Received: 19 October 2017; Accepted: 26 November 2017; Published: 3 December 2017

Transcription

1 energies Article Performance Prediction Centrifugal Compressor for Drop-In Testing Using Low Global Warming Potential Alternative Refrigerants Performance Test Codes Joo Hoon Park 1,2, Youhwan Shin 2, * Jin Taek Chung 3, * 1 Department Mechanical Engineering, Graduate School Korea University, Seoul, Korea; joohoon82@korea.ac.kr 2 Centre for Urban Energy Research, Korea Institute Science Technology, Seoul, Korea 3 Department Mechanical Engineering, Korea University, Seoul, Korea * Correspondence: yhshin@kist.re.kr (Y.S.); jchung@korea.ac.kr (J.T.C.); Tel.: (Y.S.); (J.T.C.) Received: 19 October 2017; Accepted: 26 November 2017; Published: 3 December 2017 Abstract: As environmental regulations to stall global warming are strengned around world, studies using newly developed low global warming potential (GWP) alternative refrigerants are increasing. In this study, substitute refrigerants, R-1234ze (E) R-1233zd (E), were used in centrifugal compressor an R-134a 2-stage centrifugal chiller with a fixed rotational speed. Performance predictions rmodynamic analyses centrifugal compressor for drop-in testing were performed. A performance prediction method based on existing ASME PTC-10 performance test code was proposed. The proposed method yielded expected area point centrifugal compressor with alternative refrigerants. The rmodynamic performance first second stages centrifugal compressor was calculated as polytropic state. To verify suitability proposed method, drop-in test results two alternative refrigerants were compared. The predicted based on permissible deviation ASME PTC-10 confirmed that temperature difference was very small at same efficiency. Because drop-in test R-1234ze (E) was performed within expected, centrifugal compressor using R-1234ze (E) is considered well predicted. However, predictions point R-1233zd (E) were lower than those drop-in test. The proposed performance prediction method will assist in understing rmodynamic performance at expected point area a centrifugal compressor using alternative gases based on limited design structure information. Keywords: ; point; centrifugal compressor; low-gwp alternative refrigerant; R-1234ze (E); R-1233zd (E); drop-in test; performance test code 1. Introduction At present, hydrluorocarbon (HFC)-series refrigerants are main working refrigerants used in turbo chillers. Since such refrigerants have a high global warming potential (GWP), ir use is being phased out worldwide. Refrigerants that have little or no effect on global warming have been continuously developed to replace existing high-gwp refrigerants. Low-GWP refrigerants include hydrocarbon-based refrigerants, natural refrigerants, hydrluoroolefin (HFO)-based refrigerants. Of se, HFO-based refrigerants, R-1234yf, R-1234ze (E), R-1233zd (E), R-1336mzz, are attracting attention as new alternative refrigerants owing to ir low-gwp characteristics compatibility with existing turbo chiller refrigerants. These alternative refrigerants have been subjected to drop-in Energies 2017, 10, 2043; doi: /en

2 Energies 2017, 10, tests in existing facilities for applicability rmodynamic evaluation. In this regard, studies using HFO-based low-gwp alternative refrigerants have been conducted in various ways. Wang et al. [1] introduced Air-Conditioning, Heating, Refrigeration Institute (AHRI) s Low-GWP Alternative Refrigerant Evaluation Program (Low-GWP AREP) outlined overall scope procedures program. In addition, for low-gwp cidate refrigerants, y provided cycle count up to current state actual test result. Spatz et al. [2] analysed oretical performance several low-gwp refrigerants in a volumetric chiller centrifugal chiller used for air conditioning. The centrifugal chiller is divided into a low-pressure chiller using R-123 a medium-pressure chiller using R-134a. The rotational speed, impeller outlet diameter, coefficient performance (COP) compressor were compared to those conventional high-gwp refrigerants, results were analysed. This study showed that se refrigerants could be applied to existing devices without significant hardware modifications. Ueda et al. [3] performed a drop-in test R-1234ze (E) alternative refrigerant in an R-134a centrifugal chiller. They analysed compatibility products related to lubrication performance chiller via full- partial-load refrigerator tests. The results drop-in test under full load conditions showed that cooling capacity COP were reduced by approximately 29% 3%, respectively, at same inlet volumetric flow rate. Brasz [4] predicted capacity performance a centrifugal compressor using an ideal cycle a rmodynamic property calculation for R-1234ze (E). In study, R-1234ze (E) was injected into an existing R-134a centrifugal compressor to compare aerodynamic performance with rotational speed. The author found that, when using alternative refrigerants, efficiency chiller increased because low rotational speed viscous loss compared to conventional centrifugal compressors, which are larger than centrifugal compressor used in conventional chiller. However, it was suggested that a condenser redesign is necessary to prevent a pressure drop from condenser. Pearson [5] analysed method characteristics R-134a for alternative refrigerants explained primary issues to consider in alternative-refrigerant centrifugal compressor chiller design. The performance developed centrifugal compressor was compared analysed with COP R-134a, R-1234ze (E) was found to be a viable alternative to existing R-134a centrifugal compressor. Mota-Babiloni et al. [6] performed a drop-in test two alternative refrigerants (R-1234yf R-1234ze (E)) in an R-134a vapor compression plant. It was confirmed that volume efficiency, cooling capacity, COP were decreased compared with those R-134a, use an internal heat exchanger could reduce COP difference between alternative refrigerants. Wu Thilges [7] experimentally analysed performance one centrifugal compressor with three alternative refrigerants for various types refrigerators with R-123 as working fluid. The effects isentropic exponent, Reynolds number, blade Mach number, impeller diameter, rotational speed, etc. were studied. Janković et al. [8] analysed performance system heat transfer characteristics a condenser when R-1234yf R-1234ze (E) were applied to simulation model an R-134a small-power refrigeration system, considering actual dimensions. The authors concluded that R-1234ze (E) could perform better than R-1234yf as an alternative refrigerant for R-134a if an overridden compressor with a cooling system could be used. Meng et al. [9] showed oretical performance three mixed gases consisting R-1234ze (E), R-152a, R-1234ze (E) mixed with R-152a, respectively, in drop-in tests. The COP R-1234ze (E) mixture was similar to that R-134a, but cooling capacity was significantly lower. They also showed that a refrigerant composed R-1234ze (E) R-152a with a mixing ratio 50:50 would be suitable as an alternative refrigerant for R-134a. Sethi et al. [10] analysed oretical performance when replacing R-447B R-452B with a domestic reversible heat pump a direct-expansion air-cooled chiller using R-410A. For R-410A

