a dfferental pressure transducer were used to measure the pressure of refrgerant-sde. Fgure 5 shows the heat transfer tube tested n ths experment. he effectve length of heat transfer s 6mm. wo types of heat transfer tube confguraton were tested: One s specfed n Fgure 3, the other s the conventonal tube havng smooth nner surface wthout twst. he heat transfer tube s set horzontally and thermally nsulated durng the experment. he expermental condtons are lsted n able 2. he heat flux s calculated based on nner base tube dameter. o examne the effect of nner channel, the heat transfer coeffcent of nner channel h s derved as followng equatons: h ( 0.5d + c) ln[ ( 0.5d c) / 0.5d ] A + = K ho Ao λal () where overall heat transfer coeffcent K s calculated from heat capacty of water sde Q w :

5 K = Qw A lm 2202, Page 4 (2) In Equaton (), the unknown heat transfer coeffcent of outer mult-port channels h o can be expermentally obtaned by mplementng water-to-water experment. At a constant water flow n outer channels, water flowng nner channel vares n several ponts. When the results plot wth the nverse of obtaned overall heat resstance as vertcal axs and the nverse of nner channel velocty as horzontal axs, the ntercept ndcates that the heat resstance of nner channel s neglgble; therefore, the heat transfer coeffcent of outer channel h o can be derved. hs technque s known as Wlson plot method (reference n Shah (985)). Water Pump Refrgerant Supply System P Heat ransfer ube P Water ank Sub-Evaporator Compressor Accumulator Gas Cooler Mass Flow Meter Samplng secton for OCR P Expanson Valve Ol Separator Ol Separator Ol Pump Mass Flow Meter Water Refrgerant Ol : emperature (Pt0 sensor) P : Pressure (Pressure gauge) P : Pressure drop (Dfferental pressure gauge) Fgure 4: Expermental faclty Water CO 2 Heat flux Q [kw/m 2 ] L = 6 Fgure 5: Heat transfer tube tested able 2: Expermental condtons Mass flux G [kg/m 2 s] Evaporaton temperature e [ C] Qualty [-] , , 0.5 OCR [weght %]

6 2202, Page 5 4. RESULS AND DISCUSSION Fgure 6 shows the varaton of evaporatve heat transfer coeffcent aganst qualty. In the fgure, the tube confguraton specfed n Fgure 3 s referred to as Spral, whle the conventonal tube wthout twst and nner fns s referred to as Smooth. It s found that the case of spral tube wthout ol gves the hghest heat transfer coeffcent. he results show that, mostly, the heat transfer coeffcent wthout ol gves hgher than that wth ol. he same trend was reported by Yoshoka et al (2008a, 2008b). hey assume that the presence of ol n the tube wall wll affect the heat transfer coeffcent snce the PAG ol s mmscble to CO 2 refrgerant and nvestgated the effect of ol retenton rate on heat transfer coeffcent. Exceptons are n the mddle qualty n hgh flow rate and hgh qualty regons where the results are nverted. In these regons, sgnfcant decrease of heat transfer ntates as ndcated wth dashed lnes n the fgure. hs seems to be caused by the flm comes off the nner wall to be dspersed. By addng ol, ths phenomenon, referred to as dryout n the fgure, s mtgated and delayed untl hgher qualty. he effect on the heat transfer coeffcent s more sgnfcant n hgher flow rate snce the sgnfcant decrease occurs n lower qualty. For smooth tube, comparng the effect of mass flow rate of pure CO 2, the heat transfer coeffcent n hgh flow rate s larger than that n low flow rate n low qualty. But ths trend s nverted n the mddle qualty. It s thought that when the flow rate becomes hgh, the lqud tends to flow n the tube core as droplets: therefore, thnner lqud flm on the tube wall causes nsuffcent lqud supply n evaporatng nner surface. When OCR = 0.5%, whch occurs n the real stuaton n the refrgerant cycle, the heat transfer coeffcent of spral tube s hgher than smooth tube n hgh flow rate. In low flow rate, although the heat transfer coeffcent s smlar up to mddle qualty, overall qualty s hgher than smooth tube snce the sgnfcant decrease of heat transfer coeffcent mtgated n hgh qualty. For example, n the specfc condton where qualty equals 0.8, the heat transfer coeffcent ncreases by 58% and 45% when G = 980kg/m 2 s and G = 500kg/m 2 s, respectvely. As consderng the surface enlargement rato.3, the heat transfer enhancement by twst and fns s 20% and 87%, respectvely. Fgure 7 shows the varaton of refrgerant pressure drop aganst qualty. From the fgure the pressure drop ncreases as follows: smooth wthout ol < smooth wth ol < spral wthout ol < spral wth ol n order. he results can be well predcted by the homogeneous model based on Blasus s equaton. Above results can be predcted by.0,.6, 2.6 and 4.6 tmes of homogeneous model, respectvely. As the temperature decreases by pressure drop s kpa corresponds to 0.06 ºC, the effect of refrgerant pressure drop of nner tube s mnor n nstallng the heat exchanger nto the refrgerant cycle. Heat transfer coeffcenth [kw/m 2 K] 0 Smooth, OCR=0 Smooth, OCR=0.5 e=25[ºc], q=40[kw/m 2 ], G=980[kg/m 2 s] Intaton of dryout Qualty x [-] Spral, OCR=0 Spral, OCR=0.5 Heat transfer coeffcenth [kw/m 2 K] 0 Smooth, OCR=0 Smooth, OCR=0.5 e=25[ºc], q=40[kw/m 2 ], G=500[kg/m 2 s] Intaton of dryout Qualty x [-] Fgure 6: Varaton of evaporatve heat transfer coeffcent wth qualty Spral, OCR=0 Spral, OCR=0.5

