Analysis of Two-stage High- temperature Heat Pump Efficiency

Transcription

1 Iue 5, Volume 6, 01 Analyi of Two-tage High- temperature Heat Pump Efficiency M. Jotanović, G. Tadić, J. rope, D. Goricanec Abtract The paper how the potential benefit of uing low temperature heat ource with high temperature pump. Two-tage heat pump with flah unit and heat exchanger are decribed. To determine the characteritic parameter of the two heat pump a computer program wa deigned, with which we could determine the dependence of COP and compreor preure ratio from evaporating temperature. An economic analyi of the utification of the ue of heat pump wa made. eyword Heat tranfer, high temperature heat pump, computer programme, economic analyi F I. INTRODUCTION OR a long time now, people have experienced the negative conequence of the greenhoue effect. The ituation i alarming and that i why we are earching for the way how to avoid the approaching natural catatrophy. It threaten to detroy mankind unle we do omething immediately. At variou conference and forum cientit are etablihing that the only poible method of reducing air pollution i the efficient ue of energy together with the development of new technologie and ytem, a well a the ue of renewable ource of energy. The conequence of an intenive exploitation of energy and energy dependence refer decidedly to the countrie with poor ource of energy and which are claified into a group of the mot environmentally endangered countrie in Europe. The complicated ituation in thee countrie doe not allow them to deal with energy iue by imply fulfilling demand and deire of energy conumer. With imple and cheap deciion on the rationalization, it i poible to achieve coniderable aving and have unpolluted environment. Studie and reearche in indutry uually how that diadvantage in energy converion are on the ide of technological procee; thi mean in the conumption of energy, rather than in the preparation of energy. When it come to the operation of an energy plant, there i often a dilemma how to improve pecific conumption of energy in M. Jotanović and G. Tadić are with the Faculty of Technology, Univerity of Eat Saraevoaraka bb, Zvornik, BOSNIA AND HERZEGOVINA ( otanovicm@mail.ru, gtadic.tf@gmail.com). J. rope and D. Goricanec are with the Faculty of Chemitry and Chemical Engineering, Univerity of Maribor, Smetanova ul. 17, 000 Maribor, SLOVENIA (correponding author to provide phone: ; uri.krope@uni-mb.i, darko.goricanec@uni-mb.i). procee, how to increae efficiency, how to direct and convert energy more efficiently, how to ue wate heat and replace combution of liquid and gaeou fuel with other ource of energy [1,, 3]. Reearche have come up with an innovative olution of uing wate low-temperature water (45 C) for the purpoe of long-ditance heating of building at the temperature of 90/70 C. In thi repect, the Faculty of Chemitry and Chemical Engineering of Univerity of Maribor developed an innovative high-temperature heat pump (TP) made by a Japanee company named Mycom in 010 [4, 5, 6]. Thi i the firt example in the world and ha uneen application poibilitie in indutry, which i confirmed by the Intergovernmental platform for R&D in Europe reward. The eence of thi paper i focued on the following: Computer imulation of techno-economic effect of the ue of different refrigerant, Computer imulation of the effect of different temperature of low-temperature ource and different refrigerant on COP, Selection of technically and economically optimal heat pump, Production of a computer upported model which will enable efficient uage of different low-temperature ource for the purpoe of long-ditance heating, greenhoue heating, heating of dry-kiln, in proce indutry (technological water), anitary water, etc. Thi would provide great financial benefit and environment protection. The reearch wa conditioned by: the neceity of intenive exploitation of alternative ource of energy and low-temperature ource, the neceity of a conitent and coordinated approach to the exploitation of the heat from low-temperature ource of energy, fulfilment of the recommendation and requet of the EU regarding efficient energy conumption and environment protection. II. HIGH TEMPERATURE HEAT PUMP Heat pump i a proce device which i ued for heating. It operation principle are baed on the removal of low temperature from the environment, which i then given on a higher temperature level. The ource for the removal of energy are air, water or ground. 350

