WASTE HEAT RECOVERY FROM AIR TURBO COMPRESSORS IN MINING: A COMPARATIVE STUDY BETWEEN COGENERATION AND HEAT PUMP TECHNOLOGIES

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1 WASTE HEAT RECOVERY FROM AIR TURBO COMPRESSORS IN MINING: A COMPARATIVE STUDY BETWEEN COGENERATION AND HEAT PUMP TECHNOLOGIES Raduk M. W., Natonal Mnng Unversty, Ukrane Abstract The possblty of cogeneraton and heat pump technologes mplementaton for waste heat recovery from mnng turbo-compressors has been nvestgated. It has been calculated that heat pump mplementaton for turbo-compressor heat recovery enables a reducton of compressed ar producton costs by half. Ths decrease n costs lessens as coolng water nlet temperature ncreases. It has been found that nstead of burnng fossl fuel n coal-burned bolers to meet mnes needs for thermal energy, t s possble to recover compressor statons waste heat by usng heat pumps. They can produce MW of useful heat n the form of hot water at a temperature of 50 C from a sngle 1.5 MW turbocompressor, thus reducng the fossl fuel consumpton n mnes consderably. Cogeneraton technology mplementaton for compressor waste heat recovery has shown that maxmum power generaton (up to 75 kw) can be acheved, dependng on the compressed ar temperature leavng the recovery sectons of the ar coolers (ntercoolng temperature). It can reduce turbo-compressor energy consumpton by up to 5 % and make t possble to save kwh of electrcty annually from a sngle 1.5 MW turbo-compressor. Introducton Ar compressor statons whch produce compressed ar for mnng machnery operaton and dfferent technologcal processes are one of the man energy consumers n coal and ron mnes. Therefore, the effcency of mnng greatly depends on the effcent producton, transmsson and consumpton of compressed ar. It s known that along wth the mechancal work produced, mnng ar compressors release large amounts of energy n the form of waste heat. In a typcal compressor coolng system, all heat produced as a result of a compressor operaton s lost to the envronment, not only lowerng the energy effcency but also causng the envronmental problems such as global warmng and heat polluton. Snce durng the compressor operaton the amount of waste heat produced s almost the same as the amount of electrcty consumed, t s reasonable to try to recover ths waste heat for power generaton or useful heat producton. Moreover, snce a great amount of fossl fuel (bascally coal) n Ukranan mnes s burned annually n bolers for heatng or hot water supply, any measures for the reducton of fossl fuel consumpton could brng economc and envronmental benefts, ncreasng the effcency of mnng. That s why waste heat recovery from compressor statons can ncrease the effcency of compressed ar producton, as well as that of mnng overall. In ths paper cogeneraton and heat pump technologes have been nvestgated to recover waste heat from ar compressor statons. Snce the temperature of compressed ar after compressor stages can reach C, t would be nterestng to know how much electrcty can be generated, usng cogeneraton technology based on the organc Rankne cycle (ORC). As the temperature of the coolng water leavng the compressor coolng system can reach C, t s possble to ntroduce heat pumps that could meet the mnes needs for hot water. The man objectve of ths study s to analyse and compare two technologes that can be mplemented for waste heat recovery from mnng compressor statons: cogeneraton technology based on the ORC and heat pump technology. Cogeneraton technology In ths secton the analyss of cogeneraton technology mplementaton for waste heat recovery from a mnng turbocompressor s carred out. A combned power and heatng cycle s proposed, to produce both electrcal and thermal power by utlzng waste heat from a turbo-compressor K The man objectve of the analyss s to determne the condtons under whch the maxmum power output s reached. Turbo-compressors K 250 and K 500 are the most wde-used compressors n Ukranan mnes. They are smlar n aerodynamc and desgn. These compressors are 6-stage centrfugal devces, combned n 3 sectons of uncooled stages (hereafter referred to as uncooled sectons). To cool down the compressed ar produced, there are 2 nter-stage coolers and one after-cooler. The compressors are equpped wth synchronous 1600 and 3000 kw motors, respectvely. The atmospherc ar s compressed by the turbo-compressor and cooled n the ar coolers. In order to ncrease the temperature potental of the waste heat recovered, and to keep the outlet temperature of compressed ar wthn a requred range, the mplementaton of new types of ar coolers s suggested. The new types of proposed ar coolers consst of two sectons: the frst secton s for waste heat recovery and the second one s for the further coolng of the compressed ar to the requred temperature. To convert the waste heat nto power, the ORC s used. There are several advantages n usng ORC to recover low grade waste heat, ncludng smaller sze, envronmental soundness and great flexblty. The man advantage of the ORC s ts superor performance n recoverng low temperature waste heat [1]. The ORC has receved ncreasng attentons for generatng power from dfferent heat sources such as geothermal, solar, waste heat of nternal combuston engnes and gas turbnes [2 5].

