Cycle Analysis and Turbo Compressor Sizing With Ketone C6F as Working Fluid for Water-Cooled Chiller Applications

Transcription

1 Purdue Univerity Purdue e-pub International Compreor Engineering Conference School of Mechanical Engineering 4 Cycle Analyi and Turbo Compreor Sizing With Ketone a Working Fluid for Water-Cooled Chiller Application Bruce P. Biederman United Technologie Reearch Center Jaro Mulugeta United Technologie Reearch Center Lili Zhang United Technologie Reearch Center Joot J. Braz Carrier Corporation Follow thi and additional work at: Biederman, Bruce P.; Mulugeta, Jaro; Zhang, Lili; and Braz, Joot J., "Cycle Analyi and Turbo Compreor Sizing With Ketone a Working Fluid for Water-Cooled Chiller Application" (4). International Compreor Engineering Conference. Paper Thi document ha been made available through Purdue e-pub, a ervice of the Purdue Univerity Librarie. Pleae contact epub@purdue.edu for additional information. Complete proceeding may be acquired in print and on CD-ROM directly from the Ray W. Herrick Laboratorie at Herrick/Event/orderlit.html

2 C135, Page 1 CYCLE ANALYSIS AND TURBO COMPRESSOR SIZING WITH KETONE AS WORKING FLUID FOR WATER-COOLED CHILLER APPLICATIONS Bruce P. Biederman 1, Jaro Mulugeta 1, Lili Zhang 1, Yu Chen 1, Joot J. Braz 2 1 United Technologie Reearch Center Eat Hartford, CT 618 biederbp@utrc.utc.com mulugej@utrc.utc.com zhangl1@utrc.utc.com cheny@utrc.utc.com 2 Carrier Corporation Syracue, NY joot.j.braz@carrier.utc.com ABSTRACT Ketone, which ha the chemical formula: CF 3 CF 2 C(O)CF(CF 3 ) 2, ha been introduced recently a a zero-odp, zero-gwp fire extinguihing fluid. It non-flammability and non-toxicity combined with it excellent environmental propertie make it an attractive fluid for HVAC application. It low denity limit it ue to centrifugal water-cooled chiller application. The relatively high critical temperature of the fluid promie good thermal cycle efficiency. However, the lope of the aturation dome on the temperatureentropy diagram force the ue of a vapor uction / condened liquid heat exchanger to prevent wet compreion and reduce the throttling loe. The low peed of ound of the fluid allow direct drive ingle tage compreor operation at relatively mall tonnage. The ub-atmopheric evaporator (.162 bar) and condener (.64 bar) preure neceitate the ue of a purge but eliminate the need for preure veel code certification. 1. INTRODUCTION Flammable and/or toxic natural refrigerant are unacceptable a working fluid for large air conditioning and refrigeration ytem due to the rik poed by the ubtantial charge of uch ytem. Effort to reduce the refrigeration charge by applying technologie uch a falling film evaporation reduce the charge level but till leave charge level too large to allow from a afety point of view, ue of thee ytem in the built environment. The only two natural refrigerant left for large HVAC ytem are water and CO Water a refrigerant in chiller From a afety and efficiency point of view water a refrigerant i a very attractive choice. However, the low denity of water at refrigeration temperature level reult in large volumetric flow rate require extremely large machinery with low power denity. The lowet cot compreor olution will be a turbomachine, of either the centrifugal or the axial type. The high onic velocity of water put pecial demand on the trength/weight ratio of the material to be ued in the turbo-machinery. It i therefore quetionable whether a cot effective chiller uing water a refrigerant are poible. 1.2 CO 2 a refrigerant in chiller The very high preure of CO 2 a refrigerant ha jut the oppoite drawback of that of water. It very high denity form a afety rik. A ytem leak could caue aphyxiation. Alo, the tran-critical cycle condition of CO 2 reult in poor ytem efficiencie due to the large throttling loe, which can only be overcome by the ue of power recovery during the expanion proce.