2 Iue 5, Volume 6, 01 The rapid development of heat pump tart with the great oil crii, when people tarted earching intenively for technological olution for the replacement of foil fuel with other ource of energy. Due to tricter law on pollution, the awarene of energy, environmental awarene of conumer and the rie in the price of energy, heat pump are becoming more and more intereting becaue they are energy productive and environmentally friendly. With the development of new technologie, the method of application, improvement of performance, and reduction of the dimenion and ma of thee device on the other hand, the application of heat pump increae. The mot recent type of heat pump have the poibility of operation in high-temperature ytem of longditance heating. They ue refrigerant which do not harm the environment and have a good ratio between the electric power ued and the heat obtained, which i even up to 1:9. Scientit predict that heat pump with different implementation method in the future will be the fundamental part in low-temperature and high-temperature heating ytem a an environmentally acceptable olution. An important fact in the application of heat pump i that they ue much le primary energy compared to ga and oil boiler. Thi mean that they can reduce the emiion of CO and other gae by 31 % to 60 %. For example, in cae where the operation of heat pump ue electric power from renewable ource, the emiion of greenhoue gae i reduced to minimal value, which will become a neceity ooner or later [7]. A. Operation of heat pump Compreion heat pump conit of an evaporator, condener, compreor and regulation valve. The operating principle of a one-tage heat pump (ETČ) are hown in Fig. 1 [8,9]. In the evaporator, the refrigerant, which i liquid, after the reduction of preure on the regulation valve, i evaporated at low temperature and aturation preure. Then the refrigerant receive a heat flow from the environment Φ U. Refrigerant vapour i then taken to the compreor where it i compreed to the preure where the condenation temperature i higher than the temperature of the environment. Vapour i then cooled and condened. It thu give a flow of heat Φ to the urrounding area that i heated. The liquid refrigerant i then again taken to the evaporator through the regulation valve. Since the compreor i the main proce unit, it i neceary to conider the following when planning different implementation: the ratio between preure, maximum allowed ga preure maximum allowed temperature, rotation peed, power and volume flow of the compreor, a well a other characteritic. Table 1 how characteritic of ome compreor made by Mayekawa - Japan. Table 1: Performance of the Mayekawa compreor [10] Operating characteritic Maximum allowed ga preure (Pd) [bar] Maximim preure of ga pumpimg out (P) [bar] Maxium average ga preure [bar] Minimum preure of pumping out of the ga [bar] Maximum differential preure [bar] Maximum preuure ratio [ / ] Minimum preure of introduction of oil [bar] Maximum preure of intro uction of water [bar] Minimum preure of the introduction of water [bar] Maximum allowed ga temperature [ C] Min. temp. of ga pumping out Maximum temperature of the introduction of water [ C] Min. temperature of oil introduction [ C] Maximum water temperature [ C] Minimum temperature of water introduction [ C] Maximum rotation peed [revol./min] Minimum rotation peed [revol./min] Maximum trength of. drive [kw] Volume flow of piton compreor [ /h] Type L, WA, WB WBH L, WA, WBH, WB 4WBH, 6WBH 4WB, 6WB, 14WB All model L, WA, WB WBH WBH L, WA, WBH, WB, WA L, WA,WBH, WB All model All model WA, WBH, WB,L All model WA, WBH, WB,L All model All model All model L WA WA WBH, WB L WA, WBH, WB L, WBH, WB WA WA L WBH Operating parameter Ammonia: 3 HFC: 4 3,6 0 6,86 5, revol./min; 8, 1100 revol./min; 6, revol./min; 5,38-0,733 19,6 14,7 0,0 15, Ammonia: 8 Ammonia: 9 HFC: Ammonia: 140 HFC: 135 Ammonia:140 HFC: Ammonia: 50 HFC: (1500) Comment Only for HFC Neceary for the introduction of oil Pd - P Allowed temperature of the ga With oil protection At the outlet of the refrigerant 351

3 Iue 5, Volume 6, 01 Fig. 1: High-temperature heat pump operating principlediagram B. Review of the ource for heat pump When it come to the ource of energy for the operation of heat pump, the mot advantageou today are renewable ource uch a wind, hydroenergy, olar energy, geothermal energy a well a wate energy from proce indutry. Globally peaking, geothermal reource refer to the heat accumulated deep in the ground, i.e.in the ma of tone and fluid in the Earth' crut. A cro-ection of a younger igneou intruion lie on the urface, with conduction through deep tectonic dicordance and through inlet volcanic channel. Geothermal energy i alo the reult of the decompoition of radioactive element in different chemical procee that occur in the Earth' crut. The Earth' temperature i increaed by 1 C (geothermal degree) to every 33 m of it depth. If it increae more rapidly, we peak of a poitive anomaly or increaed geothermal degree, or geothermal gradient. The main indicator for an area to have a great potential for exploitation i the geothermal anomaly, which, hydrologically and hydrochemically, give a clear decription of the poibilitie and method for the exploitation of the energy potential. Conidering the poibility of exploitation, geothermal energy i claified into: hydrogeothermal energy and petrogeothermal energy. The firt type refer to the geothermal energy of liquid and gaeou fluid, wherea the other i the geothermal energy of a ma of rock. The ignificance of geothermal energy i bet decribed by the two natural propertie: the tability and conitency of the heat flow and heat, which i kept in rock. If we ay that a heat flow which i tranferred from the inide of the Earth toward the urface of the Earth i conidered to be geothermal energy, than we can ay that thi energy i a renewable ource of energy. eeping energy encloed in rock i a complex proce which depend on the three form of heat tranfer. Convection of water and magma i uch a rapid energy tranfer that we can claify geothermal energy into renewable ource, becaue the energy removed i compenated at the ame peed. If, on the other hand, energy i tranferred only with heat exchange, we can hardly claify it into renewable ource ince the time contant of energy compenation i coniderably longer than the contant of the exploitation of the ame energy. All conventional method of the exploitation of geothermal energy are baed on the removal of energy from natural geothermal ytem in which energy i upplied by water and tranferred at the ame time within the ytem itelf. Water take energy to the urface where it i exploited. During production, preure begin to decreae, which caue the upply of additional water and energy to the ytem which i exploited. Such condition are typical for renewable ource of energy, where compenation and exploitation of energy occur at the ame rate. Geothermal energy reult form three factor: Energy which i tored in rock and fluid of the Earth' crut, Heat flow evolved by exchange and Energy flow through the Earth' crut in the form of the tranfer of matter (magma, water, vapour, gae). Occurrence on the urface (hot pring) are uually the bet indicator that geothermal ource exit in the depth of the Earth' crut; however, it i poible to recognize the ource ometime even without them. It i eentially thought that the total value of uninvetigated or unidentified geothermal ource i greater than the known ource. Low-temperature ource are more commonly found than thoe higher -temperature. It hould be known that the frequency of ditribution of ource depending on temperature i the function of an ordinary exponential curve. In order to determine the lower limit of the potential of low -temperature ource on the Earth we mut conider the following approximation: the frequency of occurrence of ource a the function of temperature ha the propertie of an exponential curve, the energy of a particular ource i linearly dependent on it temperature and the volume of a geothermal reervoir i independent of it temperature. Conventional exploitation of geothermal energy i uually claified into: high - temperature ource with the temperature of water above 150 C, which are ued for the production of electricity and low - temperature ource with the temperature of water below 150 C, which are mainly ued directly for heating. Conidering the data for the Earth' geothermal gradient, the maority of European countrie have a number of unexploited low - temperature geothermal energy ource. Only a mall number of countrie have high-enthalpy ource which are uitable for the production of electric power, on the depth that till utify the economic apect of exploitation. Direct exploitation of geothermal energy may include different market. In larger, indutrially developed countrie a large part of heat i ued at temperature between 50 and 100 C. 35

