Analysis of a Refrigerant Compressor Load Stand Incorporating Hot Gas Bypass and a Single Full Condensation Heat Exchanger

Transcription

1 Purdue University Purdue e-pubs International Compressor Engineering Conferene Shool of Mehanial Engineering 1984 Analysis of a Refrigerant Compressor Load Stand Inorporating Hot Gas Bypass and a Single Full Condensation Heat Exhanger J. A. MGovern Follow this and additional works at: MGovern, J. A., "Analysis of a Refrigerant Compressor Load Stand Inorporating Hot Gas Bypass and a Single Full Condensation Heat Exhanger" (1984). International Compressor Engineering Conferene. Paper This doument has been made available through Purdue e-pubs, a servie of the Purdue University Libraries. Please ontat epubs@purdue.edu for additional information. Complete proeedings may be aquired in print and on CD-ROM diretly from the Ray W. Herrik Laboratories at Herrik/Events/orderlit.html

2 ANALYSIS OF A REFRIGERANT COMPRESSOR LOAD STAND INCORPORATING ROT GAS BYPASS AND A SINGLE FULL CONDENSATION HEAT EXCHANGER J.A. MGovern, Leturer Department of Mehanial and Manufaturing Engineering Trinity College Dublin, Ireland ABSTRACT A refrigerant ompressor load stand is desribed and a theoretial analysis of its operation is presented. The purpose of the load stand is to establish and maintain pre-determined values of the sution pressure, the disharge pressure and the sution superheat. These are ontrolled by means of throttle valves. The power of ompression is rejeted to ooling water in a full ondensation heat exhanger whih normally operates lose to ambient temperature. The refrigerant flow is divided and the smaller part, generally less than one third, passes through the heat exhanger, while the greater part bypasses it as vapour. The two streams are re-ombined by adiabati mixing. Brief details are given of suh a load stand whih has been built reently, together with initial operating experienes. INTRODUCTION A researh programme with the title 'apaity Control of Reiproating Piston ompressors for Heat Pump or Refrigeration Appliations by Speed Variation' has been underway at Trinity College sine the Summer of The purpose of this work is to study the performane of reiproating ompressors when operated at different shaft speeds. As the programme will involve a great deal of performane testing of ompressors over a wide operating range of eah of four main parameters, the initial phase involves the onstrution and development of a suitable load stand. The four main parameters to be ontrolled are' mean sution pressure mean disharge pressure sution superheat temperature ompressor shaft speed. The first three are parameters of the load stand, while the fourth diretly influenes its apaity. In ontrolled parameters measure the following: ompressor shaft speed ompressor shaft torque (reation torque) refrigerant mass flow rate addition to the it is intended to (by means of an orifie plate) disharge vapour temperature. It is also pressure-volume measurements for onditions. intended to obtain diagrams and valve lift a full range of operating A load stand is proposed in whih the three main yle parameters are ontrolled by means of throttle valves and in whih part of the refrigerant is fully ondensed to the liquid state. Suh a system offers several potential advantages. As some exess liquid is always present it is not neessary to vary the system harge when the operating onditions are hanged. If the temperature of the ondensing heat exhanger is maintained onstant its thermal inertia and that of its ontents should be of small onsequene and a rapid system response to hanges in the operating onditions should be possible, Automati ontrol of the throttle valves should be feasible. The load stand is a variation on those desribed in referenes (1) and (2). In some respets the proposed system is analogous to a hydrauli dynamometer used in engine testing. The heat equivalent of the shaft power is rejeted to the ooling water and the onversion from work to heat is by means of totally irreversible proesses of fluid frition. A number of subroutines have for the analyses subroutines listed used for the alulations. omputer programs and been written in FORTRAN whih follow. The in referene (3) were refrigerant property 468