2 Nomenclature Acronyms Symbols AC after-cooler c p sobarc specfc heat (kj/kg*k) C condenser E d. tot total exergy losses (kw) CHP compressor heat pump E n. tot total nput exergy (kw) Ev evaporator polytropc exponent G generator L comp specfc compresson work (kj/kg) HP heat pump m ar ar mass flow rate (kg/s) HWSS hot water supply system m wf workng flud mass flow rate (kg/s) IC nter-stage cooler p n uncooled secton nlet pressure (Pa) P pump p out uncooled secton outlet pressure (Pa) PC heat pump s compressor t h2 hot water outlet temperature ( C) T turbne t nt ntercoolng temperature ( C) Th throttle valve t n uncooled secton nlet ar temperature ( C) WH water heater t out uncooled secton outlet ar temperature ( C) Q total heat transferred from the compressed ar (kw) Greeks Q cond heat removed from the condenser (kw) T temperature dfference (K) Q HW heat transferred to hot water supply system (kw) gen electrc generator effcency Q n heat pump condenser heat load (kw) s. p pump sentropc effcency Q out heat pump evaporator heat load (kw) s.t turbne sentropc effcency Q R heat transferred from the compressed ar (kw) mech.t turbne mechancal effcency W net. el net power output (kw) mot. p motor effcency W p Q total heat recovery effcency W t pump nput power (kw) turbne output power (kw) Q R heat recovery effcency x vapor extracton rate A schematc representaton of cogeneraton waste heat recovery system s shown n Fg. 1. Fg. 1. Schematc dagram of cogeneraton waste heat recovery system. A certan challenge s to choose an approprate organc workng flud whch can meet all requrements and constrants mposed on the system. The workng flud should exhbt chemcal stablty at the operatng pressure and temperature, envronmental frendlness and low toxcty. In addton, t should be non-corrosve, non-flammable, be non-auto-gntng, as well as havng good materal compatblty [6]. Another characterstc that must be consdered durng the selecton of the organc workng flud s ts shape of the saturated curve. Wth respect to the slope of the

3 saturated curve n the T s dagram, all organc workng fluds are dvded nto three groups: dry fluds have a postve slope; wet fluds have a negatve slope; whle sentropc fluds have an nfntely large slope. Bascally, dry and sentropc fluds are more preferable, snce they do not condense when the flud goes through the turbne. The am s to extend turbne blade servce lfe by preventng ther damage due to workng flud condensaton. Snce one of the man objectves of ths study s to determne the condtons under whch the hghest value of power s obtaned, t s desrable to choose the workng flud whose crtcal temperature s slghtly hgher than the temperature of the upper heat reservor. In ths case, the flud evaporaton enthalpy s mnmal, and the workng flud flow rate s maxmum, whch nfluences the hghest value of power output [7]. In vew of the aforesad, dry workng fluds such as butane, pentane, sobutene, neo-pentane, R142b, R236ea and R245fa, whose crtcal temperatures are slghtly above the outlet temperature of compressed ar comng from the uncooled sectons of the turbo-compressor have been chosen. Table 1 Crtcal parameters and envronmental propertes of workng fluds R142b R236ea R245fa butane pentane neo-pentane so-butane Crtcal temperature ( C) Crtcal pressure (MPa) GWP Bass for GWP = 100 years, GWP of CO 2 = 1. Fg. 2 shows the temperature enthalpy dagram of the combned power and heatng cycle. The cogeneraton waste heat recovery system can run n two operatng modes: a power mode, when only electrcal energy s generated, and a heatng mode, when more heat energy rather than electrcal energy s generated. Fg. 2. Schematc T s dagram of combned ORC. When the proposed cogeneraton system runs n the power mode, the workng flud leaves the condenser as saturated lqud (state pont 3). Then, t s compressed by a lqud pump to the sub-crtcal pressure (state pont 4). The workng flud s heated and boled n the evaporator (process 4 5 1) untl t becomes saturated vapor (state pont 6) or slghtly superheated vapor (state pont 1). The saturated or slghtly superheated vapor flows nto the turbne and s expanded to the condensng pressure (state pont 2). Two phase flud passes through the condenser, where heat s removed untl t becomes a saturated lqud (state pont 3). In order to recover heat more completely and more flexbly, the system can run n the heatng mode. In ths case some amount of vapor s extracted from the turbne and drected to the water heater. The condensaton temperature of the workng flud (7a) n the water heater s determned by the extracton pressure. Saturated vapor s pumped to the heat-recovery sectons of the ar coolers. When there s no need for hot water and there s zero vapor extracton, the system works n power mode ( ), generatng maxmum electrc power. The man parameters that nfluence the performance of the combned cycle are turbne nlet temperature and pressure, vapor extracton rate (a porton of the vapor extracted from the turbne for the water heater), the temperature of the compressed ar leavng the heat-recovery sectons of the ar coolers (ntercoolng temperature). These parameters effect the amount of waste heat recovered, power generated, heat recovery effcency (the effcency of waste heat recovery n the heat-recovery sectons of the ar coolers), total heat recovery effcency (the effcency of waste heat recovery n both sectons of the ar coolers) and exergy effcency.

