外文原文-耦合冷库干燥冷却除湿系统的空气循环制冷的性能分析.pdf

Performance analysis of air cycle refrigerator integrated desiccant system for cooling and dehumidifying warehouse S S Elsayeda T Miyazakia Y Hamamotob A Akisawaa T Kashiwagia aGraduate School of Bio Applications and Systems Engineering Tokyo University of Agriculture and Technology Naka cho 2 24 16 Koganei shi Tokyo 184 8588 Japan bDepartment of Mechanical Engineering Science Kyushu University 6 10 1 Hakozaki Higashi Ku Fukuoka 812 8581 Japan a r t i c l e i n f o Article history Received 12 January 2007 Received in revised form 29 June 2007 Accepted 9 July 2007 Published online 15 July 2007 Keywords Compression system Modelling Simulation Thermodynamic cycle Air Cold store Desiccant Dehumidifi cation Performance a b s t r a c t In this paper the performance of air cycle refrigerator integrated desiccant system used to cool and dehumidify warehouse is analyzed theoretically Simulation analysis is carried out to calculate the system coeffi cient of performance cooling effects and the humidity change under different values of pressure ratio storage zone temperature inside dock and outdoor air conditions Also the effect of the air cycle and the rotor parameters on the system performance is evaluated From the simulation result it is found that the des iccant system has the ability to supply air to the dock area at very low humidity The sys tem coeffi cient of performance increases due to the exhaust heat recovery on the desiccant system and this enhancement can be more than 100 The coeffi cient of performance of the proposed system is greater than that of a conventional system under the same operat ing conditions 2007 Elsevier Ltd and IIR All rights reserved Analyse de la performance d un re frige rateur a cycle d air muni d un syste me a de shydratant utilise pour le refroidissement et la de shumidifi cation d un entrepo t Mots cle s Syste me a compression Mode lisation Simulation Cycle thermodynamique Air Entrepo t frigorifi que De shydratateur De shumidifi cation Performance Corresponding author E mail address elsayed elsaid S S Elsayed www iifi ir org available at journal homepage 0140 7007 see front matter 2007 Elsevier Ltd and IIR All rights reserved doi 10 1016 j ijrefrig 2007 07 002 international journal of refrigeration 31 2008 189 196 1 Introduction Oneofthelargestitemsin thebudgetofa warehouseisthe ex pense of workers compensation Every year thousands of warehouse workers slip and fall on the ice and water which condenses on the cold fl oors of the loading docks in front of frozen food areas Applications Engineering Manual 1999 Loading docks are maintained at low temperature to avoid loading the warehouse refrigeration system with warm moist air pulled through doors air leaking into the dock area will carry enough moist to condense on the fl oor Also the fl oor near the very low temperature warehouse itself will freeze and ice will form on racks near the door inside the storage area Moisture also increases the load on the primary ware house refrigeration system which operates at very low suc tion temperature so this system is not effi cient Once the frost has formed it must be removed through defrost cycles which consume more energy Using desiccants to remove moisture reduces the load on the primary refrigeration system Also in most cases install ing a dedicated dehumidifi cation system saves energy and can reduce the cost of solving the safety problem Applica tions Engineering Manual 1999 Air cycle refrigerator has a good performance at low tem perature applications Nikai et al 1998 and rejects heat at high temperature So the air cycle refrigerator has a potential to be a good system to cool and dehumidify the warehouse where the cooling effect of the cycle is used as a part of the warehouse refrigeration load and the exhausted heat from the compressor regenerates the desiccant system An air cycle refrigerator utilizes air as refrigerant The system energy effi ciency is lower than that of a vapor com pression type refrigerating machine having phase change pro cesses Therefore the air cycle refrigerator was only used in a fi eld that was restricted for refrigerating capacity and tem perature condition Nikai et al 1998 Katoh et al 1998 Spence et al 2004 2005 The system however exhausts relatively high temperature heats It was numerically clarifi ed that if the exhausted heat would be utilized effectively the system performance would increase Nikai 2001 Recently desiccantcoolingsystemsreceiveconsiderableat