ABSTRACT 1. INTRODUCTION

Transcription

1 NUMERICAL ANALYSYS AND EXPERIMENTAL VALIDATION OF TRANS-CRITICAL CARBON DIOXIDE CYCLES FOR SMALL COOLING CAPACITIES. HERMETIC AND SEMI-HERMETIC CO 2 RECIPROCATING COMPRESSOR COMPARISON J. RIGOLA, G. RAUSH, C.D. PÉREZ-SEGARRA, N. ABLANQUE, A. OLIVA Centre Tecnològic de Transferència de Calor (CTTC), Universitat Politècnica de Catalunya (UPC) ETSEIAT, C/ Colom, 11, Terrassa (Barcelona) Spain , cttc@cttc.upc.edu J.M. SERRA, J. PONS, M. JORNET, J. JOVER ACC Spain, S.A. C/ Antoni Forrellat, s/n, Sant Quirze del Vallès (Barcelona), Spain , ABSTRACT Refrigerating systems are mostly based on cycles using H(C)FCs fluids. The EU obliges to reduce the greenhouse gases emission by 8% in 2010, following international Kyoto protocol. The use of CO 2 is the only refrigerant replacement with the security of not being harmful to the environment, non-toxic and non-flammable. Only in vending machines the CO 2 emissions can be reduced by 3 Mton of CO 2 equivalent per year replacing R134a by CO 2. The present work is a numerical and experimental comparative study of the whole refrigerating cycles in general, and the reciprocating compressors in particular, between an R134a conventional refrigerating cycle and a carbon dioxide trans-critical refrigerating system. The comparative cases presented have been specially designed for commercial refrigeration small cooling capacities. Different one-stage hermetic and semi-hermetic reciprocating compressor have been numerically modelled and experimentally validated. The conventional compressor has been experimentally tested in a calorimeter set-up, while the carbon dioxide compressors have been validated in a specific experimental unit specially designed and built to analyse high-pressure single stage vapour compression trans-critical refrigerating equipments. In order to validate this laboratory set-up, a detailed numerical simulation of the thermal and fluid dynamic behaviour of single stage vapour compression refrigerating unit has been developed and used. The numerical and experimental results obtained shows a good agreement, while the comparative global values shows very similar efficiencies between conventional system and the new trans-critical CO 2 cycle under the same working conditions. In all cases the semi-hermetic compressor present higher mass flow rate and better efficiency than the hermetic ones. 1. INTRODUCTION The Montreal protocol (UNEP, 1987) stipulated the phasing out of CFCs and HCFCs as refrigerants that deplete the ozone layer (ODP); while the Kyoto protocol (UNFCCC, 1997) encouraged promotion of policies for sustainable development and reduction of Global Warming Potential (GWP), including the regulation of HFCs. Therefore, the ecological problems caused by refrigerating units need an urgent solution. Sustainable development in the refrigeration and airconditioning fields implies the unavoidable use of natural available substances: ammonia or hydrocarbons (toxic and/or flammable), air (poor efficiency), water (limited range of applications) and carbon dioxide. Thus, the investigation and use of new and natural refrigerants is an important goal.

