The effects of non-condensable gases in domestic appliances

International Journal of Refrigeration 30 (2007) 19e27 www.elsevier.com/locate/ijrefrig The effects of non-condensable gases in domestic appliances Luca Cecchinatoa, Maurizio Dell’Evab, Ezio Fornasieria,*,1, Massimo Marcer b, Orlando Monegob, Claudio Zilioa a Dipartimento di Fisica Tecnica, Università degli Studi di Padova, via Venezia 1, I-35131 Padova, Italy b ACC Application Eng. Lab., via Salvatelli 4, I-32020 Mel (BL), Italy Received 14 December 2005; received in revised form 6 April 2006; accepted 12 April 2006 Available online 7 July 2006 Abstract It is well known that the presence of non-condensable gases inside a compression vapour refrigerating circuit introduces an additional thermal resistance at the condenser, which can significantly decrease the energy efficiency of the system. However, this problem so far has been investigated mainly for shell and tube condensers of large capacity and limited information is available on small systems, as is the case for household appliances where the internal volumes are extremely reduced and therefore a very small amount of non-condensable gas has large effect. Moreover, non-condensable gas behaves differently when condensation takes place outside tubes (shell and tube condensers) or inside tubes (condensers of small appliances); in the first case all heat transfer area is wrapped by a gas layer, whereas in the second case non-condensable gas is collected at the end of the tube. The effect of non-condensable gas in this work is experimentally investigated by injecting controlled amounts of air into a refrigerating circuit and by recording the thermal and electric variables during different modes of operation (steady state and cyclic running). The tested refrigerating circuits are part of two appliances on the market, a household refrigerator and a vertical freezer. The presence of non-condensable gas was found to spoil energy efficiency, since it brings about an increase in condensing pressure and a concomitant decrease in evaporating temperature, although larger liquid subcooling partially compensates for the first negative effects: the reason for this behaviour is the clogging action of bubbles of gaseous mixture (air and refrigerant vapour) that enter the capillary tube. Ó 2006 Elsevier Ltd and IIR. All rights reserved. Keywords: Domestic refrigerator; Experiment; Non-condensable gas; Injection; Air; Refrigerating circuit Effets des gaz non condensables dans les appareils domestiques Mots clés : Réfrigérateur domestique ; Expérimentation ; Gaz non condensable ; Injection ; Air ; Circuit frigorifique 1. Introduction * Corresponding author. Tel.: þ39 049 827 6878; fax: þ39 049 827 6896. E-mail address: [email protected] (E. Fornasieri). 1 Member of IIR Commission B2. 0140-7007/$35.00 Ó 2006 Elsevier Ltd and IIR. All rights reserved. doi:10.1016/j.ijrefrig.2006.04.002 The effect of the presence of non-condensable gases during the condensation process of a refrigerant is a topic extensively investigated (especially from a theoretical perspective) in the classical case of condensation outside 20 L. Cecchinato et al. / International Journal of Refrigeration 30 (2007) 19e27 circular tubes; the well-known result is that an additional thermal resistance arises, deriving from a diffusion process of the vapour inside the layer of non-condensable gas that tends to concentrate close to the condensing surface, being carried by the centripetal motion of the vapour toward the tube. From a different point of view, Nusselt theory explains this penalisation in terms of a decrease in the driving potential, due to the lower saturation temperature at the external interface of the falling film of the condensate, as an effect of the partial pressure of the non-condensable gas. It is manifest that a small fraction of non-condensable gases causes large damage, since they are not uniformly distributed inside the internal volume of the refrigerating systems, but gather close to the surface where condensation takes place. To get a deeper insight into this subject, the reader can refer to the wide review paper of Jensen [1] or to the specific sections of handbooks, such as the ones by Burghardt [2] and Webb [3]. As far as the present authors are aware, in the case of the condensation inside tubes, there are no contributions available in the open literature except from the present authors [4,5]. The reason for this is perhaps that this topic entails technological aspects that are not well known by scientists and academics, although they are of the greatest interest for the refrigeration industry. When a liquid receiver is present at the condenser outlet, as is recommended for predictable