4 Turbo-compressor performance modelng As ntal data, an aerodynamc compressor performance curve [8], and ar cooler characterstcs [9], are used. To recalculate the compressor performance curve under new temperature condtons (whch dffer from the rated ones), the mathematcal model [10] s used. Ths model s based on the assumpton that the compresson work of the uncooled sectons remans equal at both rated and actual modes, f volume ar flow rate also remans constant. In complance wth ths assumpton, pressure rato for an -uncooled secton at a gven nlet ar temperature s expressed as: plk k1 k1 T n. r plk 1 1 a r (1) T n. a where temperature, Т n. r s rated absolute nlet ar temperature for an -uncooled secton, n a r uncooled secton, Т. s actual absolute nlet ar s rated pressure rato for an -uncooled secton, V s volumetrc flow at the entrance to the - pl s polytropc effcency of an -uncooled secton, k s a poltropc compresson exponent. If the compressor performance curve s gven as f V and T f V the -uncooled secton are expressed as:, outlet pressure and temperature for pout p n Tout Tn a T (2) (3) Energy and exergy analyss of the cogeneraton waste heat recovery system The total power of the pumps h9s h3 h4s h10 s. p (4) h h h h W p m h h )(1 x) ( h )] (5) wf [( h10 The turbne power h1 h2 s. t (6) h h 1 2s W t mwf [( h1 h7 ) ( h7 h2 )(1 x)] (7) The condenser heat rate Q cond m ( h h3)(1 ) (8) wf 2 x The total heat transferred from the compressed ar Q m ar c p t o ut1 tout2 tout3 tn2 tn3 tout (9) The heat transferred from the compressed ar n the heat-recovery sectons Q R marc p t o ut t nt = 1, 2, 3 (10) 3 Q R QR (11) 1

5 The heat transferred to the hot water supply system Q HW m ( h 7 h8 ) (12) wf The net power output W net.el Wp Wt mech.t gen (13) mech. p mot. p Electrcal effcency Wnet.el el QR (14) Heat recovery effcency Wel QHW Q R QR (15) Total heat recovery effcency Wnet.el Q HW Q (16) Q The exergy of the state pont s expressed as: E m[( h h0 ) T0 ( s s0 )] (17) The exergy losses of each component can be found from an exergy balance. d En Eout E W (18) The overall exergy effcency s defned as: Ed. tot ex 1 (19) E n. tot For the performance smulaton of the cogeneraton system, assumptons are made as follows: 1) Ppe pressure drop and heat loss to the envronment are neglgble. 2) The condensng temperature s 30 C. The coolng water nlet temperature s assumed to be 20 C 3) Temperature dfference n the condenser and pnch pont temperature dfference are consdered to be 10 C and 5 C respectvely. 4) Isentropc and mechancal effcences for the turbne are assumed to be 0.8 and 0.97, respectvely; generator effcency s Isentropc and mechancal effcences for the pumps are assumed to be 0.7 and 0.97, respectvely; motor effcency s The thermodynamc propertes of the workng fluds were calculated by REFPROP 7.0 [11], developed by the Natonal Insttute of Standards and Technology of the Unted States. The ORC smulaton was performed usng a smulaton program wrtten by the author wth MathLab. As a result of the turbo-compressor performance smulaton, t was calculated that under the gven condtons the temperature of compressed ar leavng the uncooled sectons are 122, 132, 124 C respectvely, and ts mass flow rate s kg/s. Wth these values as nput data for waste heat recovery system smulaton, the followng results have been obtaned. The nfluence of ntercoolng temperature on power output and waste heat recovery s show on the dagrams of Fg 3, 4.