2 During the last decade, the investigation indicates that the use of carbon dioxide has an important interest as a natural fluid refrigerant (Lorentzen, 1994), (Jakobsen, 1998), (Kruse, 1999) and (Fleming, 2003). The use of carbon dioxide as a non-toxic, non-flammable, harmless and environmentally neutral fluid refrigerant must not only be an attractive alternative, but also a competitive solution. Total Equivalent Warming Impact (TEWI) is the parameter that must be taken into account. TEWI evaluates the sum of the direct contribution to GWP due to refrigerant escape, and the indirect contribution produced by carbon dioxide emissions resulting from the energy required to operate the system over its normal life. In conclusion, the way to obtain reliable cleaning systems puts up with the use of natural refrigerants, e.g. CO 2, and the obtaining a more energy efficient and emission reduction alternative refrigeration system. After 1992 the use of trans-critical carbon dioxide cycles as a natural heating and cooling alternative systems have had a revival. During last two years, it has been an important increase of reviews (Kim, 2004), seminars (Seminario Europeo sugli impianti a CO 2, 2005) and special issue journals (IRJ, 2005) about fundamental process and system design issues in carbon dioxide systems, where the important trends in CO 2 technology in refrigeration, air-conditioning and heat pumps applications are provided. The aim of this work is to present a comparative analysis between hermetic and semi-hermetic reciprocating compressor prototypes and its experimental behaviour in a whole trans-critical cycle working under similar conditions as a small cooling application. The numerical simulation model of hermetic reciprocating compressors (Pérez-Segarra, 2003), the thermal and fluid dynamic analysis of double-pipe condensers and evaporators (García-Valladares, 2004), together with the numerical simulation and experimental validation of vapour compression refrigeration systems (Rigola, 2005) has been previously presented and explained in detail. The novelty in this paper has been the analysis, development, construction and testing of semi-hermetic reciprocating compressor prototypes, modelled on the basis of the hermetic optimisation and taking also into advantage the idea to exchange the heat from the suction plenum and cylinder head with the ambient instead of the fluid flow between compression crankcase and shell. The results presented show the promising expectations of CO 2 as fluid refrigerant under the studied working conditions and both hermetic and semi-hermetic results comparison. 2. NUMERICAL SIMULATION MODELS The basic idea of the here briefly explained numerical simulation models developed, improved and used to obtain the numerical results presented below, is that the numerical simulation model of hermetic reciprocating compressors is used to develop detailed studies of the compressor behaviour analysis and also to obtain the numerical parameters which define the volumetric, isentropic and mechanical-electrical and heat transfer shell losses efficiencies evolution which feed the numerical simulation model of vapour compression refrigerating units developed and adapted to trans-critical cycles Numerical simulation model of hermetic reciprocating compressors The numerical simulation model of the thermal and fluid dynamic behaviour of hermetic reciprocating compressors is based on the integration of the fluid conservation equations (continuity, momentum and energy) in a transient and one-dimensional form along the whole compression domain (suction line, compression chamber and discharge line).

3 The governing equations are discretised using an implicit control volume formulation and a SIMPLE-like algorithm extended to compressible flow. Effective flow areas are evaluated considering a multidimensional approach based on model analysis of fluid interaction in the valve. The force balances in the crankshaft and connecting rod mechanical system are simultaneously solved at each time-step. The thermal analysis of the solid elements is based on global energy balances at each macro volume considered (shell, muffler, tubes, cylinder head, crankcase, motor, etc.). The model allows the possibility to obtain the detailed instantaneous evolution of pv diagram, suction and discharge chambers pressure and temperature evolution, together with suction and discharge valves displacement. The instantaneous motor torque, angular compressor velocity and acceleration, together with instantaneous compression work or mass flow leakage are also the information obtained. Some interesting aspects to highlight are the differences between the new carbon dioxide and the conventional R134a compressors on pv diagram or mass flow leakage like the illustrative case depicted in Figure 1. Figure 1. Detailed variables evolution over a comparative CO 2 vs. R134a cases Numerical simulation model of vapour compression refrigerating units The numerical resolution consists of a main program composed of different subroutines. The mathematical formulation of these subroutines has been carried out to solve the single phase and two-phase flow inside a characteristic duct control volume, together with the conduction heat transfer along solid tube control volume. The different elements of the equipment (evaporator, compressor, gas cooler, expansion device and auxiliary connecting tubes) are solved by means of the mentioned subroutines called in a convenient way. The double pipe heat exchangers used are directly evaluated by means of the finite-volume technique based on a one-dimensional and transient integration of the conservative equations of the fluid flow (continuity, momentum and energy), numerically integrated using a fully implicit numerical scheme and solved by means of the SIMPLEC pressure-based method. The CO 2 general correlations for heat transfer and fluid flow behaviour are detailed in (Rigola, 2005). The numerical resolution of the heat exchangers is based on standard segregated algorithm that solves iteratively the different zones (fluid refrigerant flow, secondary fluid, natural convection and radiation in the external zone and temperature distribution in the solid elements of internal and external tubes).

