Feasibility of C[O.sub.2] compressors for light commercial appliances.
The Montreal protocol (UNEP 1987) stipulated the phasing out of hydrochlorofluorocarbons (HCFCs) as refrigerants that deplete the ozone layer (ODP). Over the last few years, the Kyoto Protocol (UNFCCC 1997) has required the reduction of global warming potential (GWP). The European Union (EU) is obligated to reduce greenhouse gas emissions by 8% by the year 2010. The use of current fluid refrigerants--because of their ecological problems--needs an urgent solution. As an example, the European Climate Change Programme (ECCP 2001) indicates, as a likely scenario for Europe by the year 2010, a growth in direct emissions of around 19 Mton C[O.sub.2] equivalent and indirect emissions to be around 130 Mton C[O.sub.2] equivalent. This will equal 17% of the EU's total 2010 fluorinated gas emissions. Consequently, the investigation and use of new and/or natural refrigerants is an important goal.
Throughout the last decade, international research has indicated that the use of C[O.sub.2] has an important status as a natural fluid refrigerant (Lorentzen 1994; Jakobsen 1998; Kruse et al. 1999; Fleming 2003; Kim et al. 2004; Fornasieri and Zilio 2005; Rac Magazine 2006). It is nontoxic, nonflammable, unharmful, and has no impact on the environment. C[O.sub.2] is not only an attractive alternative but also a competitive solution. In that sense, R-744 has been the focus of research for use in many refrigeration applications, such as heat pumps (Mukaiyama 2002), air conditioners (Hafner et al. 2004), or any of several configurations of big commercial refrigeration systems (Eggen et al. 1998). All of these applications point to a promising future for R-744 use. This is why the world's leading companies in commercial refrigeration have followed this trend. A working group within the International Electrotechnical Commission (IEC 2002) recently prepared a draft proposal to modify IEC60335-2-34, which covers motor compressors using R-744 in transcritical applications. All these expectations confirm R-744 to be one of the alternatives for replacing hydrofluorocarbons (HFCs) in this sector.
The aim of this paper is to analyze in an objective and constructive way the feasibility of compressors for light commercial appliances working with R-744 in a transcritical cycle based on the development program carried out during the previous four years. The work presents a new compressor platform, designed to use R-744 as the fluid refrigerant, covering a wide range of cooling capacities from 200 to 1100 W at -10.0[degrees]C of evaporating temperature for high medium back pressure (HMBP) applications and from 200 to 800 W at -23.3[degrees]C of evaporating temperatures for low back pressure (LBP) applications.
The main difference between the R-744 platform designs in comparison with the current refrigerant compressor technology is the very high working pressure, which affects the housing design in terms of ensuring safety and preventing leakage to the ambient environment. The high-pressure difference between suction and discharge processes also affects piston-cylinder design to minimize piston force and flow leakage. Regarding this last issue, the use of piston rings is suggested (Suss and VeJe 2004). Another relevant difference is the high density of such refrigerant, which can be a constraint for compressors of small cooling capacity (i.e., lower than 1 kW approximately) because of small displacement requirements and, consequently, space limitations to allocate a reliable mechanism (cf. piston and piston-pin) and to allocate a reliable and efficient valve-valve plate system. The third difference, which contributes mainly to efficiency, is the variation of density versus temperature that is higher with R-744 than with current refrigerants (e.g., a compressor working with R-744 in a system with an evaporation temperature of -10[degrees]C and a superheat of 40 K shows an improvement in volumetric efficiency of ~13% when the superheat is reduced to 20 K but only ~8%-9% with R-134a or with R-404A). Less superheated inlet temperatures of the C[O.sub.2] compressor can also contribute to improve reliability of compressors and systems in transcritical cycles where high discharge pressures and, consequently, high discharge temperatures may be reached. All of these requirements take care of the internal suction heat transfer minimization in the new platform design.
It is difficult to exactly compare both subcritical cycles with standard compressors vs. transcritical cycles with new C[O.sub.2] compressor prototypes. It is well known that, unlike subcritical cycles, the gas cooler pressure in the transcritical cycle has an optimum value depending on working conditions. In fact, the transcritical results presented here are obtained with the optimal gas cooler pressure, while standard compressors are also presented in order to obtain a first approximation of the possibilities that C[O.sub.2] compressors offer.
After the theoretical analysis of the C[O.sub.2] behavior presented above and extensive study of the technical literature, this paper describes the advanced numerical simulation tools used to design and develop (1) C[O.sub.2] compressor laboratory prototypes, experimentally tested on transcritical cycle laboratory calorimeter tests, and (2) improved C[O.sub.2] compressor pre-industrial prototypes, experimentally tested on an instrumented freezer for LBP appliances.