3 C135, Page Are there man-made fluid with natural refrigerant GWP value? The main driver for the ue of natural refrigerant i not the fact that they happen to occur in nature but that thee fluid have a very low direct-effect global warming potential. All commonly ued fluorocarbon refrigerant have global warming potential order of magnitude larger than thoe of the natural refrigerant. Ketone - chemical formula CF 3 CF 2 C(O)CF(CF 3 ) 2 - ha been introduced recently a a zero-odp, GWP1 fire extinguihing fluid [1,2]. It ha alo been conidered a a refrigerant for centrifugal chiller but wa dimied baed on it poor cycle efficiency. Thi low cycle efficiency i omewhat urpriing becaue the fluid ha a fairly high critical temperature ( C in between the critical temperature of HFC245fa C and C), which typically indicate good thermodynamic cycle efficiency. Figure 1 how the COP of the for a number of CFC, HCFC and HFC refrigerant ued on centrifugal chiller a a function of the critical temperature of the refrigerant. The general trend for fluorocarbon-baed refrigerant i an increae in COP with critical temperature. The critical temperature of ketone uggeted a COP in the neighborhood of that of HFC245fa and. However, the cycle calculation reulted in a much lower COP R R123 COP R22 R134a R12 R245fa Tevaporation5 C Tuperheat C Tcondenation35 C T ubcooling C ientropic compreion ienthalpic expanion Critical temperature, C Figure 1. Coefficient of performance (COP) of variou fluorocarbon refrigerant and ketone a a function of critical temperature The reaon for thi deviation can be explained by comparing the TS diagram of HFC245fa, CF6 and. The aturation line for are leaned over heavily. The mall poitive lope of the liquid aturation line reult in large throttling loe. Figure 2 how the TS diagram for, HFC245fa Tcrit184 C Tcrit154 C Tcrit167 C Temperature T, C HFC245fa Normalized entropy, /((T5 C,x1)-(T5 C,h(T35 C,x))) Figure 2. TS diagram with two-phae flow dome of, HFC245fa and on a normalized entropy bai

4 C135, Page 3 and. In order to graphically compare the throttling loe the horizontal entropy axi ha to be normalized baed on the refrigeration effect by dividing the actual entropy by the difference in entropy experienced by the working fluid in the evaporator: normalized ( T 5 C, x 1) ( T 5 C, h( T 35 C, x )) evap, out evap, in The difference in the lope of the aturation line of the T diagram explain the low COP of. Figure 3a, b and c how ide by ide the TS diagram of HFC245fa, and, while maintaining the aturation line of the other two refrigerant. Since the throttle lo i equal to the triangle in the TS diagram, it can be een why the COP of ketone i o much lower than that of HFC123 or HFC245fa. Tcrit184 C Tcrit154 C Tcrit167 C Tcrit184 C Tcrit154 C Tcrit167 C Tcrit184 C Tcrit154 C Tcrit167 C Temperature T, C HFC245fa HFC245fa cycle throttle lo Normalized entropy, /(evap, out - evap,i n) Temperature T, C HFC245f a cycle throttle lo Normalized entropy, /( evap,out- evap,in) Temperature T, C HFC245f a cycle throttle lo Normalized entropy, /( evap,out - evap,in ) Figure 3a. TS diagram of HFC245fa. Figure 3b. TS diagram of. Figure 3c. TS diagram of Replacing the throttle valve in the cycle with a two-phae flow expanion turbine i one way of recovering the throttle loe. However, thi method add ubtantial complexity and cot to the ytem due to the additional rotating equipment and it control. 3. IMPROVING THE REFRIGERATION CYCLE WITH A SUCTION LINE HEAT EXCHANGER When the TS diagram ha o much lean to it that the vapor aturation line become poitive, ientropic compreion reult in wet compreor exit condition. Under thoe circumtance the refrigeration cycle doe not uffer from a de-uperheating lo and throttling i the only fundamental irreveribility of the cycle. Slightly wet ientropic compreor exit condition do not reult in compreor eroion ince actual compreor inefficiencie will caue additional heating of the working fluid reulting in dry compreor exit condition. Both and HFC245fa compre wet ientropically but exhibit uperheated compreor dicharge condition at aerodynamic efficiencie in the high ninetie, much higher than what i technically achievable. compreor exit condition are till wet at 35% compreor aerodynamic efficiency. Since actual compreor efficiencie will be higher, compreor eroion become a concern for. A uction line heat exchanger ha to be ued to uperheat the vapor before entering the compreor to prevent compreor eroion. When the heat required for uperheating the vapor leaving the evaporator i extracted from the liquid leaving the condener a hown in Figure 4b, a boot overall cycle efficiency i poible. Sub-cooling the liquid refrigerant leaving the condener reduce the throttling lo of the cycle. Heat Out Heat Out T Condener COND 35 C Throttle Valve Evaporator/Cooler T EV AP 5 C Heat In Compreor Motor Power In T COND 35 C Throttle Valve Condener Suction Line HX Evaporator/Cooler Compreor Motor Power In Figure 4a. Simple vapor compreion cycle 5 T EVAP 5 C Heat In 6 Figure 4b. Vapor compreion cycle with uction line heat exchanger