3 Energies 2017, 10, positive displacement chiller, expected conditions components were analysed with alternative refrigerants. In addition, size performance centrifugal compressors based on alternative refrigerants such as R-515A, R-1234ze (E), R-1234yf were oretically analysed for use in R-134a centrifugal chillers. To summarize previous studies to date, studies using low-gwp refrigerants have mainly focused on heat exchangers, heat pumps, rmodynamic cycle analyses. In recent years, some companies have developed centrifugal chillers [11,12] centrifugal compressors [13] that use a low-gwp alternative refrigerant as gas. However, studies on characteristics a centrifugal compressor alone using a low-gwp alternative refrigerant have not yet been published. In particular, when it is difficult to supply alternative refrigerants, a method to predict performance centrifugal compressor in advance would be very useful. In addition, it is necessary to evaluate analyse performance centrifugal compressors when using a substitute gas. In this study, two low-gwp alternative refrigerants, R-1234ze (E) R-1233zd (E), were applied to an R-134a centrifugal compressor with a two-stage turbo chiller. Then, area point centrifugal compressor were predicted via a drop-in test. The existing centrifugal compressor performance test code was improved, expected performance a centrifugal compressor using an alternative refrigerant was calculated rmodynamically analysed. Although R-1233zd (E) is a low-pressure refrigerant, it was used to examine feasibility proposed performance prediction method. 2. Centrifugal Compressor Performance Prediction Method AHRI provides an evaluation program for drop-in testing using low-gwp alternative refrigerants [14]. Based on this program, centrifugal compressor test is applied according to The American Society Mechanical Engineers, Performance Test Code-10 (ASME PTC-10) [15]. ASME PTC-10 provides a performance test procedure permissible deviation required for rmodynamic evaluation centrifugal compressor axial compressor. The ASME PTC-10 performance tests are classified into two types, type 1 type 2, depending on test gas. The type 1 method is used when components design gas test gas are same or similar. The type 2 method is used when components design gas test gas are different. In both types tests, performance centrifugal compressor is compared with design performance, feasibility result is rmodynamically evaluated. However, for each type test result, re is a difference between design performance, type comparison variable, permissible deviation. For a multi-stage compressor, performance is evaluated using information about conditions inlet outlet (inlet duct outlet duct) compressor, impeller outlet diameter at each stage, impeller outlet width at first stage. Here, mass flow rate at inlet outlet should be same. In this study, we used ASME PTC-10 type 2 method for predicting performance improved method to calculate expected expected performance centrifugal compressor using an alternative refrigerant. Performance prediction was conducted for each stage Performance Prediction Procedure 1 shows performance prediction process using an alternative refrigerant. The impeller diameter, impeller outlet width, surface roughness impeller remain constant because centrifugal compressor s shape does not vary with type refrigerant used. The performance prediction process using alternative refrigerants is as follows: STEP-1: Enter information about design conditions. The required information includes name design refrigerant, total pressure total temperature inlet outlet, mass flow rate, rotational speed. STEP-2: Calculate rmodynamic performance based on design conditions. These conditions include total pressure ratio, total temperature ratio, ratio inlet outlet specific volumes,

4 Energies 2017, 10, Energies 2017, 10, isentropic volumes, isentropic polytropic polytropic states, machine states, Mach machine number Mach (defined number in(defined Equation in Equation (3)), machine (3)), Reynolds machine number Reynolds (defined number Equation (defined(4)), Equation so(4)), on. so on. STEP-3: Enter information about expected test conditions. This input information is issimilar to to design conditions. The name selected alternative refrigerant is isentered as as test operation refrigerant. STEP-4: Calculate rmodynamic performance based on expected test conditions. The items to to be be calculated are same as those in STEP-2. STEP-5: Compare following four results obtained from from STEP-2 STEP-2 STEP-4: STEP-4: ratio ratio inlet inlet outlet outlet specific specific volumes, volumes, flow flow coefficient, coefficient, machine machine Mach Mach number, number, machine machine Reynolds Reynolds number. number. The The expected expected operation operation drop-in drop-in test is test calculated is calculated based based on ondesign design result result (STEP-2). (STEP-2). The The allowable allowable deviation deviation in in operation operation is calculated is calculated based based on on type-2 type-2 ASME ASME PTC-10 PTC-10 criteria. criteria. STEP-6: Select Select expected expected point point that meets that meets design design result inresult expected in expected operation operation derived in STEP-5. derived in The STEP-5. expected The expected point is selected point is after selected comparison after comparison with design with results design via results correction via process. correction The outlet process. temperature The outlet temperature pressure pressure selected selected point are reflectedpoint are inlet reflected conditions in inlet conditions next connected stage. next connected stage. 1. Procedure for predicting performance using alternative refrigerants. Table 1 shows permissible deviation required for performance prediction. The Table 1 shows permissible deviation required for performance prediction. The calculation calculation formulas for each comparison variable are as follows: formulas for each comparison variable are as follows: % = = (1) % vr = vr ( t vr = v ) in (1) vr d 60 v out = (2) 2 60, Φ = ṁ ρ = in 2πN, D 3 (2) imp,out (3) =, Mm, = U imp,out =,, (3) a in (4) where, %vr is percentage Rem deviation = U imp,out b imp,out between specific = U imp,out b imp,out volume ratios, vr is specific volume ratio, (4) ν µ in v in Φ is flow coefficient, Mm is machine Mach number, Uimp,out (m/s) is tangential speed at where, impeller % vr outlet, is ain percentage (m/s) is acoustic deviationvelocity, betweenrem specific is volume machine ratios, Reynolds vr is specific number, volume bimp,out is ratio, Φimpeller is flow outlet coefficient, width, νin Mm (m 2 /s) is is machine kinematic Mach viscosity, number, μin U(kg/m s) imp,out (m/s) is isabsolute tangential viscosity, speed at impeller vin (m 3 /kg) outlet, is a in specific (m/s) volume. is acoustic The subscript velocity, in Rem signifies is a machine specific point Reynolds number, inlet duct. b imp,out is