Compressor samples were tested at several reference conditions and their performances were selected to the HFC's equivalent compressors for a similar cooling capacity under the same inlet compressor and inlet expansion valve temperatures. The numerical results present a reasonably good agreement against the laboratory test data and have allowed improving the successive designs and a first C[O.sub.2] compressor pre-industrial prototype.
The experimental results in the transcritical cycle experimental unit in both HMBP and LBP applications under the selected working conditions show the possibilities of cooling capacity and COP that these new C[O.sub.2] compressor prototypes offer. A definitive experimental comparison versus standard compressors working under subcritical conditions will need the exact definition of the equivalent transcritical cycle, which is currently difficult due to the differences explained above.
Further tests with the new platform were carried out on some commercial appliances to check the performance of the overall system and confirm the preliminary results obtained in the calorimeter tests. Finally, other considerations were taken into account and analyzed to give an overall vision of the feasibility of a compressor for light commercial appliances working with R-744.
ADVANCED NUMERICAL SIMULATION TOOLS
The basic idea of this section is to show and explain the different numerical tools used for developing, designing, and improving hermetic and/or semihermetic reciprocating compressors working with C[O.sub.2] as a fluid refrigerant integrated in a single-stage transcritical cycle for light commercial appliances. The numerical simulation tools were developed and used in order to minimize the number of prototypes and reduce the experimental laboratory tests, allowing acquisition of an optimum solution in terms of efficiency, reliability, and good performance in a more timely manner. The numerical simulation tools are designed to understand and optimize the key aspects of compressor and appliance development: (1) the thermal and fluid dynamic behavior of C[O.sub.2] in semihermetic compressors, (2) the single-stage transcritical cycle performance, (3) the critical aspects of suction and discharge valve design, and (4) the shell reliability optimization.
Thermal and Fluid Dynamic Compressor Behavior Numerical Simulation Code
The numerical simulation model of the thermal and fluid dynamic behavior of hermetic reciprocating compressors is extensively explained in detail (Perez-Segarra et al. 2003) and widely experimentally validated (Rigola et al. 2003). The last version of this numerical simulation code (CTTC 2005) was adapted to use the thermodynamic and transport properties of C[O.sub.2].
The numerical simulation model for the fluid flow 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). For any of the CVs in which the refrigerant gas is flowing, the semidiscretized governing equations take the following forms:
[[partial derivative]m/[partial derivative]t] + [summation][dot.m.sub.o] - [summation][dot.m.sub.i] = 0 (1)
[[partial derivative]m[bar.v]/[partial derivative]t] + [summation] [dot.m.sub.o] [v.sub.o] - [summation][dot.m.sub.i] [v.sub.i] = [F.sub.s] (2)
[[[partial derivative]m ([bar.h] + [bar.e.sub.c])]/[partial derivative]t] + [summation][dot.m.sub.o] ([h.sub.o] + [e.sub.co]) - [summation][dot.m.sub.i]([h.sub.i] + [e.sub.ci]) = V[[partial derivative][~.p]/[partial derivative]t] + [dot.Q.sub.w] (3)
The governing equations are discretized using an implicit control volume formulation and a SIMPLE-like algorithm extended to compressible flow. Effective flow areas through suction and discharge ports are evaluated considering a multidimensional approach based on a 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.). For each one of these macro-volumes, the following fully implicit discrete energy equation can be applied, considering convection between solid k and fluid i and both conduction and radiation between solid k and its neighboring solid j.
[[[[rho].sub.k.sup.n - 1] [c.sub.pk]([T.sub.k.sup.n] - [T.sub.k.sup.[n - 1]])[V.sub.k]]/[DELTA]t] = [summation over j][[[T.sub.k.sup.n] - [T.sub.j.sup.n]]/[R.sub.kj]][A.sub.kj] + [summation over i][dot.Q.sub.ki.sup.conv,n] + [summation over j][dot.Q.sub.kj.sup.rad,n] (4)
The model allows the possibility to obtain the detailed instantaneous evolution of pV diagrams, suction and discharge chamber pressure, and temperature evolution, together with suction and discharge valve displacement. The instantaneous motor torque, angular compressor velocity, and acceleration, together with instantaneous compression work or mass flow leakage, are also obtained.
Figure 1 presents some detailed illustrative results that highlight the differences between a standard R-134a compressor of 9.0 [cm.sup.3] for HMBP applications and a standard R-404A compressor of 6.0 [cm.sup.3] for LBP applications versus the new C[O.sub.2] compressor prototype of 1.5 [cm.sup.3] for both HMBP (evaporation temperature of 7.2[degrees]C) and LBP (evaporation temperature of -23.3[degrees]C) applications on pV diagram, mass flow leakage, or compression work. The comparative illustrative results were obtained, maintaining the same subcooling, superheating, and ambient conditions.