4 Iue 5, Volume 6, 01 Reearche into the uage of geothermal energy in the world regarding the purpoe have hown that energy i motly ued for heating, half of it being for temperature regime below 100 C. Ditribution of exploitation purpoe i the following: heating 35 %, pool (balneology) 15 %, greenhoue 14 %, fihpond10 %, indutry 10 %, air-conditioning 1 %, and heat pump 13 %, other %. In Europe and other part of the world geothermal energy i ued for the production of electric power, for utility heating of reidence and indutrial facilitie, in agriculture for heating of greenhoue, in tourim, etc. The bet condition for the exploitation of geothermal energy in Europe are in Iceland, Italy and Greece. Slovenia i alo rich with low-temperature geothermal energy. refrigerant i ued for heat tranfer in both tage. The flah heat pump diagram i hown in Fig. 3. Due to the two age of compreion there i a poibility of uing lower temperature ource (10 to 30 C) for heating. In uch implementation it i important which refrigerant will be ued, becaue it ha to have good thermo-phyical characteritic a well a properly determined preure. C. The poibility of uing geothermal energy The maority of flat in citie are old, with poor inulation and uing high-temperature ytem of radiator. When the temperature of geothermal water that we wih to ue i 4 C, it i not high enough for high-temperature heating uing a heat pump. By intalling a heat pump into the heating ytem the energy of geothermal water i additionally ued allowing imultaneouly additional uer of heat energy in the heating ytem, which ha a poitive effect on the environment and economic utifiability. Fig. 3: Two-tage high-temperature heat pump with a flah expander 1) Calculation When performing calculation for the two-tage high temperature heat pump with a flah expander it i important to determine correctly the intermediate preure of the refrigerant, between the firt and the econd level/tage, which i equal to the following according to the general equation: p p p (Pa) (1) M S,1 T, where: p - vapour preure of the refrigerant on the inlet ide of S,1 the firt tage of compreion (Pa) p - vapour preure of the refrigerant on the DELIVERY T, SIDE of the econd tage of compreion (Pa) Fig. : Repreentation of the exploitation of geothermal water for long-ditance heating D. The two tage high temperature heat pump with a flah expander The two-tage high-temperature heat pump with a flah expander include two compreor and two expanion valve. The two tage of compreion are neceary due to the reduction of a high ratio of preure, which ha negative effect on the performance of the compreor (energy conumption, cooling...). Middle tage of compreion i eparated by the flah/eparate flah expander, and the ame Intermediate temperaturet M, which for the firt tage of compreion of the refrigerant i the condenation temperature T,1, and for the econd tage of compreion pa it i the temperature of evaporation TU, i calculated uing the following equation: T M pm B ± B 4 A( C log [ Pa ] A 1 () () 353

5 Iue 5, Volume 6, 01 Heat flow, of the firt tage Φ and of the econd TČ, 1 tage Φ are calculated uing the equation: TČ, Φ q h h ) (W) (3) ( T Č, 1 m, S 1 g, l, 3 q (, m, S h TČ g, 6 hl, 7 Φ ) (W) (4) where: h - pecific enthalpy of refrigerant vapour on the g,6 DELIVERY SIDE of the compreor of the econd tage of the heat pump (J/ ) h - pecific enthalpy of the liquid refrigerant in the l,3 condener of the econd tage of the heat pump (J/ ) The heat number of the two - tage heat pump with a flah expander i calculated uing the following equation: COP Φ TČ, G ( / ) (5) P,1 + P, Other parameter of the high-temperature heat pump with a flah expander and the dimenion of the compreor for both tage are calculated according to the ame principle a for one-tage high-temperature heat pump [11, 1]. ) Calculation programme DATA REFRIGERANT: R-600a (izo-butan) CONSTANTS FOR THE CALCULATION OF THE PRESSURE FOR THE TEMPERATURE AND au 0 AU 0 bu BU cu CU CONSTANTS FOR THE CALCULATION OF THE DENSITY OF THE REFRIGERANT at AT bt BT.7801 ct CT 9.66 CONSTANTS FOR THE CALCULATION OF THE DENSITY OF REFRIGERANT VAPOUR ap 0 AP 0 bp BP cp CP CONSTANTS FOR THE CALCULATION OF THEENTHALPY OF REFRIGERNAT LIQUID xt.463 XT.806 yt YT CONSTANTS FOR THE CALCULATION OF THE ENTHALPY OF REFRIGERANT VAPOUR xp 0 XP 0 yp YP zp ZP COMPRESSOR TYPE: 8RWH NUMBER OF REVOLUTIONS [1/min]: N 1000 min 1 OPERATION OF THE COMPRESSOR [%]: µ 1 UTILIZATION OF THE COMPRESSOR [/]: η 0.7 BOILING TEMPERATURE OF THE REFRIGERANT []: CONDENSATION TEMPERATURE OF THE REFRIGERANT []: MEDIUM TEMPERATURE []: ( + ) T T COMPRESSIBILITY FACTOR: χ UNUSED SPACE FOR FILLING THE COMPRESSOR ε 0.14 TWO-STAGE HEAT PUMP WITH A FLASH EXPANDER PRESSURE ON THE INLET SIDE OF THE COMPRESSOR [Pa]: au + bu + cu Pu 10 Pa Pu Pa SPECIFIC ENTHALPY OF THE REFRIGERANT VAPOUR AT THE BOILING TEMPERATURE Hpu xp yp J + + zp Hpu J SPECIFIC ENTHALPY OF THE LIQUID AT THE BOILING TEMPERATURE Htu XT + YT J Htu J REFRIGERANT VAPOUR DENSITY AT THE BOILING TEMPERATURE ρ pu ap bp + + cp ρ pu REFRIGERANT LIQUID DENSITY AT THE BOILING TEMPERATURE σtu at bt + + ct σtu PRESSURE ON THE DELIVERY SIDE OF THE COMPRESSOR [Pa]: 354