5 Energies 2017, 10, impeller outlet width, ν in (m 2 /s) is kinematic viscosity, µ in (kg/m s) is absolute viscosity, Energies v in (m , 10, /kg) is specific volume. The subscript in signifies a specific point inlet duct. Table 1. Permissible deviations from specified parameters [15]. Table 1. Permissible deviations from specified parameters [15]. Permissible Deviations from Specified Operating Parameters Permissible Deviations from Specified Limit Test Operating Values Parameters as Percentage Design Values Parameter Symbol Limit Test Min. Values as Percentage Max. Design Values Parameter Symbol Specific volume ratio %vr 95% Min. 105% Max. Flow coefficient Φ 96% 104% Specific Machine volume Mach ratio number % vr Mm 95% See 2a 105% Flow coefficient Φ 96% 104% Machine Reynolds number Rem See 2b Machine Mach number Mm See 2a Machine Reynolds number Rem See 2b 2a,b show minimum maximum deviation s for design conditions test conditions, respectively, as permissible machine Mach number machine 2a,b show minimum maximum deviation s for design conditions test Reynolds number. The minimum machine Reynolds number in 2b is greater than 90,000 [15]. conditions, respectively, as permissible machine Mach number machine Reynolds number. The minimum machine Reynolds number in 2b is greater than 90,000 [15]. (a) (b) Allowable diagram: (a) (a) machine Mach number (b) (b) machine Reynolds number adapted from from [15], [15], with with permission from from American Society Mechanical Engineers, In ASME PTC-10, performance a centrifugal compressor is evaluated with polytropic method. Because refrigerant alternative refrigerant are all real gases, method proposed by Schultz [16] is used to calculate polytropic parameters for a real gas.

6 Energies 2017, 10, In ASME PTC-10, performance a centrifugal compressor is evaluated with polytropic method. Because refrigerant alternative refrigerant are all real gases, method proposed by Schultz [16] is used to calculate polytropic parameters for a real gas Calculation Polytropic Exponents The polytropic exponents for a real gas can be calculated using compressibility function proposed by Schultz. The Schultz compressibility functions X Y are: X = T v Y = p v ( ) v 1 (5) T p ( ) v p T Here, Schultz compressibility functions X Y can be calculated by converting volume expansivity calculated using Equation (7) isormal bulk modulus calculated using Equation (8) into Equations (9) (10), respectively: X = T v α = 1 v ( ) [ v 1 = T T p Y = p v ( v p T ( ) v T p ( ) p K T = v (8) v T ( ) ] 1 v 1 = αt 1 (9) v T p ) [ ( ) ] 1 v = p = p (10) v p K T T (6) (7) X m = X in + X out 2 (11) Y m = Y in + Y out 2 where α (1/K) is volume expansivity K T (kpa) is isormal bulk modulus. Since temperature pressure states are different at compressor inlet outlet, compressibility function at compressor inlet outlet are averaged n used. X m Y m, calculated using Equations (11) (12), respectively, are used to calculate polytropic volume exponent, n, polytropic temperature exponent, m. The polytropic exponents are given by Equations (13) (14) [17]. Equation (15) gives isentropic volume exponent (k v ) used to obtain polytropic temperature exponent: 1 + X m n = ( 1 1 k V m = ) ( ηp + X Y 1 m ηp 1 ( )( ) kv Y m 1 1 k V ηp + X m (12) ) (13) (1 + X m ) 2 (14) ( ) k V = ln pout p in ) ln( vin v,is,out (15) where η p is polytropic efficiency v is,out (m 3 /kg) is specific volume at outlet duct in isentropic state.

7 Energies 2017, 10, Calculation Polytropic State In this study, method proposed by ISO 5389 [17] was used to calculate polytropic efficiency a centrifugal compressor. The formula for calculating polytropic efficiency is given by Equation (16) using compressibility functions X Y proposed by Schultz. To evaluate rmodynamic performance centrifugal compressor according to point, we can calculate polytropic work W p (N m/kg) gas power P g (kw) using Equations (17) (18). Equation (19) gives Schultz s work factor, f, for calculating polytropic work [16]: η p = k V (1+X m ) 2 k V Y m 1 1 ln T out T in ln p out p in ( ) n W p = f R Z in T in n 1 X m [ ( pout p in ) ( n 1 n ) ] 1 (16) (17) P g = ( Wp ṁ) η p (18) h f = is,out h [( ) in ] (19) kv k V 1 (p out v is,out p in v in ) where R (N m/kg K) is gas constant, Z in is compressibility factor refrigerant at inlet. duct, m (kg/s) is mass flow rate, h is enthalpy at a specific point Correction Selection Method for Drop-In Test s Expected Operating Point For a single-shaft multi-stage centrifugal compressor connected to return channel, outlet pressure return channel is designed to equalize inlet pressure next stage connected. Therefore, an appropriate predicted point should be selected as entry state next stage among calculated expected operation areas. STEP-6 in 1 is step in which an expected point is selected based on predicted area. In this step, design performance is corrected using results predicted driving. The predicted point drop-in test was selected when polytropic work value calculated correction was equal to polytropic work design performance. If value polytropic work correction result is greater or less than design performance value, outlet temperature pressure predicted point can be adjusted. To correct expected point drop-in test to match design performance, following equations are sequentially calculated: Q corr = Q t N d N t (20). m corr = Q t N ( ) d 1 (v N t v in,corr = v in,d ) in,corr (21) N corr = N t Q d (22) Q t [ (1 ) ( ) η p,corr = 1 ηp,t RAd RA t RA = RB = ( )] RBd RB t [ ] RC ( b imp,out) Rem log( Rem ) log(ε Rem ) RC = Rem (23)