5 C135, Page 4 The uction line heat exchanger introduce a thermodynamic irreveribility, ince tranferring heat from hot liquid to cold vapor require a temperature difference T a driving force, which i a thermodynamic lo. With increaed uction uperheat from the uction line heat exchanger thi thermodynamic irreveribility i reduced. Eventually another thermodynamic cycle irreveribility, the condener de-uperheat lo will arie. A mall uction-line heat exchanger with a large log-mean-temperature difference (LMTD), although good from a cot point of view, can actually woren the ytem COP. Such i the cae for. It reduction in throttling lo irreveribility i initially more than offet by the increaed irreveribility in the uction line heat exchanger. For larger heat exchanger with maller LMTD the proce revere and an overall ytem COP improvement can be achieved. Thi trend i hown in Figure 5. A uction line heat exchanger with an effectivene of 8% correponding to an LMTD of 8 C, reult in a cycle efficiency lightly better than that of HFC245fa COP Log mean temperature difference of uction line heat exchange, C Figure 5. COP of the ideal refrigeration cycle with a uction line heat exchanger a a function of heat exchanger log mean temperature difference. The uction line heat exchanger add cot to the ytem but ince it alo increae the ytem capacity per unit volume paing though the compreor, it net effect in term of increaed ytem cot for a given cooling capacity i moderate. The uction line heat exchanger bring the cycle efficiency of at an attractive level. Thi combined with it negligible GWP make an intriguing refrigerant for future chiller. Obviouly, the uction line heat exchanger ha to be deigned with minimal vapor ide preure drop ince thi preure drop will increae the required compreor head and therefore it correponding power requirement. 8.8 COP R22 R134a R12 R245fa R123 with uction line HX R Tevaporation5 C Tuperheat C Tcondenation35 C T ubcooling C 7.4 ientropic compreion ienthalpic expanion Critical temperature, C Figure 6. COP a a function of critical temperature howing that matche the traditional correlation when equipped with a uction line heat exchanger

6 imple cycle Throttle lo imple cycle with uction line HX 2 K ubcooling form uction line HX COP throttloing lo irreveribility: x % uction line HX irreveribility deuperheat horn irreveribility uction line HX LMTD Capacity increae Throttle lo C135, Page 5 Figure 7 illutrate how for the improvement in cycle efficiency i obtained. The large throttle lo of the imple refrigeration cycle i reduced dramatically and being offet by a much maller uction line heat exchange lo and a minicule de-uperheat lo. A a reult the cycle efficiency ha improved from a COP of 7.34 to 8.2. Temperature, C Entropy. kj/kg C Temperature, C COP idea l 7.34 COP idea l 8.2 De-uperheat lo Suction line HX lo LMTD7.2 C Entropy. kj/kg C Figure 7. Temperature entropy diagram with loe of in the imple cycle and the modified cycle with a uction line heat exchanger 4. CENTRIFUGAL COMPRESSOR DESIGN FOR In thi ection we will ize a ingle tage centrifugal compreor for ketone, the non-flammable natural refrigerant water and CO 2 a well a the commonly ued refrigerant and HFC134a. For each of thee fluid a compreor will be deigned for the ame pecific peed and diameter. The concept of pecific peed come from imilarity conideration and ha been promoted by Balje, Compreor rotational peed, volumetric flow rate and head (or ientropic enthalpy rie) van be combined in a non-dimenional group. Ientropic enthalpy rie ha the dimenion of J/kg or m 2 / 2 while volumetric flow rate ha the dimenion of m 3 /. Rotational peed, which ha a dimenion of -1, can be obtained taking the ratio of the head to the power ¾ and the flow to the power ½: 3 / 4 2 m 3 / 2 m [ H ] 3 / / 2 1 [ N] (1) 1/ 2 3 / 2 [ Q] 3 1/ 2 m m 1/ 2 Following thi approach a non-dimenional pecific peed n can be defined a N Q n (2).75 H Similarly, a non-dimenional pecific diameter d can be defined a:. 25 DH d (3) Q The concept of pecific peed and diameter i ueful for preliminary izing of turbo machinery equipment. Optimum efficiency of centrifugal compreor i obtained at a pecific peed between.6 and.9 and a pecific diameter between 3 and 5 a hown in Figure 8. We will ize ingle tage centrifugal compreor for a pecific peed on.76 and a pecific diameter of 3.4, reulting in the following equation defining compreor peed N and diameter D:.75 H Q N.76 (4a) D 3.4 (4b).25 Q H