6 Iue 5, Volume 6, 01 AU + BU + CU Pk 10 Pa Pk Pa SPECIFIC ENTHALPY OF REFRIGERANT VAPOUR AT THE CONDENSATION Hpk XP YP J + + ZP TEMPERATURE Hpk J SPECIFIC ENTHALPY OF REFRIGERANT LIQUID AT THE CONDENSATION TEMPERATURE Htk XT + YT J Htk J REFRIGERANT VAPOUR DENSITY AT THE CONDENSATION TEMPERATURE ρ pk AP BP + + CP ρ pk REFRIGERANT LIQUID DENSITY AT THE CONDENSATION TEMPERATURE σtk AT + BT + CT σtk INTERMEDIATE PRESSURE CALCULATION Pm PuPk Pm Pa INTERMEDIATE TEMPERATURE CALCULATION Tm log Pm 1 cu Pa Tm bu Tm SPECIFIC ENTHALPY OF REFRIGERANTE VAPOUR AT INTERMEDIATE TEMPERATURE Tm Hpm Tm XP YP Tm J + + ZP Hpm J SPECIFIC ENTHALPY OF REFRIGERANT LIQUID AT INTERMEDIATE TEMPERATURE Tm Htm XT Tm + YT J Htm J REFIRGERANT VAPOUR DENSITY AT THE INTERMEDIATE TEMPERATURE Tm Tm ρ pm AP Tm BP + + CP ρpm REFRIGERANT LIQUID TEMPERATURE AT THE INTERMEDIATE TEMPERATURE Tm Tm σtm AT BT Tm + + CT σtm Pm r1 PRESSURES RATIO FOR THE FIRST COMPRESSOR Pu r TEMPERATURE AT THE OUTLET OF THE FIRST COMPRESSOR []: χ 1 Pm χ T Pu T SPECIFIČ ENTHALPY OF REFRIGERANT VAPOUR AT THE TEMPERATURE T T Hp XP YP T J + + ZP Hp J FILLING RATE OF THE CYLINDER IN THE COMPRESSOR 1 χ λ 1 ε r1 1 λ THE REAL VOLUME FLOW OF THE COMPRESSOR v v m3 qv 0.09 m3 Pad1 v Pu Pad W COMPRESSOR STRENGTH Pad1 Pk1 η Pk W χ χ 1 VOLUME FLOW AT THE INLET SIDE OF THE COMPRESSOR v m3 Vkλ ( χ 1) χ r1 1 MASS FLOW AT THE INLET SIDE OF THE COMPRESSOR qm1 qm MASS FLOW AT THE DELIVERY SIDE OFF THE COMPRESSOR qv 0.09 m3 VOLUME FLOW OF THE REFRIGERANT IN THE COMPRESSOR qv v ρpu ρpm ADIABATIC COMPRESSION STRENGTH MASS FLOW ON THE DELIVERY SIDE OF THE COMPRESSOR qmt1 qvρpm qmt Φu1 qm1( Hpu Htm) 1000 v ρpu HEAT FLOW OF THE BOILED PART OF THE HEAT PUMP THE FIRST STAGE [W]: PRESSURES RATIO OF THE SECOND COMPRESSOR Pk r Pm Φk1 qm1( Hp Htm) 1000 r

7 Iue 5, Volume 6, 01 Φu W HEAT FLOW OF THE CONDENSATION PART OF THE HEAT PUMP THE FIRST STAGE [W]: Φk W HEAT NUMBER OF THE FIRST STAGE OF THE HEAT PUMP [/]: Φ k1 qm ( Hpm Htk ) 1000 qm.37 MASS FLOW ON THE INLET SIDE OF THE SECOND COMPRESSOR Φ k1 COPG Pk1 COPG TEMPERATURE AT THE OUTLET FROM THE SECOND COMPRESSOR []: T FILLING RATE OF THE CYLINDER IN THE COMPRESSOR 1 χ λ 1 ε r 1 λ VOLUME FLOW OF THE COMPRESSOR IN THE SECOND STAGE OF THE HEAT PUMP qm Vk ρ pm 3) Calculation reult The calculation of the two-tage high-temperature heat pump with a flah expander for refrigerant R-600a, R-90, R45fa and R-134a were conducted at two different temperature of condenation (t 70 C and t 80 C). Reult are hown in Fig. 4 and Fig. 5. Fig. 6 and Fig. 7 how the dependence of het number and ratio between preure in the compreor. The main obective of the implementation of the two - tage heat pump with a flah expander i to try to reduce the trength which i neceary for the compreor to operate in order to reduce the ratio between preure and increae the ueful heat flow of the condenation part of the pump. When conducting the calculation, the fact that the compreor in both tage are equal, wa taken into conideration. The compreor were piton compreor WBH choen from the catalogue of the manufacturer called MYCOM. The volume flow of the compreor i 637 /h, maximum operating preure i p max.,0mpaand maximum trength of the compreor i P max. 145,0 kw. Vk 0.18 m3 THE REAL VOLUME FLOW OF THE SECOND COMPRESSOR v m3 T3 Tm Pk Pm Hp J χ 1 χ SPECIFIČ ENTHALPY OF THE REFRIGERANT VAPOUR AT THE TEMPERATURE T3 T3 Hp3 XP T3 YP + + ZP J v Vk λ VOLUMSI PRETO SREDSTVA U DRUGOM OMPRESORU qv Vk ρ pm ρ pk qv m3 ADIABATIC COMPRESSION STRENGTH IN THE SECOND STAGE χ Pad v Pm ( χ 1 ) χ r χ 1 1 Pad W Pad Pk COMPRESSOR STRENGTH IN THE SECOND STAGE η Pk W MASS FLOW ON THE DELIVERY SIDE OF THE SECOND COMPRESSOR qmt qvρpk HEAT FLOW OF THE HEAT PUMP IN THE SECOND STAGE [W]: Φk qm( Hp3 Htk) 1000 Φk W HEAT NUMBER OF THE HEAT PUMP [/]: Φk COPG Pk + Pk1 COPG Fig.4: The two-tage high-temperature heat pump with a flah expander at the condenation temperature of the refrigerant of t 70 C Fig. 5: The two-tage high-temperature heat pump with a flah expander at the condenation temperature of the refrigerant t 80 C 356