6 Fg. 3. Varaton of net power output wth varous ntercoolng temperature. Fg. 4. Varaton of dfferent effcences wth ntercoolng temperature. Fg. 3 shows the varaton of net power output wth the varous compressed ar temperatures leavng the heatrecovery sectons. The power output s the product of mass flow rate of the workng flud and the enthalpy drop. As the ntercoolng temperature ncreases, the mass flow rate of the workng flud decreases and the enthalpy of the nlet flud to the turbne ncreases. As can be seen from the graph n Fg. 3, there s an optmum ntercoolng temperature for maxmum power. The hghest power s produced by R236ea at ntercoolng temperature of 60 C that almost concdes wth maxmum exergy effcency. However, the electrcal effcency at the optmum ntercoolng temperature s not at maxmum. The electrcal effcency ncreases wth the temperature; at hgher temperature the workng flud absorbs less heat and produces less power but has hgher electrcal effcency Fg. 4. Fg. 5 llustrates the varaton of electrcal effcency and the heat recovery effcency wth varous vapor extracton rate when the system runs n heatng mode. Fg. 5. Varaton of electrcal effcency and heat recovery effcency wth the varous vapor extracton rate. Fg. 5 shows the varaton of electrcal effcency and heat recovery effcency wth the varous vapor extracton rate. It can be observed that electrcal effcency slghtly decreases as vapor extracton rate ncreases. However, heat recovery effcency ncreases dramatcally due to heat usage for hot water supply.

7 Compared wth other workng fluds, R236ea produces the hghest power from waste heat delvered from turbocompressor K There s an optmum ntercoolng temperature whch les wthn the range (60 65 C) for maxmum power output. When the system runs n power mode, the maxmum power output s 75 kw; heat recovery effcency and total heat recovery effcency are and 0.053, respectvely. When the system runs n heatng mode, 891 kw of heat and 38 kw of electrc power can be generated; heat recovery effcency and total heat recovery effcency are 0.99 and 0.65, respectvely. R236ea shows a better performance over other workng fluds, but not notably. Therefore, other factors, such as turbne sze and heat exchanger surface area combned wth economc modelng should be taken nto account when decdng what workng flud to choose for waste heat recovery from mnng turbo-compressors. Cogeneraton technology mplementaton for waste heat recovery from a sngle 1.5 MW turbo-compressor can reduce compressor energy consumpton by up to 5 % and make t possble to save kwh of electrcty and 6, 42*10 6 kwh of heat energy annually, whch s equal to about 925 tons of coal. Heat pump technology Snce the amount of electrcty that can be generated from the gven heat source s sgnfcantly lower than the amount of heat energy, t would be reasonable to determne the feasblty of only heat energy producton by usng heat pumps. In ths secton the analyss of heat pump technology mplementaton for waste heat recovery s carred out. Fg. 6 shows the dagram of heat pump mplementaton for waste heat recovery from a turbo-compressor. Рис. 6. Schematc dagram of heat pump mplementaton for waste heat recovery. Accordng to the dagram n Fg. 6 the water n the turbo-compressor coolng system crculates n a closed loop. It s heated n the turbo-compressor ar coolers and cooled down n the heat pump evaporator. In a typcal compressor coolng system the temperature of the nlet coolng water depends on the ambent ar temperature. In case of heat pump technology mplementaton, ths temperature can be consderably ncreased or decreased wth respect to the ambent temperature. The temperature of the nlet coolng water greatly nfluences the cost of compressed ar producton and affects the operaton of the heat pump. However, the nature of ths nfluence s ambguous. On the one hand, the ncrease of coolng water nlet temperature leads to ncreasng COP. But on the other hand, when coolng water nlet temperature ncreases, compressed ar outlet temperature and (as a result compresson work) also ncrease. Ths results n turbocompressor performance decrease. Therefore, one of the man objectves of ths research s to determne the temperature mode of coolng water crculaton loop at whch the effcency of Turbo-compressor Heat Pump (THP) system s hghest. The producton cost of 1 м 3 of compressed ar C ar. hp and beneft B ganed from the hot water produced by the heat pump have been chosen as the effcency ndex of THP system performance. Whle calculatng the producton cost of 1 м 3 of compressed ar, the followng expendtures have been taken nto consderaton: the cost of electrcty consumed by the turbo-compressor C el. tc, and by the heat pump C el. hp, heat pump deprecaton cost C dc. hp and the cost of heat energy C h. C Cel. tc Сel. hp Сdc. hp Сh. (20) V ar hp year