performance of the refrigerating circuit, all non-condensable gas accumulates inside this volume, since a liquid seal prevents the gas from escaping, while the vapour motion draws it inside this trap. This circumstance causes the last portion of the condenser tubes to be flooded so as to cause liquid subcooling. As a matter of fact, only adequate subcooling can prevent liquid from flashing when it expands entering the receiver, where the vapour partial pressure is lower than the total one by the amount pertaining to non-condensable gas; it must be borne in mind that vapour cannot enter the receiver, as it cannot escape from it nor can it condensate, since this vessel is practically adiabatic. A moderate amount of non-condensable gas is therefore not very detrimental to performance, since a certain amount of flooding of the condenser can be useful. The circuits investigated in this work are typical of a small capacity appliance and do not include a liquid receiver, as dictated by the common practice for refrigerating systems using a capillary tube as throttling device. The analysis of the effect of non-condensable gas on the operation of the system will be proposed in Section 4 where the experimental results will be discussed; the variation of the operational variables as a function of the mass fraction of non-condensable gas present inside the circuit shows clearly how this occurrence affects the performance of the system. To summarise, when condensation takes place inside tubes, the non-condensable gas is carried away from the heat transfer surfaces and consequently the detrimental effect does not stem from heat transfer penalisation, but from the clogging effect of bubbles of gaseous mixture (air and refrigerant vapour) that enter the capillary tube, the throttling device typically used for the considered systems. In this work two different household appliances on the market (a refrigerator and an upright freezer) were tested. The tests were performed at different fractions of air injected into the refrigerating circuit and according to two basic procedures concerning continuous running or cyclic running under oneoff control, dubbed ‘‘energy consumption test’’. 2. Experimental setup 2.1. The tested appliances The experimental analysis has been carried out on two domestic appliances. (1) A single door refrigerator (cooler) of 320 litres internal volume, R600a operated (see Fig. 1a). The main components are: e A condenser, made by a coil pipe bond to a louvered plate with 10.2 m total length and 91 cm3 internal volume, directly connected by a liquid line to the dryer, just before the capillary tube. e A reciprocating hermetic compressor, ACC HQT 55AA model, lubricated with mineral oil. e An evaporator embedded in the back wall of the internal compartment with 17.7 m total length of circuit and 527 cm3 internal volume. At the outlet of the evaporator the suction line is directly connected to the compressor, while in the original appliance the suction line has part of the capillary tube inside it, forming a sort of tube-in-tube heat exchanger. This modification was introduced to simplify the circuit and to make easier the analysis of the effect of the injected air; the length of the capillary tube was suitably reduced to compensate for the effect of the absence of heat transfer. Later tests, not reported here, have shown that the penalisation arising from non-condensable gas is even worse in the original appliance; in any case the reaction of the circuit to the presence of non-condensable gas is of the same kind. The circuit is charged with the nominal amount of refrigerant R600a (isobutane), 40 g (0.2 g). (2) An upright freezer of 243 litres internal volume, R600a operated (see Fig. 1b). The main components are: e A condenser, made by a coil pipe bond to a louvered plate with 12.70 m total length and 110 cm3 internal volume. After the condenser outlet, the line of the liquid refrigerant is bonded to the frame of the door, forming the heating circuit that prevents ice formation along the door gasket (this circuit is lacking in the appliance L. Cecchinato et al. / International Journal of Refrigeration 30 (2007) 19e27 21 Fig. 1. (a) The tested single door refrigerator. (b) The tested upright freezer. (c) The doping special pipe. (d) The doping pipe connected to the compressor. A); then the liquid line passes through a dryer and afterwards enters the capillary tube. e A reciprocating hermetic compressor, ACC HQT 90AA model, lubricated with mineral oil. e An evaporator with 29.7 m total length of tube and 1012 cm3 internal volume. The tube is bent into a series of horizontal coils with two arrays of straight wire welded on the opposite sides, forming the shelves for foodstuffs. At the outlet of the evaporator, still inside the freezer compartment, a small cylindrical vessel is inserted along the vapour line to collect and to retain any liquid amount escaping from the evaporator; after this vessel the suction line forms the tube-in-tube heat exchanger with the capillary tube inside it and then is connected to the compressor. The circuit is filled in with the nominal amount of refrigerant R600a (isobutane), 110 g (0.2 g). Both the compressors are controlled according to ON/ OFF mode using a conventional thermostat; the thermostat is bound to the evaporator plate in both the appliances. 