8 Iue 5, Volume 6, 01 Analyi of the reult how that the ue of lowtemperature heat ource i poible by uing a two-tage hightemperature heat pump with a flah expander, but there are certain limitation regarding the operation of the compreor and the choice of the refrigerant. 70 C, the heat flow of the heat pump i Φ TČ 473,0 kw and the energy for the operation of the compreor i P,1 44,7 kw or P, 66,4 kw. Heat number i COP 4,56, the ratio between preure ha the value of r 1,898. For the propoed type of the heat pump, it i undertandable to ue the refrigerant R-600a, which ha good thermo-phyical characteritic in a wider temperature range (low preure and high enthalpy) For the operation of the heat pump to be practical, when other refrigerant are ued, it i neceary to determine a uitable tandard compreor for every level of compreing. E. The two-tage heat pump with a heat exchanger The two-tage heat pump with a heat exchanger conit of two one-tage heat pump. The tage are eparated by the heat exchanger, which erve a the condener for the firt tage and a evaporator for the econd tage. The tage are eparated in uch way that refrigerant are not in contact. Thi mean that each tage ha it own refrigerant. The heat pump diagram i hown in Fig. 8. [13, 14]. Fig. 6: Heat number of the two-tage heat pump Fig. 8: The two-tage heat pump with a heat exchanger Fig. 7: The ratio between the preure in the compreor of the two-tage heat pump with a flah expander The diagram how that uing refrigerant R-90 and R- 134a we achieve high value of heat flow, but for the propoed temperature the conumption of energy neceary for the operation of the compreor i to high in relation to the choen compreor. Refrigerant R-45fa give low value of the heat flow a well a the low conumption of energy. For the refrigerant R-45fa the compreor i overized. The mot uitable characteritic parameter of the two-tage heat pump with a flah expander were obtained uing the refrigerant R- 600a, which enable a wide cope of operation and, on the other hand, condition for optimal operation of the compreor are in accordance with the intruction from the manufacturer. With the ue of the refrigerant R-600a at the boiling temperature t U 0 C and condenation temperature t 1) Equation for the calculation Calculation of the characteritic parameter of the operation of the two-tage heat pump with a heat exchanger i more imply conducted according to the principle of the two one-tage pump, where it i aumed that the heat flow of the condenation of the firt tage Φ,1, i equal to the heat flow of boiling of the econd tage Φ U,. It hold that: Φ q ( h h ) (W) (6),1 m, S,1 g, l,3 Φ Φ (W) (7),1 U, It follow that the ma flow of the refrigerant on the inlet ide of the compreor of the econd tage of the heat pump i equal: 357

9 Iue 5, Volume 6, 01 q m, S, Φ U, ( h h ) g,5 l,8 (/) (8) where: h - pecific enthalpy of the refrigerant which enter the g,5 compreor in the econd tage of the heat pump (J/ ) h - pecific enthalpy of the liquid refrigerant which enter l,8 the evaporator in the econd tage of the heat pump (J/ ) Heat number of the two-tage heat pump with a heat exchanger i calculated: C O P G P Φ T Č, + P,1, ( / ) (9) Condenation temperature of the firt tage T,1 and the boiling temperature of the econd tage TU, of the refrigerant are calculated uing the method of the average value between the temperature of boiling of the firt tage TU,1 and the temperature of the condenation of the econd tage T,. T T + T U,1, M () (10) The poibility of uing refrigerant in both tage require that we determine the characteritic of the compreor for each tage eparately according to the ame principle a for the one-tage heat pump [11, 1]. ) Calculation programme DATA FOR THE CALCULATION OF THERMODYNAMIC DATA Refrigerant REFRIGERANT: R-717 Ammonia CONSTANTS FOR THE CALCULATION OF PRESSURE AT THE TEMPERATURE II au 0 AU 0 bu BU 1175 cu CU CONSTANTS FOR THE CALCULATION OF THE REFRIGERANT LIQUID DENSITY at AT bt BT ct CT CONSTANTS FOR THE CALCULATION OF THE REFRIGERANT VAPOUR DENSITY ap AP bp 0.69 BP cp CP CONSTANTS FOR THE CALCULATION OF THE REFRIGERANT LIQUID ENTHALPY xt XT yt YT CONSTANTS FOR THE CALCULATION OFTHE REFRIGERANT VAPOUR ENTHALPY xp XP yp YP zp ZP 573. TYPE OF COMPRESSOR: 8RWH NUMBER OF REVOLUTIONS OF THE COMPRESSOR [1/min]: N 1000 min 1 OPERATION OF THE COMPRESSOR [%]: µ 1 EFFICIENCY OF THE COMPRESSOR [/]: η 0.7 TEMPERATURE OF BOILING OF THE REFRIGERANT []: TEMPERATURE OF CONDENSTATION OF THE REFRIGERANT []: AVERAGE TEMPERATURE []: ( + ) T T COMPRESSIBILITY FACTOR I: χ UNUSED SPACE FOR FILLING THE COMPRESSOR ε 0.14 VOLUME FLOWOF THE COMPRESSOR [m3/h]: Vk m3 PRESSURE ON THE INLET SIDE OF THE COMPRESSOR [Pa]: au + bu + cu Pu 10 Pa Pu Pa REFRIGERANT LIQUID DENSITY AT THE BOILING TEMPERATURE σtu at + bt + ct AU + BU + CU σtu P 10 Pa PRESSURE ON THE DELIVERY SIDE OF THE COMPRESSOR [Pa]: P Pa SPECIFIČ ENTHALPY OF THE REFRIGERANT VAPOUR ENTHALPY AT THE CONDENSATION TEMPERATURE Hpk xp yp J + + zp Hpk J 358