tention as an effective technology in utilizing waste heat Such systems can be operated with waste heat source temperature below150 C Reviewsonthedesiccantsystemshavebeenper formed Waugman et al 1993 Novosel 1996 Hamamoto et al 2004 and vast amount of works such as performance analysis and experimental study of dehumidifi cation rotor Jurinak and Mitchell 1984 Kodama et al 2001 have been done The performance of air cycle refrigerator combined with desiccant rotor was studied in air conditioning application Hamamoto et al 2002a Elsayed et al 2006 It has already shown a good enhancement in the performance of air cycle In a research project the performance of air cycle refrigerator combined with desiccant rotor used in refrigerating and dehu midifying warehouse is investigated experimentally NEDO 2003 2004 The result of this study shows that compared Nomenclature COP total coeffi cient of performance COPV coeffi cient of performance of vapor compression cycle Cpbd specifi c heat of dry adsorbent bed kJkg 1K 1 Drotdiameter of a desiccant rotor m henthalpy of air kJkg 1 hfglatent heat of water kJkg 1 kb equivalent mass transfer coeffi cient m s Llength of desiccant rotor m Ppressure Pa qhheat of adsorption kJkg 1 RHrelative humidity of air RPpressure ratio of air cycle RP P3 P6 P5 PS Ttemperature C uvelocity m s V air fl ow rate kgs 1 xhumidity g kg hC isentropic effi ciency of compressor he heat to electricity converting effi ciency hEeffectiveness of sensible heat exchangers hg gas heater effi ciency hT isentropic effi ciency of turbine g specifi c heat ratio of air rdensity kgm 3 qangle of rotor Subscripts aair adadsorption badsorbent bed cinside the warehouse dedesorption ininside the warehouse dock ooutdoor ssupply to the warehouse sssupply to the warehouse in the case of conven tional system vvapor compression cycle CT M C 6 1 in 9 2 3 O 4 5 S 8 2 5 3 4 1 Desiccant rotorin Dock inside air O Outdoor air 2 Compressor 3 Expander 4 5 Heat exchangers S Supply air to cooled storage C Return air from cooled storage 2 Dehumidified air to dock area Storage area Dock area Fig 1 Desiccant air cycle international journal of refrigeration 31 2008 189 196190 with conventional system this system saves energy by 20 and reduces running cost by 30 In this study the performance of the air cycle refrigerator driving desiccant air conditioning system to cool and dehu midify the warehouse is studied by using numerical simula tion The system coeffi cient of performance and the amount of humidity removed from the dock area are calculated at different values of pressure ratio storage zone temperature inside dock and outdoor air conditions Also the effect of the air cycle and the rotor parameters on the system perfor mance is evaluated Finally the system coeffi cient of perfor mance is compared with those of conventional system composed from vapor compression cycle using R22 and des iccant rotor 2 Performance evaluation 2 1 System composition The structure of the air cycle refrigerator driving desiccant system is shown in Fig 1 The system is composed of air cycle refrigerator and desiccant rotor Waste heat of high tempera ture from the air cycle refrigerator is utilized for regenerating a desiccant rotor 2 2 Analysis model and method 2 2 1 Air cycle refrigerator The air temperature of each point in the air cycle is calculated by assuming the isentropic effi ciencies of compression and expansion processes and the effectiveness of heat exchangers as follows T03 T6 RP g 1 g 1 hC T0 3 T6 T 3 T6 2 T0s T5 1 RP g 1 g 3 hT T5 Ts T5 T0s 4 hE T3 T4 T3 To 5 hE T4 T5 T4 Tc 6 where T03and T0sare the theoretical outlet temperatures from the compressor and the turbine respectively for isentropic compression and expansion 2 2 2 Rotor performance The rotorperformanceis evaluatedaccording to the analytical model Hamamoto et al 2002b 2 2 3 Total and system coeffi cients of performance The coeffi cient of performance of the system and cooling effects are calculated as follows Sensible cooling effect Vde hc hs 7 Latent cooling effect Vadhfg xin x2 8 COP Vde hc hs Vadhfg xin x2 Vde h3 h6 h5 hs 9 2 3 Calculation conditions The baseline input data on the calculation listed in Table 1 areadsorptionanddesorptionairvelocities isentropic Table 1 Baseline input data Tin C 5 0he0 38 To C 33 0hg0 8 Tc C 40 0uad m s 2 Xin g kg 3 2ude m s 1 Xo g kg 16N rph 11 0 RP2 0hE0 85 hT0 86L m 0 1 hC0 83qad qde2 5 0 10 20 30 40 50 60 0 0 5 1 1 5 2 2 5 3 Tc 60 0 C a 0 10 