8 Energies 2017, 10, Energies 2017, 10, , =, Rem corr = η (25) p,d (24) where Q (m 3 /s) is volume flow rate at inlet. ηra, p,t RB, RC are machine Reynolds number correction constant constants. The subscript corr ( signifies ) 2 Nd corrected state. W p,corr = W p,t Rem corr (25) N t 2.5. Thermodynamic Properties Refrigerants where Q (m 3 /s) is volume flow rate at inlet. RA, RB, RC are machine Reynolds number The rmodynamic properties all refrigerants used for performance prediction were correction constant constants. The subscript corr signifies corrected state. obtained from NIST REFPROP 9.1 [18]. However, those R-1234ze (E) R-1233zd (E) were 2.5. obtained Thermodynamic from Properties file uploaded Refrigerants to FAQ NIST REFPROP [19]. The refrigerant state equation is Helmholtz equation state recommended by NIST [20]. Microst Excel was used for The rmodynamic properties all refrigerants used for performance prediction were obtained performance prediction, it was possible to acquire state quantity refrigerant in from NIST REFPROP 9.1 [18]. However, those R-1234ze (E) R-1233zd (E) were obtained from conjunction with REFPROP data. file uploaded to FAQ NIST REFPROP [19]. The refrigerant state equation is Helmholtz equation 3. Comparison state recommended Thermodynamic by NIST Characteristics [20]. Microst Excel Low-GWP was used Alternative for performance Refrigerants prediction, it was possible to acquire state quantity refrigerant in conjunction with REFPROP data. 3a,b compare rmodynamic properties R-134a, R-1234ze (E), R-1233zd (E). 3. Comparison 3a shows Thermodynamic saturation curve Characteristics three refrigerants Low-GWPin Alternative a pressure-enthalpy Refrigerants diagram, 3b shows pressure-temperature plot. At same pressure, we can observe that R-1234ze 3a,b compare rmodynamic properties R-134a, R-1234ze (E), R-1233zd (E). (E) is similar to R-134a deviation is not large. However, R-1233zd (E) has a relatively higher 3a shows saturation curve three refrigerants in a pressure-enthalpy diagram, enthalpy higher temperature than R-134a. Although it is a qualitative comparison, R-1234ze (E) 3b shows pressure-temperature plot. At same pressure, we can observe that R-1234ze (E) is considered to be suitable as an alternative refrigerant for R-134a, which is well explained in is similar to R-134a deviation is not large. However, R-1233zd (E) has a relatively higher previous research. Table 2 lists environmental impact (ODP GWP) rmodynamic enthalpy higher temperature than R-134a. Although it is a qualitative comparison, R-1234ze (E) properties for each refrigerant. is considered to be suitable as an alternative refrigerant for R-134a, which is well explained in previous research. Table 2 lists environmental impact (ODP GWP) rmodynamic properties for each refrigerant. (a) 3. Cont.

9 Energies 2017, 10, Energies 2017, 10, (b) Comparison rmodynamic diagrams for for R-134a, R-1234ze (E), (E), R-1233zd (E). (E). (a) (a) Pressure-enthalpy diagram (b) (b) pressure-temperature diagram. Table Table 2. Environmental 2. impact impact rmodynamic properties refrigerants. Refrigerant R-134a R-1234ze (E) R-1233zd (E) Refrigerant R-134a R-1234ze (E) R-1233zd (E) ODP ODP GWP GWP Molar mass (kg/kmol) Molar mass Critical (kg/kmol) temperature (K) Critical temperature (K) Critical pressure (kpa) Critical pressure (kpa) Critical density (kg/m 3 ) Critical density (kg/m 3 ) Description Centrifugal Compressor in a Centrifugal Chiller for a Drop-In Test 4. Description Centrifugal Compressor in a Centrifugal Chiller for a Drop-In Test 4.1. Specifications Operational State 4.1. Specifications Operational State 4a shows R-134a centrifugal chiller (500 RT) for drop-in test using an alternative 4a shows R-134a centrifugal chiller (500 RT) for drop-in test using an alternative refrigerant. 4b shows meridional plane two-stage centrifugal compressor used in refrigerant. 4b shows meridional plane two-stage centrifugal compressor used in chiller. The centrifugal compressor is composed two stages but with a single shaft. There is a chiller. The centrifugal compressor is composed two stages but with a single shaft. There is a return channel between first second stage, economizer is connected before second return channel between first second stage, economizer is connected before second stage s entrance. Therefore, mass flow rate refrigerant entering second stage is larger stage s entrance. Therefore, mass flow rate refrigerant entering second stage is larger than that entering first stage. The first- second-stage diffusers are all vaneless. 5 shows than that entering first stage. The first- second-stage diffusers are all vaneless. 5 shows oretical cycle R-134a centrifugal compressor for a 100% load condition via a oretical cycle R-134a centrifugal compressor for a 100% load condition via pressure-enthalpy diagram. There is no superheat degree at first-stage inlet centrifugal a pressure-enthalpy diagram. There is no superheat degree at first-stage inlet centrifugal compressor, pressure at first-stage outlet second-stage inlet are same. The inlet compressor, pressure at first-stage outlet second-stage inlet are same. The inlet temperature second stage is reduced by approximately 0.3 K from temperature firststage outlet via flow from economizer. temperature second stage is reduced by approximately 0.3 K from temperature first-stage outlet via flow from economizer. The measurement device installation position in performance test centrifugal The measurement device installation position in performance test centrifugal chiller are as follows. A relative pressure transmitter was used for static pressure measurement. The chiller are as follows. A relative pressure transmitter was used for static pressure measurement. accuracy pressure transmitter is 0.1%FS. For temperature measurement, a platinum resistance The accuracy pressure transmitter is 0.1%FS. For temperature measurement, a platinum resistance temperature detector (RTD) with a two-wire-type 1000-Ω sensor was used. The class RTD temperature detector (RTD) with a two-wire-type 1000-Ω sensor was used. The class RTD sensor is A (±(0.15 C 0.002t) C), temperature coefficient resistance is Ω/ C. The sensor is A (±(0.15 C t) C), temperature coefficient resistance is Ω/ C. temperature pressure sensors are installed at first-stage inlet second-stage volute The temperature pressure sensors are installed at first-stage inlet second-stage volute outlet. The measured data were acquired through an in-house system, measured results were outlet. The measured data were acquired through an in-house system, measured results were used to verify performance prediction results. used to verify performance prediction results.