Single-Stage Vapor-Compression Transcritical Refrigerating Cycles Numerical Analysis
The numerical resolution consists of a main program composed of different subroutines. 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. A detailed description of the numerical model is referenced in Rigola et al. (2005). The objective of this numerical model is to validate the experimental results of the laboratory calorimeter test presented below. In this case the internal heat exchanger is not necessary due to the fact that both gas cooler and evaporator outlet conditions are controlled by secondary loops. Obviously, to validate the whole cycle in a real platform, it will be necessary to introduce this element.
[FIGURE 1 OMITTED]
The double-pipe heat exchangers 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 general correlations for heat transfer and fluid flow behavior were adapted to evaluate C[O.sub.2] as a fluid refrigerant.
The expansion device is evaluated in a similar way as the fluid refrigerant in the heat exchangers when a capillary tube is considered. In this case, it is necessary to take into account that 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 of a commercial valve as an expansion device, the numerical model is based on considering the flow through the valve as a sudden contraction along the tube: [dot.m] = [A.sub.D] x [C.sub.D] x [square root of (2[[rho].sub.i]([p.sub.i] - [p.sub.o]))].
The compressor inside the global cycle simulation is modeled 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 behavior is characterized by the following parameters: volumetric efficiency ([[eta].sub.v] = ([dot.m]/[[rho].sub.i])/[dot.G.sub.s.sup.[c = 0]]); isentropic efficiency ([[eta].sub.s] = [w.sub.s]/[w.sub.cp]); mechanical and electrical efficiency ([[eta].sub.me] = [w.sub.cp]/[w.sub.e]), and heat transfer loss efficiency ([[eta].sub.Qsh] = 1 - [[epsilon].sub.Qsh] = 1 - ([q.sub.sh]/[w.sub.e])). This last parameter represents the ratio between the heat transfer losses through the shell to the ambient versus the power consumption. The above-mentioned compressor parameters are functions 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 Perez-Segarra et al. (2005). In this paper, this information is obtained by means of the numerical simulation model hermetic reciprocating compressors presented above.
Figure 2 shows the differences between the numerical results and the experimental data on two C[O.sub.2] CL15 compressor prototypes and a standard R-134a compressor GLY80 model. A detailed description of compressor geometries is referenced in Rigola et al. (2005). The comparative results of the C[O.sub.2] transcritical cycle are presented under different evaporation temperatures of 7.2[degrees]C, 0[degrees]C, and -10.0[degrees]C, with an inlet fluid compressor temperature and outlet gas cooler temperature of 32[degrees]C. The comparative results of the GLY80 standard compressor working with R-134a are presented under the same evaporation temperature and the same boundary conditions as the C[O.sub.2] cycle, looking for the same comparative cooling capacity. The only difference between both C[O.sub.2] and R-134a cycles are the gas cooler pressure of the first at 85 bar and the condenser temperature of the second at 55[degrees]C. In all studied cases the ambient temperature is around 32[degrees]C. The results show a reasonably good agreement between the experimental data and the numerical data. Differences in mass flow rate are lower than 6%, while differences in power consumption are lower than 4% in all cases studied. The main differences take place on COP, with a maximum of 10% and lower end of 8%, depending on the cases studied.
Finite Element Analysis Suction and Discharge Valve Designs
The optimization of performances must always be within the limits of reliability. For this reason, structural analysis is carried out after a checkpoint of performances, which is reached by simulation according to the diagram shown in Figure 3.
[FIGURE 2 OMITTED]
[FIGURE 3 OMITTED]
Suction and discharge valve reliability was calculated through finite element analysis (FEA), applying a structural nonlinear analysis by means of a commercial code (ANSYS 2002). Even if bending fatigue or pressure fatigue is considered, it is simulated by the contact of elements--the abrupt change in stiffness that occurs when the valve begins contact with the stopper (bending fatigue) or contacts the valve-plate seat. A force value is applied on the valve on the point that coincides with the projection of the center of the valve plate suction/discharge port. This force is applied from zero to the maximum value in several steps in such a way that allows the software to simulate the true contact phenomena.
[FIGURE 4 OMITTED]
Adequate meshing has to be applied, dividing the domain into different areas in order to optimize computational time, as shown in Figure 4a. On the other hand, Figures 4b and 4c show an example of the detection of the contact surface for the suction valve just before the contact or during the flexion.