4 The expansion device is evaluated in a similar way than fluid refrigerant in the heat exchangers when capillary tube is considered, and taking into account that in this case, all the inflow conditions cannot be simultaneously input data, as the critical mass flow rate is fixed for a given capillary tube. In the case herewith presented a commercial valve as expansion device, have been used and selected to provide an accurate desired control flow rates. In this case, the numerical model is based on considering the flow through the valve as a sudden contraction along the tube m& = A C ρ ( p ). The contraction of the tube is function of the C D, which represents the D D p6 coefficient of flow C D = ρ / ρ, obtained from the characteristic behaviour of the 6 valve. The A D is function of the number of turns open of the valve. The compressor inside the global cycle simulation is modelled on the basis of global energy balance between the inlet and outlet cross-sections of the compressor considering cyclical steady state. In this case, the compressor behaviour is characterized by the following parameters: volumetric c= 0 efficiency ( η = ( m & / ρ ) & ); isentropic efficiency ( η = w / w ); mechanical and electrical v 1 / G s efficiency ( η = w / w ), and heat transfer losses efficiency ( η = ε = 1 ( q / w ) ). The me cp e above mentioned compressor parameters are function of the compressor geometry, fluid refrigerant, compression ratio, inlet compressor temperature, etc. A detailed description of the parameters used in this paper is presented in (Pérez-Segarra, 2005). This formulation requires additional empirical information for the evaluation of volumetric, isentropic and mechanical-electrical efficiencies; together with heat transfer losses efficiency. In this paper, this information is obtained by means of the numerical simulation model hermetic reciprocating compressors presented above. s s cp Qsh 1 Qsh sh e The algorithm to solve the global equations system is based on a successive substitution method. Thus, at each time step or outer iteration, the subroutines that solve all the different elements are called sequentially, transferring adequate information to each other until convergence is reached. Transferred information depends on whether transient or steady state is considered. The boundary conditions for the simulation of the whole system are the inlet temperature, pressure and mass flow rate of the secondary flow in the gas cooler and evaporator, the compressor speed, the ambient temperature and pressure, and the number of turn over in the valve. 3. EXPERIMENTAL SET-UP DESCRIPTION An experimental set-up has been specially designed to evaluate the thermal and fluid dynamic behaviour of carbon dioxide trans-critical cycles and to validate the numerical simulation results of the numerical model presented in the companion paper. A schematic diagram and a general view of the refrigeration system are depicted in Figure 2. The experimental unit is made up of the following elements: a one stage carbon dioxide hermetic reciprocating compressor prototype, dual heat transfer coil gas cooler and evaporator together with a metering valve. The auxiliary fluid used in the gas cooler and the evaporator annuli is water. Tables 1 and 2 show the cycle components and instrumentation elements parameters, respectively. Two thermostatic heating and cooling units control the inlet auxiliary water temperature in the condenser and evaporator auxiliary circuits, respectively. The volumetric flow in these secondary circuits is controlled by two modulating solenoid valves and measured by means of two magnetic flow-meters, with an accuracy of ± 0.01 l/min. from 0 to 2.5 l/min., and ± 0.5 % F.S. from 2.5 l/min. to 25 l/min. Compressors HCL15 and SHCL15 are the CO 2 hermetic and semi-hermetic compressor prototypes, respectively.

5 Table 1. Experimental set-up components Stainless steel tubing ¼ OD Stainless steel tubing ¼ OD Dual HTC gas cooler ¼ OD Dual HTC evaporator ¼ OD Dual HTC annulus ½ OD Dual HTC annulus ½ OD Dual HTC length 4.5 m Dual HTC length 4.5 m Dual HTC insulation 20 mm Dual HTC insulation 20 mm Compressor capacity 1.5 cc Metering valve Cv 0.04 cc Figure 2. Experimental set-up unit scheme and illustration. Table 2. Experimental set-up measurements Mass flow meter kg/h Security valve X T 0.67 Limit, accuracy % Security valve Cv 0.41 Pt100 error <0.05ºC Security valve spring bars Pt100 accuracy <+-0.1ºC K thermocouples error <0.05ºC Pressure limits 0-100/0-160 bars K thermocouples accuracy ºC Pressure accuracy <0.1% span Metering valve limits bars 4. RESULTS Three aspects are analysed in the present section. First of all, a thermodynamic analysis of the transcritical cycle is studied in order to obtain the working conditions that offer the optimum COP. Subsequently; a second version of the hermetic reciprocating carbon dioxide compressor prototype has been numerically studied, and experimentally validated. Finally, a detailed analysis of the conventional R134a sub-critical cycle is compared against the detailed results of the second hermetic reciprocating prototype and an improved semi-hermetic reciprocating prototype. This second version of semi-hermetic reciprocating carbon dioxide compressor prototype has been developed using the numerical results of the hermetic ones obtained from the numerical simulation tool briefly explained above and the experimental results of the previous prototypes (Raush, 2005) (Rigola, 2005) Optimum gas cooler pressure The trans-critical CO 2 refrigerating systems are based on the Gustav Lorentzen cycle (Lorentzen, 1994). At super-critical conditions, the high pressure is independent of the temperature, and some kind of pressure adjustment is necessary, as both cooling capacity and COP depend on gas cooler pressure.