3 THERMODYNAMIC PROCESSES A iruit diagram to implement the basi proesses of the load stand is shown in Fig. 1. It is assumed that ondensation to the saturated liquid state ours in the heat exhanger and that the saturation pressure an be maintained onstant. The sution and disharge pressures of the ompressor are established with respet to the heat exhanger pressure by the throttle valves D and A. By some, as yet unspeified means, part of the total refrigerant flow leaving valve A is aused to pass through the heat exhanger, while the balane is bypassed to valve D. A w , D ' ' (..'~-- 4 adiabati mixing Figure 1. Load stand basi iruit By varying the proportion of refrigerant whih is bypassed the sution superheat may be ontrolled. The idealised thermodynami proesses are shown on a P-h diagram in Fig. 2. By applying the Steady Flow Energy Equation the flow rate of refrigerant through the heat exhanger an be expressed as a proportion of the total as followsb hi - h" 1 - (1) Table 1 is a set of results of this alulation for a wide range of operating onditions for refrigerant 12 when the saturation temperature in the heat exhanger is 20 deg. C. Similar tables have been produed for other values of the heat exhanger operating temperature. For example, at a saturation temperature of 30 deg. C the proportions of flow rate whih pass through the heat exhanger are higher by from 5% to 7%, depending on the pressure and superheat onditions. The main onlusion from Table 1 is that it is possible, in priniple, to produe any operating ondition of interest by varying the relative amounts of refrigerant whih are ondensed and bypassed. There are pratial diffiulties in implementing the yle as desribed thus far. There will be a pressure fall, however small, in any real heat exhanger and so a orresponding pressure fall will be required in the bypass line. These pressure losses will not upset the operation of the yle. A means of proportioning the flow is also neessary. This an be done by introduing throttle valves B and C as shown in Fig. 3. A 3 2.,.----4,..._------, e. al5 ~u Figure 2. Idealised thermodynami proesses h B 7--' 5 I ',. Figure 3. Ciruit inluding bypass and liquid line throttle valves The effet these valves have on the thermodynami yle is shown on a P-h diagram in Fig. 4. The inlusion of these extra throttle valves does not alter the values presented in Table 1 for the proportion of the refrigerant flow whih passes through the heat exhanger. The pressure at whih adiabati mixing of the ondensed Iiquid and bypassed vapour ours may lie at any level between that of the 469

4 heat exhanger and that at sution, depending on the settings of 'valves B, C and D. ~is phx pmix- Ps" Figure 4. inluding throttling. Thermodynami liquid line The Ideal Load Stand and Compressor 2 A 3 B 6 h proesses, and bypass For the purposes of designing or sizing a load stand an ideal ompressor whih an be desribed in terms of its displaement rate and learane ratio is a suitable model. The volumetri effiieny of suh a ompressor is given by '1 ( 2) By using this equation in onjuntion with the displaement rate and the thermodynami relationships for the yle, the ideal mass flow rate, the proportions of refrigerant ondensed and bypassed, and the rate of heat rejetion in the heat exhanger an be alulated for any ombination of operating onditions. The results of suh alulations, assuming a displaement rate of 1 litre/se., a learane ratio of 3% and a heat exhanger saturation temperature of 20 deg. C are given in Tables 2 and 3. The values of heat rejetion rate, given in Table 3, apply regardless of the heat exhanger operating temperature. These figures an be used in seleting a suitable heat exhanger for the load stand. Effetive Valve Flow Areas The value hosen for the mixing pressure is a matter of ontrol strategy and, when this has been deided upon, the required flow areas for the four throttle valves an be alulated. The valves are modeled as orifies having a oeffiient of disharge of 1 and an thus be fully desribed by an effetive flow area. For valve, whih the refrigerant enters as a liquid, inompressible flow is assumed and the effetive flow area is given by my;-;;; ( 3) A.~ For valves A, B AND D the flow is ompressible and, depending on the pressure ratio in eah ase, may be hoked. In the alulations either the atual or the riti~l pressure ratio, whihever is the greater, is used. For the purposes of alulating the required orifie area isentropi flow is assumed, although, for the throttling proess as a whole, enthalpy is the property whih remains unhanged. The isentropi proess is modeled as being polytropi, i.e. one for whih onst. ( 4) While this assumption may be onsidered reasonably valid for refrigerant vapour, even within the saturation region, the ideal gas equation should not be assumed to apply, espeially when the refrigerant is saturated for part of the proess. The equation for the required flow area should therefore be in terms of the properties pressure and volume, rather than, for example, pressure and temperature. The following equation is used: A m [Pu ~ (rpfi - r~)~-.s ( 5) Vu n-1 J It should be noted that for some sution onditions point 7 in Fig. 4 lies within the saturation region on the P-h diagram. The speifi volume at this point an be found sine the speifi enthalpy is the same as at point l. Values of the required effetive flow areas for the valves are presented over a range of operating onditions in Table 4. It has been assumed in the table that one fifth of the pressure fall from the heat exhanger saturation pressure to the sution pressure ours aross the bypass and liquid line throttle valves B and C. It an be noted that any hange in the operating onditions requires a hange in the areas of all four valves. The required flow areas for valve, the liquid line throttle valve, are very muh smaller than for the other valves whih throttle vapour. HEAT EXCHANGER SELECTION, CHARACTERISTICS AND ANALYSIS The displaement rate of the ompressor to be tested was about 3.7 litres/seond and, using data from Table 3, the required maximum heat rejetion rate was found to be about?.2 kw, whih also oinided with the 470