Finite Element Analysis Shell Optimization
When designing the new housing for an R-744 compressor, two kinds of nonlinear effects are present: an abrupt change of stiffness due to contact within the internal parts of the compressor, and plastic deformation when the maximum pressure, corresponding to that stated by present safety regulations, is applied.
For the new concept of platform, the internal volume is reduced and the shape is modified irregularly with no constant thickness. Simplification to more simple types of elements is not possible.
Due to all the above-mentioned points, computational cost at this point is very high, despite using the same FEM commercial code (ANSYS 2002). In order to reduce it, a symmetry technique was used and local concentrated meshing was applied (see Figures 5a and 5b). A different type of material compared to the conventional carbon steel commonly used for the hermetic reciprocating compressor shell was applied (see Figure 5c).
LABORATORY CALORIMETER TEST
An experimental setup was specially designed to evaluate the thermal and fluid dynamic behavior of C[O.sub.2] transcritical cycles and to validate the numerical simulation results of the one-stage reciprocating compressors. The experimental unit is made up of the following elements: a one-stage C[O.sub.2] compressor prototype, dual heat transfer coil gas cooler, and evaporator, together with a metering valve. A schematic representation and a general view of the refrigeration system are depicted in Figure 6.
The mass flowmeter, with a limit between 1 and 100 kg/h, guarantees an accuracy of [+ or -]0.047% of the span, with a repeatability and stability of [+ or -]0.015% nominal flow. Fluid flow refrigerant temperature sensors are K-type thermocouples with an accuracy of [+ or -]0.2[degrees]C, while secondary fluid flow temperature sensors are Pt100 thermoresistance with an accuracy of [+ or -]0.05[degrees]C. Pressure transducers with limits between 0-100 and 0-150 bars have an accuracy of [less than or equal to]0.05%. A detailed description about instrumentation accuracy and components geometry is referenced (Rigola et al. 2003, 2005).
[FIGURE 5 OMITTED]
[FIGURE 6 OMITTED]
The auxiliary fluid used in the gas cooler and the evaporator annuli is water. 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 flowmeters, with an accuracy of [+ or -]0.01 L/min from 0 to 2.5 L/min and [+ or -]0.5% F.S. from 2.5 to 25 L/min.
COMPRESSOR PLATFORM DESCRIPTION
Due to the order of magnitude of pressure ratios in comparison with the conventional fluid refrigerants, the compressor volume is reduced seven to eight times. The single compression stage platform is the one considered in this work. The range covered by this platform goes from 200 to 1100 W at a -10[degrees]C evaporating temperature for HMBP applications, and from 200 to 800 W at a -23.3[degrees]C evaporating temperature for LBP applications. This range is made up of seven displacements covered by the combination of two bore diameters. Its corresponding stroke is 1.42 cc (CL15TB). No piston rings are used.
[FIGURE 7 OMITTED]
Different versions of mechanisms were designed, and several life tests were conducted to arrive at a successful implementation of the definitive system.
Two versions of semihermetic housings were designed. The first generation, depicted in Figure 7a, was not an optimized prototype and was mainly used for investigation purposes and the collection of "know-how" data (compressor performance, mechanism life testing, sealing performance and its durability, etc.). After this step, and as a result of all the investigations carried out before, a second generation of housing, depicted in Figure 7b, was projected and built, complying with the latest version of the draft of the IEC60335-2-34 (IEC 2002) for motor compressors working with R-744 where the strength resistance test is proposed to be 286 bar.
Considering the higher pressure conditions and the new bore dimensions required due to small displacements for R-744, adequate suction and discharge valve reeds were designed to ensure safety, taking care to not penalize performances much. Safety design criteria currently used for standard refrigerants based on internal background and literature (Grolier 2000) were for this new refrigerant. The motor technology is the same as that currently used in our production facilities.
PLATFORM PERFORMANCE RESULTS
The compressor platform, described in the paragraph above, was numerically verified, experimentally validated, and widely tested in the laboratory calorimeter also described above. The results presented are the final part of different steps carried out before (Rigola et al. 2004, 2006; Raush et al. 2005).
In this section the graphical evolution of the volumetric and isentropic efficiencies are depicted for a reference cycle, the noise aspects are evaluated, and, finally, an experimental study of the new C[O.sub.2] compressor platform is shown for HMBP and LBP range working conditions. The results show the possibilities of this compressor in both ranges for light-commercial appliances.