20 30 40 50 60 0 5 1 1 5 2 2 5 3 COP Xin X2 g kg Sensible cooling effect Latent cooling effect COP Tc 40 0 C Cooling effect kJ kg b Xin X2 0 10 20 30 40 50 60 1 01 52 02 53 03 5 Pressure ratio 0 5 1 1 5 2 2 5 3 Tc 20 0 C c Fig 2 The effect of pressure ratio on the system performance international journal of refrigeration 31 2008 189 196191 effi ciencies of turbine and compressor effectiveness of heat exchangers and the rotor characteristics 3 Result and discussion The system coeffi cient of performance the amount of humid ity removed from the dock area sensible cooling effect and latent cooling effect are calculated for different values of pres sureratio RP warehousetemperature Tc the effectofinside dockconditions Tin RHin andoutdoorairconditions To RHo Also the effect of the air cycle and the rotor parameters on the system performance is evaluated 3 1 The effect of warehouse temperature The values of COP sensible cooling effect latent cooling effect and the change in the humidity are estimated at different valuesofTc andRPinFig 2 Fromthisfi gure forthesamevalue of Tc the values of sensible cooling effect latent cooling effect and the change in the humidity increase by increasing RP On the other handCOP increasesby increasing RP beforereaching to its maximum value at about the value of RP 2 0 after this the COP value decreases slightly by increasing RP AtthesamevaluesofRP thechangein thehumidity latent cooling effect and sensible cooling effect slightly increase by increasing Tc Also the COP increases by increasing Tc Also from the fi gure it is clear that the latent cooling effect in most cases is nearly the same as the sensible cooling effect this means that using desiccant system increases coeffi cient of performance of the cycle by about 100 The change in the humidity Xin X2 reaches to about 2 8 g kg as a maxi mum value so the air supplied to the dock area has a very low absolute humidity about 0 4 g kg At this condition air dew point is about 25 C having a potential to remove a large amountofmoisturefromthedockareaandprevent condensation 10 15 20 25 30 35 40 0 5 1 1 5 2 2 5 Tin 0 0 C a 10 15 20 25 30 35 40 Cooling effect kJ kg 0 5 1 1 5 2 2 5 COP Xin X2 g kg Sensible cooling effect Latent cooling effect COP Tin 5 0 C b Xin X2 10 15 20 25 30 35 40 5060708090 RHin 0 5 1 1 5 2 2 5 Tin 10 0 C c Fig 3 The effect of dock air conditions on the system 0 5 10 15 20 25 30 0 5 1 1 5 2 2 5 To 25 0 C a 0 5 10 15 20 25 30 Cooling effect kJ kg 0 5 1 1 5 2 2 5 COP Xin X2 g kg Sensible cooling effect Latent cooling effect COP To 30 0 C b X in X2 0 5 10 15 20 25 30 4050607080 RHo 0 0 5 1 1 5 2 2 5 To 35 0 C c Fig 4 The effect of outdoor air conditions on the system performance international journal of refrigeration 31 2008 189 196192 3 2 The effect of inside dock air conditions Fig 3 shows the effect of inside dock conditions Tin RHin on the system performance By increasing RHinat the same value of Tin COP latent cooling effect and the change in humidity increase the reason behind this could be that increasing RHinincreases the potential to remove moisture from air Also at constant inside dock air relative humidity RHin by increasing Tin COP latent cooling effect and the change in humidity increase On the other hand the sensible cooling effect is nearly the same for different values of Tin and RHin 3 3 The effect of outdoor air conditions The effect of outdoor air temperature Toand relative humidity RHoon the system performance is shown in Fig 4 As can be seen from the fi gure at the same outdoor air temperature COP latent cooling effect and the change in humidity de crease by increasing RHo this is because that the capability of outdoor air for adsorbing humidity from desiccant rotor decreases by increasing RHo At the same outdoor air relative humidity RHo by increas ing To sensible cooling effect latent cooling effect the change in humidity and COP slightly decrease 3 4 The effect of air velocity and mass fl ow rate ratio Figs 5 9showtheeffectofairvelocityandmassfl owrateratio on the system performance for different values of rotor speed As we can see from these results by increasing mass fl ow rate ratio COP signifi cantly increases On the other hand COP de creases by increasing air velocity The reason behind that is when the air velocity is increased mass transfers between air and adsorbent are restrained because of a reduced residence time of the air in the adsorbent wheel Heat transfers are also restrained but to a small extent because