5 motor rating. A water ooled ondenser heat exhanger with some apaity to hold exess liquid was desired and a shell and tube type was seleted. Either a single pass or a two pass water piping arrangement was possible. The single pass arrangement was hosen in order to maximise the flow rate and to minimise the temperature hange of the water in passing through the heat exhanger. Heat Transfer Rate The manufaturer presented the heat transfer rate as a funtion of the water flow rate and the temperature differene defined in terms of the saturation temperature and the water inlet temperature. AT ( 6) Heat transfer orrelations suggest that the water side heat transfer oeffiient depends on the flow rate, raised to the power of 0.8 approximately. Hene, the following expression was derived for the rate of heat transfer Q [ a v 8 l 1 + b v ~ OT (7) By regression analysis of points taken from the manufaturer's graphs the following oeffiients were obtainedr a b l. 933 (8) ( 9) A graph of this expression is given in Fig. 5. Log Mean Temperature Differene The onstants given in equations 8 and 9 were re-evaluated for the ase where the log mean temperature differene is used in equation 7. The values are a b (11) ( 12) A similar fit to that shown in Fig. 5 was obtained using these onstants. RW/K '<I 3 I G.2.I 0 0. '.2..3 WATER f"low RAT Figure 5. Heat transfer rate temperature differene as a water flow rate. Water Pressure Loss.s. 6 1/s per degree funtion of It was found that the manufaturer's data ould be represented reasonably losely by the following expression: Care should be taken in applying equations 6 to 9 for flow rates lower than those quoted in the manufaturer's data (0.19 litres/se.), as no impliit aount is taken of the water exit temperature. 6P where and k = j X (13) (14) (15) The temperature differenes were onverted to log mean values based on the refrigerant saturation temperature and the water inlet and outlet temperatures using the following equation (Ts - T,.,.,l - ('I's - Twi) In (~: - ~:~) ( 10) WATER CIRCUIT ANALYSIS The maximum available water supply rate was about 4 litres per minute and the maximum head between a onstant head tank installed in the laboratory and the drain was about 1.7 metres. The water iruit whih was adopted is shown diagramatially in Figure 6. Valve E ontrols the rate of water flow to drain and hene the saturation temperature within the heat exhanger. 471