8 where V year s the amount of compressed ar produced annually. The beneft ganed from heat pump technology mplementaton for waste heat recovery has been estmated as a dfference between the cost of compressed ar produced by the turbo-compressor wthout heat pumps C., and after heat pump technology mplementaton B C ar. comp C ar. hp V year C ar. hp. ar comp (21) When the pressure on the nlet and outlet sde of the compressor, nlet ar temperature, coolng water flow rate and ts temperature on the nlet sde of the ar coolers t w1 are gven, the problem of turbo-compressor mode calculaton comes n determnng the coolng water outlet temperature t w2 and the temperature of the compressed ar on the outlet sde of the ar coolers, the dscharge temperature and pressure of the compressed ar on the outlet sde of the uncooled sectons. Also to be consdered are the amount of heat rejected n the ar coolers Q (whch s heat pump evaporator heat load) and power consumed by the turbo-compressor s motor W comp. The values t w2 and Q n, obtaned through compressor operaton mode calculatons, are nput data for heat pump mode calculatons and energy analyss. The objectve of the heat pump operaton mode calculatons when t w1, t w2, Q n, t h2 are gven nvolves the calculaton of heat pump thermodynamc cycle parameters, heat pump electrc power nput W hp and ts thermal power output Q out (whch s heat pump condenser heat load) as well as COP. For the CPH system performance smulaton, the followng assumptons are made: 1) Coolng water nlet temperatures t w1 vared between 5 and 30 С. 2) Atmospherc pressure s assumed to be 0,1 MPa, and ambent temperature s 15 С. 3) Dscharge absolute pressure s 0.9 MPa. 4) Coolng water mass flow rate s consdered to be 11.0 kg/s. 5) Hot water outlet temperature s consdered to be 50 С. 6) Cost of heat energy s assumed to be UAH/MJ; heat pump captal costs are consdered to be 820$/kW. 7) Temperature dfferences n the evaporator and n the condenser are assumed to be 5 С. 8) Isentropc, mechancal and motor effcences for the heat pump are assumed to be 0.75, 0.97 and 0.94 respectvely. 9) All heat exchangers n the cycle (evaporator and condenser) are consdered to be shell-and-tube counter flow heat exchangers. 10) As workng flud, refrgerant R134a s used. The coolng system of a turbo-compressor K has been chosen as a source of low-grade heat for the heat pump. The nfluence of the coolng water temperature w1 s shown on Fg Twenty-four-hours mode of operaton has been chosen for the heat pump. n t on the energy and cost parameters of THP system performance Fg. 7. Influence of coolng water nlet temperature on energy parameters of THP system.

9 Fg. 7 shows the nfluence of t w1 on compressor performance and energy parameters of the system, such as: condenser heat load consumpton Q out, evaporator heat load Q n, heat pump W comp. From these results, t s evdent that as w1 W hp and turbo-compressor electrc power t decreases, all energy parameters ncrease. The ncrease of evaporator heat load Q n can be explaned by a better coolng of the compressed ar and an ncrease n compressor performance. Ths, n turn, results n an ncrease n both the condenser heat load Q out and electrc power W hp consumpton. The ncrease of W hp s also due to the decrease of COP as temperature t w1 decreases. Fg. 8. Influence of coolng water nlet temperature on COP and compresson work. Fg. 8 llustrates that as t w1 ncreases, COP also ncreases. It s shown that at the pont when temperature t w1 s 30 С and hot water outlet temperature s 50 С, COP can reach 5.7 whch s consdered to be a promsng value. Fg. 7 t decreases, both compressor performance V and as a result electrc power 8 show that n spte of the fact that as w1 consumed by the turbo-compressor W comp ncrease, specfc compresson work L comp decreases. For a gven turbocompressor, the amount of waste heat recovered amounts to kw. Heat pump heatng capacty s kw. Fg. 9 shows the nfluence of w1 C ar. hp and beneft B ganed from the thermal energy produced. t on some economc parameters such as producton cost of 1 m 3 of compressed ar Рис. 9. Influence of coolng water nlet temperature on compressed ar producton cost and beneft ganed from thermal energy generaton. The results show that as t w1 ncreases, the producton cost of 1 m 3 of compressed ar C ar. hp decreases and beneft B ganed from thermal energy ncreases. From these results, t s evdent that turbo-compressor specfc compresson work