In the following sections of this paper all compressors for HMBP and LBP are evaluated following the ASHRAE conditions detailed below. In the transcritical C[O.sub.2] cycles, inlet compressor temperature has an important influence on compressor efficiency (i.e., an internal heat exchanger is necessary to reduce outlet gas cooler temperature and inlet evaporation conditions), while gas cooler pressure gives an optimum value depending on working conditions (i.e., the gas cooler pressure is not the equivalent saturation pressure of condenser subcritical cycle temperature). These are the reasons why the transcritical cycles are evaluated slightly differently than subcritical cycles. Gas cooler pressure is obtained from optimization analysis (Rigola et al. 2006), while inlet compressor and outlet gas cooler temperatures are controlled in the laboratory calorimeter by the auxiliary fluid in both evaporator and gas cooler heat exchangers. Therefore, the last subsection presents the experimental results of standard compressors under subcritical cycle conditions with the aim to obtain an approximation of the possibilities that C[O.sub.2] can produce. An exact comparative analysis is not possible if both refrigeration cycles are not evaluated under the same system working conditions (refrigerating chamber temperature, ambient conditions, etc.).
Volumetric and Isentropic Efficiencies
Two of the more important nondimensional parameters that define the compressor behavior are the volumetric and isentropic efficiencies extensively described by Perez-Segarra et al. (2005). Most of these efficiencies are defined by comparison of the actual compressor behavior versus an "ideal" one.
It is assumed the ideal compressor has the same geometry (swept and clearance volume) and working conditions (inlet state, outlet pressure, and nominal frequency) as the actual one but operates reversibly according to adiabatic compression and expansion and to isobaric suction and discharge processes. The refrigerant gas is considered to behave as a perfect gas. To define this ideal or reference compressor, only a few parameters are needed. Specifically (1) inlet state and outlet pressure: [T.sub.i], [p.sub.i], [p.sub.o]; (2) type of fluid: r, [c.sub.p]; and (3) basic compressor characteristics: [V.sub.cl], c, [f.sub.n]. The volumetric flow rate pumped by the ideal (isentropic) compressor and the specific input work required can be easily obtained.
[dot.G.sub.S] = [V.sub.cl][1 - c([[product].sup.[1/[gamma]]] - 1)][f.sub.n] (5)
[w.sub.S] = r[T.sub.i][[gamma]/[[gamma] - 1]][[[product].sup.[([gamma] - 1)/[gamma]]] - 1] (6)
where [product] = [p.sub.o] / [p.sub.i] is the compression ratio, and [gamma] is the isentropic index [gamma] = [c.sub.p] / ([c.sub.p] - r) or adiabatic exponent. The specific heat is assumed as constant and evaluated from the real fluid thermophysical properties at the arithmetic mean suction/discharge ideal compressor temperatures and pressures.
Using the ideal compressor behavior determined, the volumetric efficiency is defined as the ratio of the actual volumetric flow rate at inlet conditions and the maximum one (ideal compressor without clearance volume c = 0): [[eta].sub.v] = ([dot.m]/[[rho].sub.i])[dot.G.sub.s.sup.[c = 0]]. In the same way, the isentropic (or compression) efficiency is defined as the ratio of the specific work delivered to the gas by the ideal compressor (with or without clearance volume c = 0) and the specific work delivered by the actual one, i.e., [[eta].sub.s] = [w.sub.s]/[w.sub.cp]. The mechanical-electrical efficiency [[eta].sub.me] defined above is necessary to obtain the ratio between the ideal work and the power consumption. Then the final parameter evaluated is: [[eta].sub.sme] = [w.sub.s] / [w.sub.e].
Due to the nature of the R-744 transcritical cycle, for ambient temperatures above 30[degrees]C it is going to be unavoidable to work at high discharge pressures so as not to lose cycle efficiency. For this reason the compressor has been tested at different discharge pressures for an evaporation temperature of -10.0[degrees]C and an inlet compressor temperature, outlet gas cooler temperature, and ambient temperature of 32[degrees]C. Figure 8 shows the illustrative experimental results of volumetric and isentropic efficiencies under the conditions described above.
Noise was measured on the compressor alone, not with the appliance. As is well known, the most important characteristics affecting the overall noise of the appliance are the sound power level, vibration level, and pulsation level. Regarding sound power levels (SPLs), Table 1 shows the preliminary results measured at 1 m of distance. SPL presents higher levels on a CL15 compressor prototype than the equivalent compressor in R-134a. Vibrations are equivalent to the existing platforms with traditional refrigerants. Spectrum measurement into the compressor chamber is especially important due to the fact that it is one of the most relevant sources of noise, and its smoothing reduction is the target to be considered (Raush et al. 2004). Gas pulsation is higher than current platforms, and some modifications must be implemented to achieve acceptable results. This work was already performed with success.