10 Iue 5, Volume 6, 01 SPECIFIČ ENTHALPY OF REFRIGERANT LIQUID AT THE TEMPERATURE OF CONDENSATION Htk xt + yt J Htk J REFRIGERANT VAPOUR DENSITY AT THE TEMPERATURE OF CONDENSATION ρ pk ap + bp + cp ρpk REFRIGERANT LIQUID DENSITY AT THE CONDENSATION TEMPERATURE σtk at bt + + ct σtk P r THE RATIO BETWEEN THE PRESSURES Pu r.164 THE TEMPERATURE AT THE OUTLET FROM THE COMPRESSOR []: χ 1 P χ T Pu T SPECIFIČ ENTHALPY OF THE REFRIGERANT LIQUID AT THE BOILING TEMPERATURE Htu Htu XT J + YT J REFRIGERANT VAPOUR DENSITY AT THE TEMPERATURE OF BOILING ρ pu 8.75 ρ pu ap bp + + cp SPECIFIC ENTHALPY OF REFRIGERANT VAPOUR AT THE TEMPERATURE OF CONDENSATION T Hp xp T yp J + + zp Hp J FILLING RATE OF THE CYLINDER IN THE COMPRESSOR 1 χ λ 1 ε r 1 λ THE REAL VOLUME FLOWOF THE COMPRESSOR v Vkλ v m3 qv VOLUME FLOW OF THE REFRIGERANT IN THE CMPRESSOR v ρpu ρpk qv 0.04 m3 ADIABATIC COMPRESSION STRENGTH ( χ 1) χ Pad vpu χ r χ 1 1 Pad W COMPRESSOR STRENGTH Pad Pk η Pk 943.8W VOLUME FLOWON THE INLET SIDE OF THE COMPRESSOR v m3 MASS FLOW ON THE INLET SIDE OF THE COMPRESSOR qm vρpu qm 0.64 VOLUME FLOWON THE DELIVERY SIDEOF THE COMPRESSOR qv 0.04 m3 MASS FLOW ON THE DELIVERY SIDE OF THE COMPRESSOR qmt qvρpk qmt 0.64 HEAT FLOW OF THE EVAPORATED PART OF THE FIRST STAGE OF THE HEAT PUMP [W]: Φu1 qm( Hpu Htk) 1000 Φu W HEAT FLOW OF THE CONDENSATION PART OF THE FIRST STAGE OF THE HEAT PUMP [W]: Φk1 qm( Hp Htk) 1000 Φk W HEAT NUMBER OF THE FIRST STAGE OF THE HEAT PUMP [/]: Φk1 COPG Pk COPG 6.76 HEAT FLOW OF THE CONDENSER IN THE FIRST STAGE IS EQUAL TO THE HEAT FLOW OF THE EVAPORATOR IN THE SECOND STAGE Φ Φk1 ΦU Φ ΦU W DATA FOR THE CALCULATION OF THERMODYNAMIC DATA FOR REFRIGERANTS IN THE SECOND STAGE HLADIVO: R-600a (izo-butan) CONSTANTS FOR THE CALCULATION OF PRESSURE AT THE TEMPERATURE I au 0 AU 0 bu BU cu CU CONSTANTS FOR THE CALCULATION OF REGRIGERANT LIQUID DENSITY at AT bt BT

11 Iue 5, Volume 6, 01 ct CT 9.66 CONSTANTS FOR THE CALCULATION OF THE REFRIGERANT VAPOUR DENSITY ap 0 AP 0 bp BP cp CP CONSTANTS FOR THE CALCULATION OF REFRIGERANT LIQUID ENTHALPY xt.463 XT.806 yt YT CONSTANTS FOR THE CALCULATION OF REFRIGERANT VAPOUR ENTHALPY xp 0 XP 0 yp YP zp ZP TYPEOF THE COMPRESSOR: 8RWH Bro obrtaa OF THE COMPRESSOR [1/min]: N 1000 min 1 OPERATION OF THE COMPRESSOR [%]: µ 1 UTILIZATION OF THE COMPRESSOR [/]: η 0.7 TEMPERATURE OF BOLINIHG OF THE REFRIGERANT []: TU TEMPERATURE OF CONDENSATION OF THE REFRIGERANT []: T AVERAGE TEMPERATURE []: ( TU + T ) TS TS COMPRESSIBILITY FACTOR: χ 1.07 PRESSURE ON THE INLET SIDE OF THE COMPRESSOR [Pa]: au + bu + cu TU PU 10 TU Pa UNUSED SPACE FOR FILLING OF THE COMPRESSOR ε 0.14 PU Pa SPECIFIC ENTHALPY OF REFRIGERANT VAPOUR AT THE BOILING TEMPERATURE TU HpU xp yp TU J + + zp HpU J SPECIFIC ENTHALPY OF THE REFRIGERANT LIQUID AT THE BOILING TEMPERATURE HtU HtU XT TU J J + YT REFRIGERANT VAPOUR DENSITY AT THE BOILING TEMPERATURE TU ρpu ap TU + bp + cp ρpu REFRIGERANT LIQUID DENSITY AT THE BOILING TEMPERATURE TU σtu at TU + bt + ct σtu PRESSURE ON THE DELIVERY SIDE OF THE COMPRESSOR [Pa]: AU BU + + CU T P 10 T Pa P Pa SPECIFIČ ENTHALPY OF THE REFRIGERANT VAPOUR AT THE CONDENSATION TEMPERATURE Hp XP T YP T J + + ZP Hp J SPECIFIC ENTHALPY OF THE REFRIGERANT LIQUID AT THE CONDENSATION TEMPERATURE Ht XT T + YT J Ht J REFRIGERNAT VAPOR DENSITY AT THE CONDENSATION TEMPERATURE T ρ p AP T BP + + CP ρp REFRIGERANT LIQUID DENSITY AT THE CONDENSTATION TEMPERATURE T σt AT T BT + + CT σt THE RATIO BETWEEN PRESSURESOF THE COMPRESSOR P r PU r TEMPERATURE AT THE OUTLET OF THE COMPRESSOR []: χ 1 P χ T3 TU PU T ) Reult of the calculation Calculation of characteritic parameter wa conducted for the following refrigerant: R-407c in the firt tage and R-600a in the econd tage R-717 in the firt tage and R-600a in the econd tage R-90 in the firt tage and R-600a in the econd tage 360