7 2202, Page 6 0 Smooth, OCR=0 Smooth, OCR=0.5 Homogeneous Homogeneous x.6 Spral, OCR=0 Spral, OCR=0.5 Homogeneous x2.6 Homogeneous x4.0 e=25[ºc], q=40[kw/m 2 ], G=980[kg/m 2 s] 0 Smooth, OCR=0 Smooth, OCR=0.5 Homogeneous Homogeneous x.6 Spral, OCR=0 Spral, OCR=0.5 Homogeneous x2.6 Homogeneous x4.0 e=25[ºc], q=40[kw/m 2 ], G=500[kg/m 2 s] Pressure drop P [kpa/m] Pressure drop P [kpa/m] Qualty x [-] Qualty x [-] Fgure 7: Varaton of refrgerant pressure drop wth qualty From the above, t s confrmed that the newly desgned heat transfer tube wth fns and twst s effectve to enhance the heat transfer by not only ncreasng nner surface area but mtgatng the dryout, sgnfcant decrease of heat transfer, to control the evaporatve lqud located to the tube wall surface. It also confrmed that the heat transfer enhances n the condton wth ncludng ol, and the hgher refrgerant flow s, the more effectve heat transfer can be obtaned. 5. VARIFICAION WIH HEA EXCHANGER ASSEMBLY IN HE REFRIGERAN CYCLE From the above results, a heat exchanger nstalled n the CO 2 refrgerant cycle was desgned as shown n Fgure 8 and the effect of newly desgned heat exchanger s examned. he estmaton was mplemented at the economzer condton of rated coolng mode, lsted n able 3. In order to obtan the same heat capacty, the lengths of the two heat exchangers tested n ths study were estmated to understand the effect of newly desgned heat exchanger based on the expermental results obtaned n prevous secton. For estmaton, the length of heat exchanger s dvded nto 50 segments and the heat capacty of each segment s calculated by usng two-phase heat transfer coeffcent obtaned n prevous secton and the heat transfer coeffcent of vapor phase and hgh temperature sde whch are used the equaton of sngle phase flow. he result s shown n Fgure9. From the fgure, t s found that the new type heat exchanger can reduce the sze to 23%. Fgure 8: Heat exchanger confguraton for nstallng the refrgerant cycle

8 2202, Page 7 able 3: he condton of the economzer n the refrgerant cycle (Rated coolng mode) Hgh temperature sde Inlet pressure Inlet temperature Mass flow rate Low temperature sde Saturaton temperature Mass flow rate Heat capacty 9.36 [MPa] 38.0 [ºC] 294 [kg/hr] 30.0[ºC] 97 [kg/hr] 2.58 [kw] 7 Length of Heat exchanger[m] % 0 Fgure 9: he effect of the newly desgned heat exchanger wth the same capacty estmated 6. CONCLUSIONS For a refrgerant-to-refrgerant heat exchanger used as an economzer n the CO 2 refrgerant cycle for ar-condtoner, a newly desgned alumnum double-tube heat exchanger n whch outer channel s formed surroundng mult-ports nstead of outsde annulus s proposed to amng at weght savng and compactness. It can be confrmed that the proposed nner surface geometry enhances the heat transfer and the effect s verfed expermentally by measurng heat transfer tube and estmatng the performance of the heat exchanger assembly n the refrgerant cycle. NOMENCLAURE A surface area (m 2 ) a outer thckness (mm) b beam thckness (mm) c bottom thckness (mm) d dameter (mm) G mass flux (kg/m 2 s) h heat transfer coeffcent (W/m 2 K) K heat transfer rate (W/m 2 K) L length (mm) N number (-) OCR ol crculaton rato (w%) Q heat capacty (W) q heat flux (W/m 2 )

9 2202, Page 8 e evaporaton temperature ( o C) P pressure drop per length (kpa/m) lm log mean temperature (K) γ helx angle ( o ) λ thermal conductvty (W/mK) Subscrpt Al f fh f o t w alumnum fn hgh fn low fn nner outer tube water REFERENCES ANSI/ASHRAE Standard (RA 2006), (2006). Standard method for measurement of proporton of lubrcant n lqud refrgerant. Amercan Socety of Heatng, Refrgeratng, and Ar-Condtonng Engneers, Inc., Atlanta. Shah, R.K. (985). Heat exchangers. In W.M. Rohsenow, J.P. Hartnett & E.N. Ganc (Eds.), Handbook of Heat ransfer Applcatons (4-2). New York, NY: McGraw-Hll. ara, S. & Hakawa,. (204). Evaluaton of performance of Heat pump system usng R32 and HFO mxed refrgerant. Internatonal Refrgeraton and Ar Condtonng Conference, Purdue Unversty, West Lafayette. IN: paper Yoshoka, S., Km, H. & Kasa, K. (2008a). Effect of PAG ol crculaton rate on the heat transfer performance of ar cooled heat exchanger n carbon doxde heat pump system. Proc. IEA Heat Pump Conference, Zurch: paper 3.. Yoshoka, S., Km, H. & Kasa, K. (2008b). Heat transfer performance and ol behavor for R744 wth PAG ol n arcooled heat exchanger. Proc. 8th IIR Gustav Lorentzen Conference on Natural Workng Fluds, Copenhagen: paper M4-04. ACKNOWLEDGEMEN he work presented n ths manuscrpt was supported by the New Energy and Industral echnology Development Organzaton (NEDO), JAPAN.