10 Energies 2017, 10, Energies 2017, 10, Energies 2017, 10, (a) (a) (b) 4. Centrifugal chiller its drop-in (b) test using low-gwp alternative refrigerants. (a) Centrifugal chiller for drop-in test (b) meridional plane Centrifugal chiller chiller its drop-in its drop-in test usingtest low-gwp using alternative low-gwp refrigerants. alternative (a) refrigerants. Centrifugal chiller (a) Centrifugal for drop-in chiller test for drop-in (b) meridional test plane. (b) meridional plane. 5. Pressure-enthalpy diagram with oretical operational cycle R-134a centrifugal compressor 5. Pressure-enthalpy at 100% load. diagram with oretical operational cycle R-134a centrifugal compressor at 100% load. 5. Pressure-enthalpy diagram with oretical operational cycle R-134a centrifugal compressor at 100% load.

11 Energies 2017, 10, Constraint Conditions Centrifugal Compressor for Predicting Performance Table 3 lists inlet pressure temperature, specific volume ratio, flow coefficient for a full load (100% load) R-134a centrifugal compressor. In centrifugal chiller used in this study, outlet temperature chilled water evaporator was ± 0.3 K even if refrigerant was changed. Therefore, performance compressor was predicted while assuming that inlet temperatures first stage compressor were all identical. Because rotational speed compressor is fixed, re is no change in rotational speed, even with an alternative refrigerant. The percentage deviation specific volume ratio was calculated within tolerance described in Table 1. In addition, flow coefficient was assumed to be unchanged. The performance each stage centrifugal compressor was analysed based on above constraints. In addition, in this study, centrifugal compressor using R-1234ze (E) was designated as DR-34, centrifugal compressor using R-1233zd (E) was designated as DR-33 for convenience. Table 3. Inlet conditions R-134a compressor constrain conditions alternative refrigerants for performance predictions at 100% load state. 1st stage 2nd stage Refrigerant Symbol Unit R-134a (Ref.) R-1234ze (E) R-1233zd (E) Rotational speed N rpm Fixed rotational speed Inlet total pressure p in kpa Inlet total temperature T in K Specific volume ratio vr ~ ~1.65 Deviation specific volume ratios % vr - - 5%~+5% 5%~+5% Flow coefficient Φ Inlet total pressure p in kpa Using calculated data Inlet total temperature T in K from 1st stage Using calculated data from 1st stage Specific volume ratio vr ~ ~1.67 Deviation specific volume ratios % vr - - 5%~+5% -5%~+5% Flow coefficient Φ Results Discussion 5.1. Comparison Permissible Deviation by Alternative Gases 6 shows a comparison between results DR-34 DR-33 in permissible diagram 2. 6a shows results for each stage DR-34 DR-33 in permissible machine Mach number. All stages except for first stage DR-33 were within permissible. For first stage DR-33 to return to permissible, rotational speed must be reduced or sound speed at inlet must increase, as indicated by Equation (3). There are two oretical methods for increasing sound speed at inlet: increasing inlet temperature decreasing inlet pressure. In former case, if a superheat degree approximately 4 K is applied under assumption that re is no pressure decrease by absorption at inlet compressor, machine Mach number shifts into permissible. In latter case, inlet pressure compressor can be reduced by controlling chilled water, but it moves furr away from permissible because inlet temperature is also decreased. Therefore, effect adding superheat needs to be considered. However, for a more precise analysis, this effect should be analysed comprehensively via a cycle simulation performance test. 6b shows results each stage DR-34 DR-33 in permissible machine Reynolds number. The ratio machine Reynolds number DR-33 was relatively lower than that DR-34, but points both refrigerants were in permissible.