6 250 6 qev, wcp (kj/kg) qev Tev = -25 C wcp Tev = 7.2 C Tev=-10C Tev = 0 C Tev=-25C Tev = -10 C Tev = 0 C Tev=7.2C COP Tev = 7.2 C Tev=0C Tev=-10C Tev = -25 C gas cooler pressure (bar) gas cooler pressure (bar) Figure 3. Thermodynamic evolution of specific evaporation cooling, compression work and COP depending on gas cooler pressure. Figure 3 shows the thermodynamic analysis of the trans-critical cycle (isentropic compression, isenthalpic expansion and isobaric heat exchangers), considering inlet compressor and outlet gas cooler temperatures of 32ºC at different evaporation temperatures from 25ºC to 7.2ºC. In all cases, the specific compression work increases linearly, while the COP shows a maximum value for all evaporation temperatures around 80 bars. The final optimum gas cooler pressure selected has been 85 bars, taking into account the evolution of the specific evaporation cooling from 80 to 90 bars. Thus, the final optimum value is a compromise between the maximum COP and a minimum cooling capacity. The present results have been obtained not only considering the numerical analysis presented but also an experimental study (Raush, G. 2005) that observe the same conclusion with variations around ± 2.5 bar under the same working test conditions of Figure Carbon dioxide trans-critical cycle results A second version of CO 2 hermetic compressor prototype (HCL15(2)), detailed on (Rigola, 2005), has been numerically tested with the numerical simulation models of section 2 and experimentally validated with the set-up described in section 3. Table 3 shows the comparative results of the carbon dioxide trans-critical cycle under three different evaporation temperatures of 10, 0 and 7.2ºC, respectively, with an inlet fluid compressor and outlet gas cooler temperatures of 32ºC and a gas cooler pressure of 85 bars, in all three cases. The values of gas cooler pressure in brackets are boundary conditions. Table 3. Global carbon dioxide cycle numerical results vs. experimental data. pgc pev T2 T4 T7 8 x T g 6 m& (bar) (bar) (C) (C) (C) (C) (kg/h) numerical (84.95) experimental (84.95) numerical (84.98) experimental (84.98) numerical (85.78) experimental (85.78) The results of Table 3 show the reasonable good agreement between the experimental data and the numerical ones. Differences on mass flow rate are lower than 8%. The main difference on temperatures is observed at the outlet compressor temperature T 2, which is lower than 9%. Differences on outlet gas cooler T 4 and outlet evaporator T 8 are less than 2% in studied cases.

7 4.3. Comparative analysis between CO 2 hermetic and semi-hermetic compressors Once different carbon dioxide hermetic reciprocating compressors have been numerically analysed and optimised, built and experimentally tested, semi-hermetic reciprocating compressors have been build following the same strategy, taking the advantage to design a direct solution and a cylinder head in contact with the ambient. Both changes and some news improvements on geometry after the first SHCL15(1) version have allowed to obtain this second and improved SHCL15(2) semihermetic compressor described on (Raush, 2005) with some minor changes. Table 4 shows the experimental global comparative results between the hermetic HCL15(2) and semi-hermetic SHCL15(2) carbon dioxide compressors working in the trans-critical set-up cycle described on section 3 under the conditions of subsection 4.2, against R134a commercial ACC hermetic reciprocating compressor working in a sub-critical conventional cycle following ISO 917, under the working conditions of inlet compressor temperature 35ºC, condenser temperature 55ºC and outlet fluid temperature 46ºC. Table 4. Global experimental comparative results (conventional R134a compressor vs. hermetic and semi-hermetic carbon dioxide compressors) Tevap m& ηv W & e ηsme Q & evap COP (C) (kg/h) (%) (W) (%) (W) GLY HCL15(2) ,0 333,8 53,0 463, SHCL15(2) GLY HCL15(2) ,1 353,5 59,1 681, SHCL15(2) ,5 374,4 64,6 788, GLY HCL15(2) ,7 365,2 61,1 841, SHCL15(2) ,6 353,6 72,4 964, The results of Table 4 indicate that hermetic and semi-hermetic improved carbon dioxide compressor prototypes of 1.5 cc are able to produce a cooling capacity of 400, 600 and 800 W at 10, 0 and 7.2ºC of evaporation temperatures, respectively. The mass flow rate of CO 2 hermetic compressor is quite similar than GLY80 conventional R134a hermetic one. However, the semihermetic CO 2 presents a higher mass flow rates around 16, 14 and 10% for the evaporation temperatures of 10, 0 and 7.2ºC respectively. Despite of this, the semi-hermetic compressor still shows a 10% COP lower against conventional R134a compressor, although the same semi-hermetic compressor presents a higher COP around 3% and 6% in comparison with R134a conventional compressor at 0ºC and 7.2ºC of evaporation temperatures, respectively. In all studied cases, the CO 2 semi-hermetic compressor indicates a better COP around 4, 8 and 15% in comparison with the CO 2 hermetic compressor at 10, 0 and 7.2ºC, respectively. 5. DISCUSSION A numerical and experimental comparative study between conventional sub-critical cycle and carbon dioxide trans-critical refrigerating system has been presented. The CO 2 trans-critical cycle has been numerically solved and experimentally validated. Two CO 2 improved hermetic and semihermetic compressor prototypes have been designed, build, tested and compared. The comparative results conclude that trans-critical CO 2 semi-hermetic compressor not only present higher improvement on mass flow rate and COP in comparison with the trans-critical CO 2 hermetic compressor in all studied cases and evaporation temperatures, but also better cooling capacities than conventional R134a sub-critical compressors, an equal COP at 0ºC of evaporation temperature and only a 10% COP lower at 10ºC of evaporation temperature. Then, results show promising perspectives due to the fact that CO 2 compressors are still development prototypes in progress.