12 Iue 5, Volume 6, 01 The reult are given in Fig. 9 and Fig. 10, which how the trength which i neceary for the operation of the compreor and the heat flow of the heat pump. Calculation were conducted for different temperature (t U ) of the refrigerant: from 5,0 C to 5,0 C and the contant temperature of condenation t 70 C. The value of the heat number of the two-tage heat pump with a heat exchanger were calculated for the condenation temperature of the refrigerant t 60 C and t 80 C. Fig. 11 how the dependence between the COP and the refrigerant boiling temperature. For the calculation of all uggeted combination of refrigerant the compreor choen were: a piton compreor with the flow of 300 /h for the firt tage, and a piton compreor with the flow of 700 /h for the econd tage. The main obective of the two-tage pump with a heat exchanger i the ue of a wider range of temperature between the boiling temperature and the condenation temperature. With the propoed way of implementation of the heat pump, the low-temperature ource are maximally ued. Moreover, the heat obtained can be ued for long-ditance or combined heating Fig. 10: The econd tage of the two-tage heat pump with a heat exchanger Fig.11: COP of the two-tage heat pump witha heat exchanger Fig. 9: The firt tage of the two-tage heat pump with the heat exchanger The following combination of refrigerant are optimal for the operation of the two-tage heat pump with a heat exchanger: R-717 in the firt tage and R-600a in the econd tage, R-407c in the firt tage and R-600a in the econd tage. The firt combination give the highet heat flow for the purpoe of heating, the ue of trength i equal or lower than that of the other propoed combination of refrigerant. The problem of thi combination i a narrow range of application, becaue at larger difference between the the condenation temperature and boiling temperature the operation of the compreor i relimited due to high operating preure R-717. For wider operating condition it i neceary to deign the type of compreor which operate at higher operating preure The other combination of the refrigerant give lower value of the heat flow. Here we are not retricted concerning the operation of the compreor in both phae. From the environmental apect, we ued F-gae in the firt tage, which have a negative effect on global warming; however, R-407c refrigerant i currently ued a a replacement for the HCFC refrigerant. Other combination of refrigerant give poorer reult. The operating condition of the compreor are limited or completely impoible. III. ECONOMIC ANALYSIS Buine deciion which determine buine condition of a company affect market competition and innovation and uually direct our buine deciion a to how to make the bet buine policy in a certain period of time. When deciding whether to refue or accept an indutrial proect, the key factor i the economic analyi which i baed on the two aumption: 361

13 Iue 5, Volume 6, 01 invetment, which i a ingle um of money for the implementation of the proect in practice the urplu income in comparion to the expene, which mean a difference between the revenue and the expene for maintenance and operation. The economic analyi of a proect refer to the determination of the value of the above mentioned aumption uing economic method which differ in requirement and accuracy. The choice of the method depend on the tage of the development of the proect. For the economic analyi of the utification of the application of the high -temperature heat pump for heating the method of net preent value wa ued (NPV). The annual inflation rate i alo taken into conideration. NPV repreent the um of all value of money flow. The rule on deciding on the invetment baed on NPV i that the invetment i accepted if NPV i higher than 0 or vice vera. Apart from thi method, the uccefulne of the invetment can be aeed uing a profitability coefficient (), which repreent the relation between the net preent value from elling and the um of value of all expene, on one ide, and the deadline of return on the other [15]. A. Economic utifiability of the application of the hightemperature heat pump The deciion on the invetment i much eaier and impler if it i baed on the calculated parameter for individual implementation of heat pump. The calculation included the model with mot uitable refrigerant. The economic analyi tated that the cot of the invetment are covered from peronal ource and from loan in 30% to 70% ratio. The preent value of the invetment cot C INV e wa calculated according to the equation; C INV N an CTČ C0 + (EUR/year) (11) 0 (1 + r) where: - own fund (EUR) C 0 C - the cot of the heat pump (EUR) TČ r - dicount rate ( / ) N - lifetime of the ytem (year) Annuity factor a n i calculated uing the equation: a n r (1 + r ) + ( / ) (1) n1 a a (1 n1 ra ) 1 r a - dicount tage annuity ( / ) n 1 - time of return annuity ( / ) Maintenance cot of the heat pump were evaluated at % of the purchae price. Net preent value of expene with inflation rate included wa calculated uing the equation: C S N 0.0 C (1 + r ) TČ (EUR/year) (13) (1 + r + r) 0 Net preent value of electric power cot for the operation of the compreor of the pump were calculated uing the equation: C PS N C P t t (1 + r ) (EUR/year) (14) 0 E 1 (1 + r + r) Net preent value of the income from the heat produced conidering the inflation rate i calculated uing the equation: C P N 0 C Φ t t (1 + r ) T T Č 1 (1 + r + r ) (EUR/year) (15) Net preent value of the heat income, conidering the invetment cot and maintenance cot, a well a the conumption of electric energy for the operation of the compreor i calculated uing the equation: C C ( C + C + C ) (EUR/year) (16) P INV S PS where : r - inflation rate ( / ) C - electric power cot (EUR/kWh) E C -the cot of heat for heating (EUR/kWh) T t - operation time per day (h/day) 1 t - operation time per year (day/year) The calculation of economic parameter wa conducted for all three heat pump. The cot of the heat pump wa calculated baed on the cot of individual proce device. The cot of electric power i a known value. The time of operation of the pump i 160 working hour [1]. Table contain data for the calculation of different implementation of heat pump. Due to eaier economic evaluation and comparion of the very method of implementation, the reult for high-temperature pump are hown in Fig. 1. It how NPV at different temperature of boiling of refrigerant, and at a contant condenation temperature t 70 C. 36