12 Energies 2017, 10, Energies 2017, 10, (a) (b) Comparison Comparison allowable allowable diagrams diagrams for for alternative alternative gases. gases. (a) (a) Allowable Allowable machine machine Mach Mach number number (b) (b) allowable allowable machine machine Reynolds Reynolds number. number Prediction Allowable Range Theoretical Operating Point 5.2. Prediction Allowable Range Theoretical Operating Point 7a,b show changes in total temperature, polytropic efficiency, polytropic work as 7a,b show changes in total temperature, polytropic efficiency, polytropic work as a a function changing total pressure in first- second-stage outlets DR-34, respectively. function changing total pressure in first- second-stage outlets DR-34, respectively. Here, flow coefficients DR-1 DR-2 were fixed at , respectively. The deviation Here, flow coefficients DR-1 DR-2 were fixed at , respectively. The deviation specific volume ratios varied from 5% to +6%, which is permissible error. The polytropic specific volume ratios varied from 5% to +6%, which is permissible error. The polytropic efficiency was compared by limiting calculation from 50% to 95%. 7a shows that, as efficiency was compared by limiting calculation from 50% to 95%. 7a shows that, first-stage outlet pressure DR-34 increases, outlet temperature polytropic work increase, as first-stage outlet pressure DR-34 increases, outlet temperature polytropic work while polytropic efficiency decreases. When specific volume ratio increases, minimum increase, while polytropic efficiency decreases. When specific volume ratio increases, maximum values s outlet pressure, outlet temperature, polytropic minimum maximum values s outlet pressure, outlet temperature, work increase. When specific volume ratio decreases, opposite occurs. In 7b, changes polytropic work increase. When specific volume ratio decreases, opposite occurs. In 7b, in outlet temperature polytropic efficiency as a function increasing DR-34 outlet pressure changes in outlet temperature polytropic efficiency as a function increasing DR-34 outlet at second stage show same tendency as that at first stage. However, at first stage, pressure at second stage show same tendency as that at first stage. However, at first polytropic work decreases, minimum maximum values

13 Energies 2017, 10, Energies 2017, 10, stage, polytropic work decreases, minimum maximum values polytropic work increase identically. To analyse difference in tendency polytropic polytropic work increase identically. To analyse difference in tendency polytropic work work first second stages, work factor isentropic volume exponent each stage first second stages, work factor isentropic volume exponent each stage were were compared, compared, as shown as shown in in 8. The 8. work The work factor factor increases increases gradually gradually with increasing with increasing pressure ratio pressure ratio in in first first stage, stage, but it but decreases it decreases in second in second stage. To stage. compare To compare values values isentropic volume isentropic volume exponent exponent in inwork-factor work-factor equation, equation, Equation Equation (19), even (19), when even when pressure pressure ratio changes, ratio changes, isentropic volume exponent is is greater than than 1 (kv 1 (k> V 1) > in 1) in first first stage stage but but less less than than 1 in 1 insecond second stage stage (k V (kv < 1). < 1). Therefore, first term denominator in inequation (19) (19) becomes negative negative in in second second stage. stage. Furrmore, product product outlet outlet pressure pressure outlet outlet isentropic isentropic specific specific volume volume second term second becomes term smaller becomes thansmaller product than product inlet pressure inlet pressure inlet specific inlet volume. Thus, specific second volume. term Thus, also becomes second term negative. also becomes Because negative. Because product product first second first terms second terms denominator, total value denominator becomes greater than denominator, total value denominator becomes greater than difference between difference between enthalpies numerator. As a result, work factor decreases. enthalpies numerator. As a result, work factor decreases. (a) (b) 7. Comparison outlet total temperature, polytropic efficiency, polytropic work for a 7. Comparison outlet total temperature, polytropic efficiency, polytropic work for permissible 5%vr to +5%vr for DR-34. (a) First-stage compressor at Φ = (b) secondstage compressor at Φ = vr to +5% vr for DR-34. (a) First-stage compressor at Φ = a permissible 5% (b) second-stage compressor at Φ =

14 Energies 2017, 10,10, 2043 Energies 2017, 2043 Energies 2017, 10, Comparison Comparison work factor isentropic volume exponent for each stage DR-34 0%. 8. work factor isentropic volume exponent for each stage DR-34 atatat 0% vr.vr 8. Comparison work factor isentropic volume exponent for each stage DR-34 0% vr. 9a shows expected point first-stage outlet 9a shows expected point first-stage outlet 9a shows expected point first-stage outlet DR-34 calculated oretically via proposed method. 9b shows expected DR-34 calculated oretically via proposed method. 9b shows expected DR-34 calculated oretically via proposed method. 9b shows expected expected point inlet outlet in second stage based on expected expected point inlet outlet in second stage based on expected expected point inlet outlet in second stage based on expected point outlet in first stage DR-34. point point outlet outletinin first firststage stagedr-34. DR-34. (a)(a) (b)(b) Calculated point forfor DR-34. (a)(a) First stage (b)(b) second Calculated point DR-34. First stage second 9. Calculated point for DR-34. (a) First stage (b) second stage. stage. stage.

15 Energies 2017, 10, Energies 2017, 10, The expected is defined as pressure when deviation specific The expected is defined as pressure when deviation specific volume ratio is between 5% +5% enthalpy when calculated polytropic volume ratio is between 5% +5% enthalpy when calculated polytropic efficiency efficiency is between 50% 95%, under condition that specific volume ratio R-134a is between 50% 95%, under condition that specific volume ratio R-134a design design DR-34 deviation specific volume ratio are 0%. Within expected DR-34 at deviation this time, centrifugal specific volume compressor ratio has are various 0%. Within expected points. Therefore, oretical at this time, expected centrifugal compressor point needs to has be various calculated using STEP-6 points. in Therefore, 1. The outlet oretical expected points each point stage needs in to be calculated 9a,b were using obtained STEP-6 under in following 1. The outlet limitations: percentage points each stage deviation in 9a,b specific were obtained volume under ratio is 0% following deviation limitations: ratio percentage polytropic work deviation between specific design volume condition ratio is 0% correction result deviation is 0%. ratio polytropic work between design condition correction result 10a,b show is 0%. expected expected points in each stage DR-33. The 10a,b overall showresults expected are similar to those for DR-34 owing expected to rmodynamic points incharacteristics each stage DR-33. R-1233zd The overall (E). results However, areit similar has a lower to those density for DR-34 compared owing to R-1234ze to rmodynamic (E); thus, outlet characteristics pressure R-1233zd in each stage (E). However, DR-33 is itlower has a than lower that density DR-34. compared to R-1234ze (E); thus, outlet pressure in each stage DR-33 is lower than that DR-34. (a) (b) 10. Calculated point for DR-33. (a) First stage (b) second stage. 10. Calculated point for DR-33. (a) First stage (b) second stage.