10 ncreases as t w1 ncreases. However, the economc beneft ncreases, whch contrbutes to a decrease n producton cost of 1 m 3 of compressed ar. It should be stressed that the ncrease of t w1 and as a result COP, can result n a decrease n turbo-compressor performance that can nfluence the normal work of machnery that utlze compressed ar. Conclusons Cogeneraton technology mplementaton for waste heat recovery from a sngle 1.5 MW turbo-compressor can reduce compressor energy consumpton by up to 5 %. It makes t possble to save kwh of electrcty and 6, 42*10 6 kwh of heat energy annually, whch s equal to about 925 tons of coal. In comparson, heat pump technology mplementaton can generate kw of thermal power, whch s equal to 1660 tons of coal. In addton, the producton costs of compressed ar can be almost halved. The results of THP system smulaton show that f compressed ar and hot water are consdered to have equal sgnfcance, certan recommendatons can be made. It can be recommended to cool down the compressed ar at a slghtly hgher temperature, when coolng water nlet temperature s С just lke n a typcal coolng system. However, snce the man am of a compressor s to produce compressed ar, preference should be gven to those operaton modes of the system under whch the coolng water nlet temperature would not exceed the temperature specfed for typcal turbo-compressor coolng systems. In ths case, however, an economc beneft ganed from heat pump technology mplementaton for waste heat recovery from turbo-compressors stll remans consderable. Another mportant advantage of the techncal soluton proposed, s a closed crcut of the coolng water system. Ths prevents heat exchange surfaces of ar coolers from foulng and contrbutes to an ncrease n nterrepar mantenance perod of the ar coolers, as well as less spendng on repars. Moreover, waste heat recovery from compressor statons leads to the reducton of boler plant emssons whle burnng fossl fuel to meet the mne s needs for heatng and hot water. In addton, envronmental heat polluton s also reduced sgnfcantly. Although heat pump technology mplementaton for waste heat recovery from turbo-compressors s more benefcal than cogeneraton n terms of coal saved, cogeneraton technology s more flexble. It makes t possble to generate both heat and electrcty accordng to the mne s needs. For nstance, n summer tme when demand for heat energy s low, the cogeneraton waste heat recovery system can work n power mode. In wnter, when demand for heat energy ncreases, the cogeneraton system can run n heatng mode, generatng mostly heat energy. References 1) Tamamoto T, Furuhata T, Ara N, Mor K. Desgn and testng of the organc Rankne cycle. Energy 2001;26: ) H.D. Madhawa Hettarachch, Mhajlo Golubovc, Wllam M. Worek, Yasuyuk Ikegam Optmum desgn crtera for n Organc Rankne cycle usng low-temperature geothermal heat sources. Energy 2007;32(9): ) Agustín M. Delgado-Torres, Lourdes García-Rodríguez. Analyss and optmzaton of the low-temperature solar organc Rankne cycle (ORC). Energy Converson and Management 2010;51(12): ) IacocpoVaja, Agostno Gambarotta. Internal combuston engne (ICE) bottomng wth organc Rankne cycle. Energy 2009;35(2): ) Costante Invernzz, Paolo Iora, Paolo Slva. Bottomng organc Rankne cyle for mcro-gas turbnes. Appled Thermal Engneerng 2007;27(1): ) V. Mazza, A. Mazza. Unconventonal workng fluds n organc Rankne cycles for waste energy recovery systems. Appled Thermal Engneerng 2001;21(3): ) Aleksandra Borsukewcz-Gozdur, Wladyslaw Nowak. Comparatve analyss of natural and synthetc refrgerants n applcaton to low temperature Clausus-Rankne cycle. Energy 2007;32(4): ) Цейтлин Ю. А., Мурзин В. А. Пневматические установки шахт. М.: Недра, ) Берман Я.А, Маньковский О.Н., Марр Ю.Н., Рафалович А.П. Системы охлаждения компрессорных установок. Л.: Машиностроение, ) Рис В.Ф. Центробежные компрессорные машины. Л.: Машиностроение, ) REFPROP Verson 7.0, NIST Standard Reference Database 23, 2002.