14 Iue 5, Volume 6, 01 Table : Data for the calculation of the economic analyi of different implementation of the heat pump Implementation of the heat pump ETČ ETČ DTČR DTČP DTČP Refrigerant R-600a R-600a/R-90 R-600a R-717/R-600a R-407c/R-600a Temperature of boiling (t U) [ C] 0,0 0,0 0,0 0,0 0,0 Temperature of condenation(t ) [ C] 70,0 70,0 70,0 70,0 70,0 Heat flow of the heat pump (Φ TČ) [kw] 88,6 585,36 473,0 674, 631,1 Compreor trength (P ) [kw] 75,3 116,0 111,1 161,6 116,9 Operation time per day (t 1) [h/day] Operation time per year (t ) [day/year] Heat pump lifetime (N) [year] Own fund (C o) [EUR] Cot of the heat pump (C TČ) [EUR] Electric power cot (C E) [EUR/kWh] 0,07 0,07 0,07 0,07 0,07 Cot of heat for heating (C T) [EUR/kWh] 0,035 0,035 0,035 0,035 0,035 Dicount rate (r) [%] 7,00 7,00 7,00 7,00 7,00 Inflation rate (r ) [%] 1,0 1,0 1,0 1,0 1,0 tage i economically utifiable at the boiling temperature of t U 10 C. In the firt tage, the type of the compreor which i recommended i WA or WBH, in the econd tage only WBH. The heat flow of the two-tage heat pump with a heat exchanger in that cae i 489,4 kw, the total conumption of energy for the operation of both compreor i 113,9 kw. The uggeted implementation of the heat pump i one of the method for maximum utilization of low -temperature ource of energy for heating of reidential unit. Fig. 1: NPV depending on boiling temperature of the refrigerant IV. CONCLUSION With a two-tage heat pump with a flah expander, the bet operating condition are with the refrigerant R-600a. For the propoed implementation of the heat pump, the ratio between the preure of the compreed refrigerant decreae. Boiling of the refrigerant at the temperature of t U 0 C i economical. The condenation temperature i t 70 C. Then the heat flow i 408,0 kw, the conumption of energy for the operation of both compreor i 111,1 kw. It i uggeted to ue the ame compreor in the firt and the econd tage a with the one-tage pump. The two-tage heat pump with a heat exchanger conit of two one-tage pump. It advantage i that we can ue, in each tage, different refrigerant. The analyi of thermo-phyical propertie of refrigerant etablihed that the operation of a two-tage heat pump with a heat exchanger i the cheapet uing refrigerant R-407c in the firt tage and R-600a in the econd tage of the heat pump. If the condenation temperature of the refrigerant in the econd tage i t 70 C, the operation of the heat pump of the firt REFERENCES [1] J. rope, D. Doberek, D. Goricanec, Economic evaluation of poible ue of heat of flue gae in a heating plant. WSEAS tran. heat ma tranf., 006, vol. 1, i. 1, pp [] E. Berglez-Matavž, J. rope, D. Goricanec, Techno-economic comparion of the efficiency of variou energy ource heating ytem. WSEAS tranaction on power ytem, 007, vol., i. 1, pp [3] D. Doberek, D. Goricanec, J. rope, Economic analyi of energy aving by uing rotary heat regenerator in ventilating ytem. IASME Tran., Nov. 005, vol., i. 9, pp [4] Heat pump, Mayekawa, Mycom, [5] D. Goricanec, J. rope, etc., High temperature heat pump for exploitation of low temperature geothermal ource, 010 [6] B. ulcar, D. Goricanec, J. rope, Economy of replacing a refrigerant in a cooling ytem for preparing chilled water, International Journal of Refrigeration 33, 010, pp [7] IBRAHIM DINCER, Refrigeration Sytem and application, 16, 003. [8] [9] [10] MYCOM EUROPE, Reciprocating Compreor WBH Serie, Firt Edition, March 006. [11] Zoran Rant, Termodinamic, Mechanical faculty. Univerity of Lublana, Lublana, 000. [1] Recknagel Sprenger, Book for heating and climatiation, Beograd, 006. [13] Z. Črepinšek, D. Goričanec, J. rope,. Comparion of the performance of aborption refrigeration cycle. WSEAS tran. heat ma tranf., vol. 4, i. 1, 009, pp [14] B. ulcar, D. Goricanec, J. rope, Economy of exploiting heat from low-temperature geothermal ource uing a heat pump, Energy and Building 40 (008) [15] urtz, Ruth, Handbook of engineerin economic, Guide for engineer, technician cientit, and manager, McGrawe-Hill,