16 Energies 2017, 10, The minimum maximum values pressure temperature in expected at outlet in each stage DR-34 DR-33 are listed in Table 4. At same deviation specific volume ratios, lower efficiency is, higher are pressure temperature. When deviation specific volume ratios increases from 5% to +5%, both pressure temperature increase. When difference deviation, % vr, is analysed, pressure difference DR-34 is greater than that DR-33, pressure difference second stage is greater than that first stage. Furrmore, pressure difference increases furr if efficiency decreases. The temperature difference is approximately 3 K at a polytropic efficiency 95% for both DR-34 DR-33 approximately 6 K at 50% efficiency. Because results are generally more sensitive to temperature than to pressure at same efficiency point, more attention should be focused on temperature control during drop-in test. Table 4. Minimum maximum value pressure temperature at outlet by deviation specific volume ratios for polytropic efficiencies 50% 95%. DR-34 Deviation Specific Volume Ratios P min (kpa) T min (K) P max (kpa) T max (K) 1st stage 2nd stage 5% vr % vr % vr % vr (=5% vr ( 5% vr )) % vr % vr % vr % vr (=5% vr ( 5% vr )) Polytropic efficiency 95% 50% DR-33 Deviation Specific Volume Ratios P min (kpa) T min (K) P max (kpa) T max (K) 1st stage 2nd stage 5% vr % vr % vr % vr (=5% vr ( 5% vr )) % vr % vr % vr % vr (=5% vr ( 5% vr )) Polytropic efficiency 95% 50% 5.3. Validation Thermodynamic Performance Prediction 11a shows a diagram comparing prediction results for first second stages DR-34 with measurement results from drop-in test. The static pressure p s (kpa) static temperature T s (K) at inlet duct first stage centrifugal compressor at outlet second stage were converted to total pressure total temperature using Equations (26) (29), respectively, compared with prediction results. In ASME PTC-10, when fluid Mach number at measured position, Ma, is less than 0.2, it is regarded as incompressible because compressive effect decreases, total pressure is calculated as shown in Equation (26). If fluid Mach number is 0.2 or higher, total pressure should be calculated as shown in Equation (27). Because fluid Mach numbers at inlet outlet ducts are both less than 0.2, total pressure was calculated using Equation (26). Equation (30) gives average fluid speed V (m/s) using cross-sectional area at measured position: p = p s + V2 2 v s (Incompressible flow, Ma < 0.2) (26) [ p = p s 1 + k 1 ] k Ma 2 k 1 (Compressible flow, Ma 0.2) (27) 2

17 where k is ratio specific heats Cp (N m/kg K) is specific heat at constant pressure. The subscript s signifies static state. The measured point at inlet in first stage DR-34 is lower by approximately 11 kpa (approximately 4%) than oretically predicted point. The deviation inlet Energiestemperature 2017, 10, 2043 seems to be due to environment during actual operation closed-1circuit cycle chiller. 21 Under assumption that inlet temperature Ma = V is identical to design condition deviation specific volume ratio is 0%, predicted point is indicated by a solid (28) symbol a solid line in 11a. The pressure at a measured point outlet in second ( V 2 ) stage is lower by approximately 8.7 kpa (approximately 1.2%) than predicted point, but it is in predicted T. = T s Under + (1 0.65) assumption (29) 2 that C inlet condition in first stage p,s DR-34 is converted to measured temperature pressure deviation specific volume. ratio is 0%, predicted point is indicated m vby s an open symbol a short-dashed line in V = (30) 11a. A where k is ratio specific heats C p (N m/kg K) is specific heat at constant pressure. The subscript s signifies static state. Energies 2017, 10, (a) (b) 11. Comparison 11. Comparison predicted results drop-in test test results for for (a) (a) DR-34 DR-34 (b) DR-33. (b) DR-33. When inlet temperature pressure decrease, predicted points at first second-stage outlets decrease. Even when inlet conditions are changed, measured point is within predicted. 11b compares prediction results in first second stages DR-33 with measurements obtained via drop-in test. The solid symbol solid line indicate predicted point at 0%vr. The pressure temperature measured at