7 C135, Page 6 pecific peed/diameter election point Figure 8. Selection of pecific peed n and diameter d on the Balje diagram for centrifugal compreor, reproduced with permiion from [3]. The compreor izing will be done for a refrigeration duty of ton (176 kw) auming identical evaporation and condenation temperature of 5 and 35 C, repectively. Cycle analyi will then give the value for head and volumetric flow rate. Uing equation 4a and 4b will then give the peed and impeller diameter of the centrifugal compreor for each of thee working fluid. The reult of thee calculation are ummarized in Table 1. water CO 2 HFC134a ODP.2 GWP h evap kj/kg h,comp kj/kg Head m mdot kg/ ρ kg/m 3 Vdot m 3 / N rpm D m u m/ a m/ u 2 /a Pr COP Table 1. Calculation of impeller peed and diameter of a ton (176 kw) centrifugal compreor for ketone and other working fluid Thee calculation how that with a uction line heat exchanger reult in a low rpm N, a low impeller tip peed u 2 and a large impeller diameter D compreor deign with a COP approaching that of. The low ientropic enthalpy rie of the compreor correpond to a low head and reult in a low tip peed

8 water C135, Page 7 u 2. Combined with the low denity of comparable to, but lower than that of it reult in a compreor rotational peed cloe to /6 Hz allowing ingle tage direct drive operation at the appropriate pecific peed election. To illutrate the relative ize of a compreor Figure 9 how it ton ingle tage centrifugal compreor for three low GWP refrigerant ize in relatio n centrifugal compreor with the natural refrigerant water and CO 2 while Figure 1 make a compreor ize comparion with the two currently mot popular centrifugal chiller refrigerant and HFC134a.. ketone CO 2 Impeller diamet er: m.991 m.85 m Impeller rotational peed: 8,966 rpm 2,587 rpm 51,177 rpm Impeller tip peed: 711 m/ 134 m/ 228 m/ ton ingle tage ketone compreor relative to and HFC134a compreor Figure 9. Illutration of relative, ize and peed of a ton centrifugal compreor with ketone, water and CO 2 a refrigerant ketone HFC134a Impeller diamet er:.991 m.68 m.23 m Impeller rotational peed: 2,587 rpm 5,423 rpm 14,497 rpm Impeller tip peed: 134 m/ 173 m/ 18 m/ Figure 1. Illutration of relative, ize and peed of a ton centrifugal compreor with ketone, and HFC134a a refrigerant

9 C135, Page 8 5. CONCLUSIONS 1. Ketone hare the environmental advantage of natural refrigerant (zero ODP and negligible GWP) with the engineering advantage of man-made fluid (reaonable power denity and nonflammable and non-toxic). 2. The low cycle efficiency of ketone in the imple vapor compreion cycle i dramatically improved by the ue of a uction line heat exchanger, reulting in COP value in line with expectation baed on it critical temperature. 3. The low head requirement for ketone combined with it large volumetric flow rate requirement enable gearle ingle-tage direct-drive /6 Hz compreor operation for compreor erving chiller in the to ton capacity range. 4. The low impeller tip peed allow the ue of alternate material for impeller fabrication 5. The ub-atmopheric evaporator (.162 bar) and condener (.64 bar) preure neceitate the ue of a purge but eliminate the need for preure veel code certification. REFERENCES 1. Owen, J.G., Phyical and Environmental Propertie of a Next Generation Extinguihing Agent, NIST SP 984; June 2, Halon Option Technical Working Conference, 12th. Proceeding. HOTWC 2. April 3-May 2, 2, Albuquerque, NM 2. Owen, J.G., Undertanding the Stability and Environmental Characteritic of a Sutainable Halon Alternative, NIST pecial publication 984-1, 3 3. Balje, E.O., Turbomachine, A Guide to Deign, Selection and Theory, John Wiley and Son, New York, 1981.