they are more rapid than mass transfers so the rotary dehumidifi er tends to work as a sensible heat exchanger which is not effi cient for the de humidifi cation process Also from these fi gures there is an 0 7 0 9 1 1 1 3 1 5 23456789101112 N rph COP uad 3 0 m sec ude 3 0 m sec Vad Vde 4 Vad Vde 5 Vad Vde 6 Vad Vde 7 Vad Vde 3 Fig 7 The effect of mass fl ow rate ratio on the system performance uad 3 0 m s and ude 3 0 m s 0 8 1 1 2 1 4 23456789101112 N rph COP uad 1 0 m secude 1 0 m sec Vad Vde 4 Vad Vde 5 Vad Vde 6 Vad Vde 7 Vad Vde 3 Fig 5 The effect of mass fl ow rate ratio on the system performance uad 1 0 m s and ude 1 0 m s 0 8 1 1 2 1 4 23456789101112 N rph COP Vad Vde 4 Vad Vde 5 Vad Vde 6 Vad Vde 7 uad 2 0 m sec ude 2 0 m sec Vad Vde 3 Fig 6 The effect of mass fl ow rate ratio on the system performance uad 2 0 m s and ude 2 0 m s 0 8 1 1 2 1 4 23456789101112 N rph COP uad 2 0 m sec ude 1 0 m sec Vad Vde 4 Vad Vde 5 Vad Vde 6 Vad Vde 7 Vad Vde 3 Fig 8 The effect of mass fl ow rate ratio on the system performance uad 2 0 m s and ude 1 0 m s international journal of refrigeration 31 2008 189 196193 optimum rotation speed for every operating condition this is duetotherotationspeedwhichshouldbelowenoughforcom plete regeneration or rapid cooling of the rotor but enough to keep the adsorbent far from equilibrium The result of these confl icting effects yields the optimum rotation speed Fig 10 summarizes the optimum values of COP for the above operating conditions According to this result optimum COP increases by increasing mass fl ow rate ratio On the other hand optimum COP decreases by increasing air velocity 3 5 The effect of rotor length The system performance at different values of rotor length is shown in Figs 11 13 From these fi gures COP generally in creases by increasing rotor length Also the optimum rotating speed decreases by increasing the rotor length the reason be hind is that by increasing the rotor length the adsorption ca pacity of desiccant rotor increases so it needs a long time to reach the equilibrium condition Also from these results for higher values of air velocities COP signifi cantly increases by increasing rotor length 3 6 The effect of air cycle effi ciencies The effect of air cycle effi ciencies hC hTand hE on the system performance is shown in Fig 14 As can be seen from the fi g ure COP signifi cantly increases by increasing both the values of isentropic effi ciency of compressor and the turbine Also from the fi gure COP increases by increasing the effectiveness of the heat exchanger Also this means that by developing the air cycle components compressor turbine and heat ex changers the potential of using air cycle refrigerator in refrig eration and air conditioning signifi cantly increases 4 Comparison between the performance of the present system and the performance of the conventional system The performance of the proposed system is compared with the performance of a conventional system The conventional 0 9 1 1 1 1 2 1 3 2345678 Vad Vde optimum COP uad ude 1 0 m sec uad ude 3 0 m sec uad 2 0 m sec ude 1 0 m sec uad 3 0 m sec ude 1 0 m sec uad ude 2 0 m sec Fig 10 Maximum COP at different mass fl ow rate ratios and air velocities 0 8 1 1 2 1 4 123456789101112 N rph COP L 0 10 m L 0 15 m L 0 20 m uad 1 0 m secude 1 0 m sec Fig 11 The effect of rotor length on the system performance uad 1 0 m s and ude 1 0 m s 0 8 1 1 2 1 4 23456789101112 N rph COP uad 3 0 m secude 1 0 m sec Vad Vde 4 Vad Vde 5 Vad Vde 6 Vad Vde 7 Vad Vde 3 Fig 9 The effect of mass fl ow rate ratio on the system performance uad 3 0 m s and ude 1 0 m s 0 8 1 1 2 1 4 1 6 123456789101112 N rph COP L 0 10 m L 0 15 m L 0 20 m uad 2 0 m sec ude 2 0 m sec Fig 12 The effect of rotor length on the system performance uad 2 0 m s and ude 2 0 m s international journal of refrigeration 31 2008 189 196194 system consists of a vapor compression cycle and a desiccant rotor as shown in Fig 15 For the vapor compression cycle R22 is used as a refrigerant the compression process is adiabatic with isentropic effi ciency equals to 75 the difference be tween condensation temperature and outdoor air tempera tureis10 C andthedifferencebetweenevaporation temperature and the warehouse temperature is 10 C Refrig erant enters the expansion device as saturated liquid and leaves the evaporator as saturated vapor The desiccant rotor operates at the conditions as the desiccant rotor combined with the air cycle and regenerated by using a gas heater For air cy