As a conclusion, two of the three main contributors to noise coming from the compressor are for the new platform inline with existing technology. The effect of the third one, the sound power level, will have to be confirmed by new experimental noise measurements on real appliances looking for the need--or not--of improvements.
[FIGURE 8 OMITTED]
Even if pressure conditions are not exactly the same, the results seem to confirm that the noise level is acceptable, considering that the target is to get close/similar levels to the current ones with standard refrigerants.
Results of C[O.sub.2] Compressor Prototype on Calorimeter Test
The present compressor platform has been designed to work with HMBP applications, to cover a cooling capacity between 200 and 1100 W at a -10[degrees]C evaporation temperature, and with LBP applications, to cover a cooling capacity between 200 and 800 W at a -23.3[degrees]C evaporation temperature.
Table 2 shows the experimental results on the laboratory calorimeter described above for the new compressor CL15TB also detailed under the specific calorimeter test working in transcritical cycle conditions.
The reference transcritical cycle selected for the new C[O.sub.2] compressor is defined by similar ASHRAE conditions at an inlet compressor temperature of 32[degrees]C and an outlet gas cooler temperature of 32[degrees]C, with a gas cooler pressure of 85 bar. This second set of working conditions does not have a standard test yet. Therefore, those proposed are the conditions that present the optimum COP values following the cooling capacities expected.
These results show an improvement of COP between 3% and 10% in comparison with the last C[O.sub.2] compressor prototype referenced (Rigola et al. 2005) for the same HMBP conditions. Results in Table 2 indicate that COP decreases when evaporation temperature decreases almost linearly, approximately 25%-30% between each one of the presented cases. It is interesting to highlight that COP decreases with the same slope along both HMBP and LBP working conditions. The prototypes tested before these results presented important decreases from HMBP to LBP evaporation temperatures.
Results of Standard Compressor Calorimeter Test
The aim of this section is to have the experimental information of the conventional compressors used on both studied applications, for standard compressors tested under the ASHRAE conditions, following ISO917 (ISO 1989), in order to have an approximation versus the new C[O.sub.2] compressor prototype that is not an experimental comparison.
Table 3 shows two compressor models of R-134a for HMBP applications. The first one, GP12TB, refers to a standard efficiency product, while the second one, GLY90RAb, refers to the current highest efficiency in the market of its cooling capacity level. Both compressors have been tested under a subcritical cycle with a condenser temperature of 55[degrees]C, an inlet compressor temperature of 35[degrees]C, and a subcooling of 9 K.
In addition to this, Table 3 also shows two more compressor models of R-404A for LBP applications. The first one, ML80FB, refers to a standard efficiency product, while the second one, MLY60LAb, refers to the current highest efficiency in the market of its cooling capacity level. Both compressors have been tested under a subcritical cycle with a condenser temperature of 55[degrees]C, an inlet compressor temperature of 32[degrees]C, and a subcooling of 23 K.
Results in Table 3 show that the COP of R-134a compressors for HMBP applications decrease around 15%-20% between each one of the tested cases, while the COP of the R-404A compressors for LBP applications decreases around 25%-30% between each one of the tested cases.
A direct comparison between the conventional compressors and the new C[O.sub.2] compressor platform due to the differences on working cycle conditions is not possible. Regarding efficiency, the new platform is expected to perform between the standard and the high efficiency levels on HMBP applications and should keep a good cooling capacity tendency with no important fall at low evaporating temperatures for LBP applications.
At this point all these results will have to be verified, only with the confirmation of real applications. At least the results are very promising and further optimization tasks can still be carried out.
RESULTS ON LBP APPLIANCE
In order to show the real behavior of the C[O.sub.2] compressor CL15TB under commercial light appliance conditions in general and LBP conditions in particular, a conventional freezer originally designed to work with R-134a and R-404A has been adapted and instrumented to use C[O.sub.2] as the fluid refrigerant (Jornet 2006). The aim of this investigation is to know the limits of this new platform and confirm the results obtained numerically with the models and experimentally with the calorimeter tests.
Figure 9 depicts the front and back freezer test with an inner volume of 258 [dm.sup.3] and shows the experimental freezer unit scheme instrumented with several pressure transducers and temperature probes. Figure 9 also indicates where a suction accumulation and a suction gas heat exchanger (SGHX) were added. Two expansion stages were studied in this appliance: a first stage with a discharge pressure control (DPC) and a second stage with a capillary tube. A pre-gas cooler has been implemented.