18 Energies 2017, 10, The measured point at inlet in first stage DR-34 is lower by approximately 11 kpa (approximately 4%) than oretically predicted point. The deviation inlet temperature seems to be due to environment during actual operation closed-circuit cycle chiller. Under assumption that inlet temperature is identical to design condition deviation specific volume ratio is 0%, predicted point is indicated by a solid symbol a solid line in 11a. The pressure at measured point outlet in second stage is lower by approximately 8.7 kpa (approximately 1.2%) than predicted point, but it is in predicted. Under assumption that inlet condition in first stage DR-34 is converted to measured temperature pressure deviation specific volume ratio is 0%, predicted point is indicated by an open symbol a short-dashed line in 11a. When inlet temperature pressure decrease, predicted points at first second-stage outlets decrease. Even when inlet conditions are changed, measured point is within predicted. 11b compares prediction results in first second stages DR-33 with measurements obtained via drop-in test. The solid symbol solid line indicate predicted point at 0% vr. The pressure temperature measured at inlet first stage are very close to oretically predicted point with a difference less than 0.5%. However, pressure temperature at measured point second-stage outlet are higher by approximately 20 kpa (approximately 15%) by approximately 2.4 K than those at predicted point, respectively. This shows that predicted point is low. The predicted point when operated at +5% vr in first stage DR-33 is indicated by an open symbol a short-dashed line in 11b. Even when deviation specific volume ratio increases in first stage, predicted point at second-stage outlet is low, predicted is lower than measured point. When we examine operation characteristics a general centrifugal compressor, if flow rate decreases at a fixed rotational speed, outlet pressure temperature become higher than design points. Because a decrease in flow rate implies operation at an f-design point, operation information for f-design point R-134a needs to be compared. On or h, from systematic viewpoint chiller, because a low-pressure refrigerant was injected into a medium-pressure refrigerant system, rmodynamic problem occurring in condenser evaporator can be considered. If drop-in test R-1234ze (E) R-1233zd (E) can be performed on a turbo chiller for a low-pressure refrigerant, rmodynamic influence heat exchanger can be analysed with complexity. Table 5 compares total pressure ratio, polytropic efficiency, gas power at calculated point alternative refrigerant with design conditions R-134a. This shows that pressure ratios first second stages are similar even with a different refrigerant. In case DR-34, efficiency second stage is lower than that first stage, but y are more similar in case DR-33. The gas power increases in second stage for both compressors because increase flow rate due to economizer in second stage R-134a was reflected when predicted performance drop-in test was calculated. Furrmore, DR-33 has a very low gas power owing to lower density flow rate compared to or refrigerants. Table 5. Performance prediction results alternative refrigerants at predicted points. Refrigerant Symbol R-134a R-1234ze (E) R-1233zd (E) Deviation specific volume ratio % vr st stage 2nd stage Total-to-total pressure ratio pr Polytropic efficiency (%) η p Gas power (kw) P g Total-to-total pressure ratio pr Polytropic efficiency η p Gas power (kw) P g

19 Energies 2017, 10, Conclusions In this study, a low-gwp alternative refrigerant was used in centrifugal compressor a conventional R-134a turbo chiller, performance centrifugal compressor in each stage, which is required for drop-in test, was predicted. A performance prediction method was proposed by changing existing performance test code, predicted point were calculated using proposed method. The following conclusions were obtained from this study. Polytropic work, which is one rmodynamic performance variables a centrifugal compressor, was found to be affected by isentropic volume exponent value. In predicted s DR-34 DR-33, temperature difference between deviations specific volume ratios was approximately 3 K at 95% polytropic efficiency approximately 6 K or less at 50% polytropic efficiency. This shows that predicted calculated based on permissible deviation ASME PTC-10 is sensitive to temperature at same efficiency point. The predicted DR-33 was small because pressure difference between deviations specific volume ratio was smaller than that DR-34. The predicted point at stage-two outlet DR-34 was higher by approximately 1.2% than that drop-in test. If point was predicted by changing stage-one inlet condition to test result, pressure at stage-two outlet was lower by approximately 2.6% compared to measured value. Although re was a quantitative deviation between measured predicted points because test result was in predicted s for two conditions, proposed method predicted DR-34 well. In case DR-33, predicted point were generally lower than test results. A more accurate evaluation requires additional analysis based on f-design points R-134a rmodynamic influence heat exchanger. In this study, we proposed a performance prediction method that can be used to predict point drop-in test alternative gases by using performance test code an existing centrifugal compressor. The proposed method can be used to predict evaluate performance in each stage a multi-stage centrifugal compressor. The results proved that performance a centrifugal compressor can be predicted when operated with an alternative gas based on limited information about factors such as design conditions shapes. Furrmore, additional information about f-design points a centrifugal compressor will be helpful for identifying predicted point more precisely. Acknowledgments: This research was supported by Technology Innovation Program (Grant No ) by Ministry Trade, Industry & Energy. Author Contributions: All authors contributed to this work by collaboration. Joo Hoon Park is main author this manuscript. Youhwan Shin Jin Taek Chung assisted provided useful suggestions in contents research. All authors revised approved publication. Conflicts Interest: The authors declare no conflict interest. Nomenclature GWP Global warming potential HFOs Hydrluoroolefins ODP Ozone depletion potential A Cross-sectional area flow channel (m 2 ) a Acoustic velocity (m/s) b Tip width (m) C p Specific heat at constant pressure (N m/kg K) D Diameter (m) f Work factor Schultz h Enthalpy total state (kj/kg) K T Isormal bulk modulus (kpa)

20 Energies 2017, 10, k V Isentropic volume exponent k Ratio specific heats (C p /C v ) Mm Machine Mach number Ma Fluid Mach number m Polytropic temperature exponent N Rotational speed (rpm) n Polytropic volume exponent P g Gas power (kw) p Pressure total state Q Volume flow rate at inlet (m 3 /s) R Gas constant (N m/kg K) RA Machine Reynolds number correction constant RB Machine Reynolds number correction constant RC Machine Reynolds number correction constant Rem Machine Reynolds number T Temperature total state (K) U Tangential speed (m/s) V Averaged fluid velocity (m/s) v Specific volume (m 3 /kg) vr Specific volume ratio % vr Percentage deviation between specific volume ratios W Work (N m/kg) X Compressibility function Schultz Y Compressibility function Schultz Z Compressibility factor. m Mass flow rate (kg/s) α Volume expansivity (1/K) η Efficiency µ Absolute viscosity (kg/m s) ν Kinematic viscosity (m 2 /s) ρ Density (kg/m 3 ) Φ Flow coefficient Subscripts imp, out corr d in is m max min out p recal s t vr Impeller outlet Corrected state Design condition or result Inlet Isentropic state Arithmetic mean Maximum value Minimum value Outlet Polytropic state Recalculation Static state Test condition or result Specific volume ratio References 1. Wang, X.; Amrane, K.; Johnson, P. Low global warming potential (GWP) alternative refrigerants evaluation program (Low-GWP AREP). In Proceedings International Refrigeration Air Conditioning Conference, West Lafayette, IN, USA, July 2012; Volume 2233, pp. 1 7.

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