6 1 I. 7tn shut-off valves were used for throttling. Valves A, B and D were three quarter inh flare and valve C was quarter inh flare. It was found diffiult to ontrol the sution superheat temperature and this was attributed to the insensitivity of valve. This was replaed with a one eight inh needle valve and it was then found possible to ahieve and maintain desired operating onditions. However, the operating experiene to date is very limited. Figure 7 is a photograph of the load stand. A sight glass loated upstream of the liquid line throttle valve, C, is used to ensure that liquid is present. CONCLUSIONS PUMP Figure 6. Heat exhanger water iruit The equation representing the pressure loss through the heat exhanger has already been mentioned, The Dary equation was used to determine pressure losses in the pipes and approximate empirial loss oeffiients were used to take aount of the bends and tees. A standard domesti heating irulation pump, having three speeds, was used and its pressure versus flow rate harateristis were obtained from the suppliers. The maximum flow rate through the heat exhanger will our when there is no flow to drain and is limited by the available head between points X and Y in Figure 6. This flow rate was alulated to be about 20 litres per minute. At its lowest speed the pump would produe a flow rate of about 16 litres per minute and at the seond speed a flow rate of about 21 litres per minute. Thus the alulations indiated that a flow to drain was possible at the lowest pump speed, or at a higher speed with throttle valve F partially losed. The onlusion from the analysis was that a onsiderable inrease in the flow rate through the ondenser ould be produed by reirulation, but, that this was not essential to the operation of the system. One effet of a high reirulation rate is that subooling of the refrigerant in the heat exhanger is minimised. A straightforward analysis of the load stand has been presented and the results indiate that it is feasible in priniple. From the analysis and the limited operating experiene to date it is felt that purpose designed throttle valves of higher sensitivity will be required. NOTATION A a,b h j,k n p ~p Q r., rp T v 4T v '1 "' Subsripts effetive flow area of speified valve, sq. m onstants speifi enthalpy of refrigerant at speified state, J/kg onstants in pressure loss equation mass flow rate of refrigerant, kg/s polytropi index for a reversible adiabati proess of the refrigerant pressure, N/(sq. m) pressure differene, N/(sq. m) rate of heat transfer, kw learane ratio (= ompressor learane volume divided by ompressor swept volume) pressure ratio or ritial pressure ratio for a valve temperature, K temperature differene, K volume flow rate of water, litres per seond speifi volume at speified state, (u. ml/kg ompressor volumetri effiieny THE LOAD STAND AS BUILT A load stand has been built with the analyses presented. operated diaphragm-sealed in aordane Standard hand refrigeration dis hx mix s sue disharge heat exhanger mixing refrigerant saturation property sution 472

7 tot u wi wo 1-7 REFERENCES total upstream property water inlet property water outlet property refrigerant equilibrium states (Fig. 4) 1. sodel, w., "Introdution to Computer Simulation of Positive Displaement Compressors", Purdue University, Marriott, L.W.; F.R. Lady; L.L. Evans, "Aelerated Test for Rating Positive Displaement Compressors Using Computer Control", Proeedings 6f the 1974 Purdue Compressor Tehnology Conferene. 3. Kartsounes, G.T.; R.A. Erth, (1971), "omputer Calulation of the Thermodynami Properties of Refrigerants 12, 22 and 502", ASHRAE Transations, vol. 77, part II, pages 88-l03. Figure 7. The refrigerant ompressor load stand 473

8 HEAT EXCHANGER MASS FLOW RATE AS A PROPORTION OF THE TOTAL REFRIGERANT NO.: 12 COMPRESSOR MASS FLOW RATE HEAT EXCHANGER OPERATING TEMP., DEG. C: 20.0 SUCTION SUPERHEAT, K: o.o DEG SUCTION SUPERHEAT, K: 10.0 DEG SUCTION SUPERHEAT, K: 20.0 DEG o.o ~ DEG Table 1 474