Table 4 shows the experimental results of the freezer tested at three different ambient temperatures of 25[degrees]C, 35[degrees]C, and 43[degrees]C, respectively. The main conclusions of the results obtained are: (1) at ambient temperatures over 30[degrees]C the discharge gas cooler pressure must increase to avoid a decrease in the performance of the system, which is already known and experimentally verified for a transcritical cycle with high efficiency dependence over system ambient temperatures due to an R-744 critical temperature of 31[degrees]C (Lozza et al. 2004); (2) at a 43[degrees]C ambient temperature, evaporating temperatures of -22[degrees]C and an averaged cooled room temperature of -15[degrees]C were reached; (3) at a 35[degrees]C ambient temperature, the conventional system with R-404A shows an 18% reduction of energy consumption over R-744; (4) the requirement of rather low temperatures in the cooled compartment (below -15[degrees]C) with ambient temperatures above 35[degrees]C highly penalizes the efficiency of the system at a low ambient temperature (25[degrees]C). Power input (W/K) means power consumption divided by the final temperature difference between ambient and the averaged temperatures of the interior of the freezer.
[FIGURE 9 OMITTED]
Thus, further improvements on the application can still be carried out as well as on the compressor, while it seems feasible to get the levels of standard efficiency with a single stage but difficult to reach the high efficiency levels.
OVERALL FEASIBILITY CONSIDERATIONS
When analyzing the feasibility of the new R-744 platform, one of the key aspects, apart from the others analyzed in this paper, will be the weight-cost relation.
Concerning weight, to have a comparative idea of this parameter, the weight of the R-744 platform is 16.2 kg, while the R-134a equivalent for a GP12TB is 12.1 kg. Thus, the R-744 platform is in a worse position compared to its R-134a equivalent.
Concerning cost, although no deep cost analysis was carried out, several aspects are already clear, which allows a rough approximation: (1) the new platform contains fewer parts than the existing designs, as inner volume was considerably reduced and the traditional compressor suspension system was changed in such a way that several parts were eliminated from the original HFC compressor; (2) the new housing will have a very big impact on final cost; and (3) the cost is expected to have less of an increase in the segment corresponding with the highest cooling capacities in comparison with the segment corresponding with the lower cooling capacities. Furthermore, the impact on cost will depend on R-744 remaining a marginal refrigerant or becoming the natural refrigerant of the future, replacing a significant amount of current HFCs.
With the aim of summarizing the contents of this paper and the overall feasibility considerations studied and examining all that is explained throughout the previous points, Table 5 shows a qualitative evaluation of all the key aspects that can influence the success of the R-744 platform consolidation, as analyzed in this paper.
A new platform design suitable to work with refrigerant R-744 was presented here. The advanced numerical simulation tools used to analyze, design, and optimize the compressor platforms were described, together with the details of the laboratory calorimeter test used to experimentally validate the compressor prototypes. Several aspects related to these performances were analyzed. The results obtained showed good agreement with the numerical analysis obtained and showed the possibilities against the standard ones. The new platform performs well for HMBP applications, although it needs to be improved for LBP applications so as not to penalize the TEWI. The noise of the new technology does not seem to be a hard constraint to make it feasible. The experimental results obtained under LBP appliance tests confirm the experimental results explained above, together with some new considerations depending on ambient temperature according with appliance boundary conditions. Finally, other parameters were analyzed and it seems to be clear that, at the end, the cost will depend to a great extent on the consolidation of this new refrigerant. Government implications will have great influence over this.