8 NOMENCLATURE A D metering valve section (m 2 ) T fluid temperature (ºC) C D contraction coefficient tube w specific work (J/kg) COP Coefficient of Performance W & e power consumption (W) c=0 G & isentropic volumetric flow (m 3 /s) x g vapour mass fraction m& mass flow rate (kg/h) ρ fluid density (kg/m 3 ) p fluid pressure (bar) η sme iso-mechanical-electrical efficiency (%) Q & cooling capacity (W) evap η v volumetric efficiency (%) REFERENCES International Institute of Refrigeration, 1975, New International Dictionary of Refrigeration, 3 rd ed., IIF/IIR, Paris, 560 p. Optimal IR, Fee W. 1986, Calculation of conference attendance budgets, Int. J. Cost-Effective Research, 7(4): UNEP 1987, United Nations Environmental Program Montreal Protocol on substances that deplete the ozone layer. UNFCCC 1997, United Nations framework convention on climate change. Kyoto Protocol. Lorentzen G, 1994, Revival of carbon dioxide as a refrigerant. International Journal of Refrigeration, 17(5): Jakobsen A, 1998, Improving efficiency of trans-critical CO 2 refrigeration systems for reefers. In IIF-IIR Commission D2/3, pages , Cambridge, UK. Kruse H, Heidelck R, Süss J, 1999, The application of CO 2 as a refrigerant. Bulletin of the International Institute of Refrigeration, 79(1): Fleming J, 2003, Carbon dioxide as the working fluid in heating and/or cooling systems. Bulletin of the International Institute of Refrigeration, 83(4):7 15. Kim M, Pettersen J, Bullard CW, 2004, Fundamental process and system design issues in CO 2 vapor compression systems. Progress in energy and combustion science, 30: Seminario Europeo sugli impianti a CO 2, 2005, Componenti per impianti a CO 2, Fornasieri E, Zilio C, Università di Padova, Italy. International Journal of Refrigeration, Special Issue, 2005, CO 2 as Working Fluid Theory and Applications, 28:8. Pérez-Segarra CD, Rigola J, Oliva A, 2003, Modeling and numerical simulation of the thermal and fluid dynamic behaviour of hermetic reciprocating compressors. Part 1: Theoretical basis. International Journal of Heating, Ventilating, Air-Conditioning and Refrigerating Research, 9(2): García-Valladares O, Pérez-Segarra, CD Rigola J, 2004, Numerical simulation of double-pipe condensers and evaporators. International Journal of Refrigeration, 27: Rigola J, Raush G, Pérez-Segarra, CD, Oliva, A, 2005, Numerical simulation and experimental validation of vapour compression refrigeration systems. Special emphasis on CO 2 trans-critical cycles, International Journal of Refrigeration, 28(8): Pérez-Segarra CD, Rigola J, Sòria M, Oliva A, Detailed thermodynamic characterization of hermetic reciprocating compressors. International Journal of Refrigeration, 28: , Raush G, Rigola J, Pérez-Segarra CD, Oliva A, Thermal and fluid dynamic behaviour of a trans-critical carbon dioxide small cooling system: Experimental investigation. In International Conference on Compressors and Their Systems. L15-C639/45 City University, London, UK Rigola J, Raush G, Ablanque N, Pérez-Segarra CD, Oliva A, Comparative analysis of R134a subcritical cycle vs. CO 2 trans-critical cycle. numerical study and experimental comparison. In IIR Gustav Lorentzen Natural Working Fluids Conference 3/A/10.10, University of Glasgow, UK, 2004.