9 HEAT EXCHANGER Jl'iASS FLOW RA'I'E IN KG/S/1000 FOR AN IDEAL COMPRESSOR REFRIGERANT NO.: 12 OF GIVEN DISPLACEMENT RATE AND CLEARANCE RATIO HEAT EXCHANGER OPERATING TEMP., DEG. C: 20.0 COMPRESSOR DISPLACEMENT RATE, LITRES PER SECOND: 1.00 CLEARANCE RATIO:.030 SUC'l'ION SUPERHEAT, K: 0.0 SUCTION SAT. TEMP. f'isf'<ll\fgf' SPTVRrTI0N TEMP., f:p('. DEG C ] l oooo UC'TI0N SUPERHF:AT, K; 10.0 SUCTION SAT. 'I'El-lP. Dl f.c'l-jllrgf: SP':'ORM'H!N TgMP. L'Jo:G. DEG , SUC'l'ION SUPERHEAT, K: 20,0 SUC'l'ION Sl).T. TEMP. DISCHARGE SATURATION TEMP., DEG. DEG so.o I l l ] ] ,9593.oooo SUCTION SAT. TEMP. DISCHARGE SATURATION TEMP., DEG. DEG ] l l ] oooo Table 2 475

10 HEAT EXCHANGER BEAT TRANSFER RATE IN KW FOR AN IDEAL COMPRESSOR OF GIVEN DISPLACEMENT RATE AND CLEARANCE RATIO REFRIGERANT NO.: 12 COMPRESSOR DISPLACEMENT RATE, LITRES PER SECOND: 1.00 CLEARANCE RATIO:.030 SUCTION SUPERHEAT, K: o.o DEG oooo ] SUCTION SUPERHEAT, K: 10.0 DEG , SUCTION SUPERHEAT, K: 20.0 DEG , ,5903 SUCTION SAT. TE!YIP. DISCHARGE SATURATION TEMP., DEG. DEG , Table 3 476

11 EFFECTIVE FLOW AREAS OF VALVES ON THE TEST S'I'AND FOR AN IDEAL COMPRESSOR OF GIVEN DISPLACEMENT RATE AND CLEARANCE RATIO **** UNITS SQUARE MM **** SYMBOLS C - CHOKED FLOW, I - IMPOSSIBLE OPERATING CONDITION, 0 - VALVE FULLY OPEN Z - ZERO COMPRESSOR FLOW IT IS ASSUMED THAT THE ADIABATIC MIXING PRESSURE (=P7=P6=P5) LIES BETWEEN THE HEAT EXCHANGER PRESSURE (=P3=P4) AND THE SUCTION PRESSURE (=P1) AND THAT P7 = P3- (P3-P1)*(.20) REFRIGERANT NO.: 12 ADIABATIC INDEX (BASED ON P AND V): HEAT EXCHANGER OPERATING TEMP. DEG. C: 20.0 COMPRESSOR DISPLACEMENT RATE, iitres PER SECOND: 1.00 CLEARANCE RATIO:.030 *** VALVE A *** SUCTION SUPERHEAT, K: 0.0 SUCTION SAT. TEMP. DISCHARGE SAT. TEMP., DEG. DEG SUCTION SAT. TEMP. DISCHARGE SAT. TEMP., DEG. DEG l SUCTION SUPERHEAT, K: 0.0 *** VALVE B *** SUCTION SAT. TEMP. DISCHARGE SAT. TEMP., DEG. C DEG. C ~ ~~ ~ SUCTION SAT. TEMP. DEG. C SUCTION SUPERHEAT, SUCTION SAT. TEMP. DEG SUCTION SUPERHEAT, SUCTION SAT. TEMP. DEG l K: K: SUCTION SUPERHEAT, K: 0.0 DISCHARGE SAT. TEMP., *** VALVE C *** DISCHARGE SAT. TEMP., DISCHARGE SAT. TEMP., *** VALVE D *** DEG. C DEG. DEG SUCTION SAT. TEMP. DISCHARGE SAT. TEMP., DEG. C DEG. C ~~------~ ~ ~~ ~ ~ ~ SUCTION SAT. TEMP. DEG. C DISCHARGE SAT. TEMP., DEG. C Table 4 477