A = heat transfer area, [m.sup.2]
[A.sub.D] = expansion valve cross section, [m.sup.2]
c = clearance volume
[c.sub.p] = specific heat capacity at constant pressure, J x [kg.sub.-1] x [K.sup.-1]
[C.sub.D] = expansion valve flow coefficient
COP = coefficient of performance
[e.sub.c] = kinetic energy per mass unit, J x [kg.sup.-1]
[f.sub.n] = nominal frequency, Hz
[F.sub.s] = forces in flow direction, N
[dot.G.sub.s] = suction compressor volumetric flow, [m.sup.3] x [s.sup.-1]
[dot.G.sub.s.sup.[c = 0]] = suction compressor volumetric flow without clearance volume, [m.sup.3] x [s.sup.-1]
h = enthalpy per mass unit, J x [kg.sup.-1]
[dot.m] = mass flow rate, kg [s.sup.-1]
m = control volume mass, kg
p = pressure, Pa
[q.sub.sh] = heat compressor shell losses per mass unit, J x [kg.sup.-1]
[dot.Q] = input heat flux, W
[dot.Q.sub.ev] = cooling capacity, W
[dot.Q.sub.w] = wall heat exchanger flux, W
r = gas constant per mass unit, J x [kg.sup.-1] x [K.sup.-1]
R = contact thermal resistance, [m.sup.2] x K x [W.sup.-1]
t = time, s
T = temperature, K
[T.sub.ev] = evaporation temperature, K
v = flow velocity, m x [s.sup.-1]
V = volume, [m.sup.3]
[V.sub.cl] = cylinder volume, [m.sup.3]
[w.sub.cp] = compression work per mass unit, J x [kg.sup.-1]
[w.sub.e] = electrical work per mass unit, J x [kg.sup.-1]
[w.sub.s] = isentropic work per mass unit, J x [kg.sup.-1]
[gamma] = isentropic index
[DELTA]t = time step
[rho] = density, kg x [m.sup.-3]
[[epsilon].sub.Q.sub.sh] = deviation with respect to the reference case
[[eta].sub.me] = mechanical-electrical efficiency
[[eta].sub.Q.sub.sh] = compressor heat transfer shell losses efficiency
[[eta].sup.s] = isentropic efficiency
[[eta].sub.sme] = isentropic mechanical-electrical efficiency
[[eta].sub.v] = volumetric efficiency
[product] = pressure ratio
i = inlet section
o = outlet section
ki = fluid and solid neighbors
kj = solid neighbors
conv = convection
n = current time instant
rad = radiation
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A. Oliva, DrEng
C.D. Perez-Segarra, DrEng
J. Rigola, DrEng
Received October 6, 2006; accepted January 11, 2007
J. Jover is an electrical engineer and responsible for product planning, M. Jornet is an electrical engineer and applications engineering director, J. Pons in an electrical technical engineer and research and development laboratories director, and J.M. Serra is a mechanical engineer and research and development director at ACC Spain, SA. A. Oliva is a professor at the Technical University of Catalonia (UPC) and director of the Heat and Mass Transfer Technological Centre (CTTC) in the Mechanical and Aeronautical School of UPC, Barcelona, Spain. C.D. Perez-Segarra is a professor at UPC and a CTTC researcher. J. Rigola is an associate professor at UPC and a CTTC researcher. G. Raush is a PhD student at UPC and a CTTC researcher.
Table 1. Sound Pressure Level Preliminary Results Comparison Compressor Model GLY80 CL15TB Appliance cans number 378 480 Evaporation temperature -10[degrees]C 0[degrees]C Discharge pressure 14 bar 88 bar Noise level (SPL) (a) 59 dB 61 dB a. Sound pressure level measure at 1 m of distance. Table 2. Global Experimental Results of the New C[O.sub.2] Compressor Prototype under Transcritical Laboratory Test Conditions [T.sub.ev] ([degrees]C) +7.2 0.0 -0.0 -23.3 -35.0 [dot.Q.sub.ev] (W) 976 768 534 306 181 COP 2.69 2.04 1.45 1.04 0.72 Table 3. Global Experimental Results of the Standard Compressors under Subcritical Laboratory Test Conditions [T.sub.ev] GP12TB GLY90RAb ([degrees]C) [dot.Q.sub.ev] (W) COP [dot.Q.sub.ev](W) COP +7.2 1070 2.06 896 2.63 0 800 1.79 673 2.24 -10.0 494 1.41 425 1.73 [T.sub.ev] ML80FB* MLY60LAb* ([degrees]C) [dot.Q.sub.ev] (W) COP [dot.Q.sub.ev](W) COP -10.0 686 1.43 582 1.80 -23.3 370 1.09 326 1.36 -35.0 187 0.79 166 0.96 Table 4. Global Experimental Results under the LBP Appliance Test Ambient Temperature ([degrees]C) 25 35 45 Throttle (a) (b) (a) (b) (a) Evaporation temperature -36.4 -31.4 -29.3 -28.8 -22.3 ([degrees]C) Discharge pressure (bar) 76.5 83.6 86.9 91.6 110.9 Discharge temperature 71.4 74.1 88.8 91.5 109.6 ([degrees]C) Average cooled room ([degrees]C) -31.2 -26.8 -1.6 -22.4 -15.8 Power input (W/K) 4.94 5.84 8.83 5.61 6.57 Table 5. Overall Feasibility Picture R744 HMBP Platform R744 LBP Platform GWP +++ +++ EFFICIENCY* + - NOISE* ~ ~ WEIGHT - - COST - - +++ best, ++ good, + promising, ~ equivalent to current refrigerants, - bad *To be confirmed on appliance.
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|Author:||Jover, J.; Jornet, M.; Pons, J.; Serra, J.M.; Oliva, A.; Perez-Segarra, C.D.; Rigola, J.; Raush, G.|
|Publication:||HVAC & R Research|
|Date:||May 1, 2007|
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