2.2. The measuring system The two appliances were instrumented in order to measure their performance (i.e. the electrical power spent to keep the inside volume at the required temperature) and to determine the temperature profiles along the heat exchangers so as to make it possible to get information on the change in operating conditions caused by non-condensable gas. 22 L. Cecchinato et al. / International Journal of Refrigeration 30 (2007) 19e27 Coppereconstantan thermocouples (T type) were used as temperature sensors. The estimated accuracy of the entire system for temperature measurements is 0.36  C. Twenty-one and 10 thermocouples were thermally anchored on the pipe wall of the condensers, respectively in the case of the refrigerator and of the freezer, for tracing the temperature profiles; for the evaporators, 10 and 11 thermocouples, respectively were used to the same purpose; for the refrigerator and the freezer; 6 thermocouples were used to monitor the temperature of the compressor shell and adjacent pipes (bottom and top of the shell, suction line, discharge line). Since this research work was conducted at the laboratory of an industrial firm in the framework of a cooperation aimed at a technological investigation, measuring instruments and procedures were typical of industrial applications; of course the sensor temperature does not perfectly match the temperature of the refrigerant inside the tube, but the deviation is low enough to support the theoretical analysis and the experimental results were always consistent with the theory. For determining the temperature inside the compartment, different procedures were used for the refrigerator and for the freezer. According to UNI/EN/ISO 7371 [10], three thermocouples were put inside the refrigerator compartment, which was kept empty, while in the case of the freezer, test packages were put inside the foodstuffs compartment according to UNI/EN/ISO 5155 [6] and temperatures were measured by one thermocouple put in the middle of each of eight M packages (measurement packages), placed using the loading scheme stated by the manufacturer. Two thermocouples were put in the climatic room according to UNI/EN/ISO 5155 [6] and CECED [7]. Pressures were measured at the suction and discharge sides of the compressor. The accuracy of the transducer (FS ¼ 10 bar A for suction side and FS ¼ 20 bar A for discharge side) was 0.04% FS, as declared by the manufacturer. The accuracy of the entire system of pressure measurements is estimated as 0.01 bar A. Electrical parameters (absorbed power, current and applied voltage) were recorded through a power analyser with an estimated accuracy of the entire measurement system of 0.7 W for the power, 3 mA for the current and 0.3 V for the voltage. The tests were carried out inside a climatic room built according to UNI/EN/ISO 5155 [6] and CECED [7]. 3. Testing procedures 3.1. Continuous running test Continuous running tests were carried out conforming to ANSI/AHAM HRF-1-1988 [8]. To summarise, the compressor runs continuously (no thermostatic control) for at least 24 h until thermal equilibrium is established, when for 5 h the difference between the highest temperature and the lowest one recorded by each sensor is lower than 0.5  C and when there is no significant deviation of other important thermodynamic parameters, according to CECED [7]. Thereafter the data from the measuring sensors are acquired for 1 h at intervals of 20 s. The resulting outputs are the time-averaged values of the relevant variables. The ambient temperature was 32  C (0.5  C). 3.2. Energy consumption test For the energy consumption test, the appliance is tested at a fixed thermostat set point and ambient temperature of 25  C (0.2  C). The compressor runs cyclically for at least 24 h or until thermal equilibrium is established as specified in UNI/EN 153 [9] and in UNI/EN/ISO 7371 [10], for the refrigerator, and in UNI/EN/ISO 5155 [6], for the freezer. For the refrigerator the equilibrium is reached when for 24 h the measured temperature of the refrigerator compartment does not differ more than 0.5  C from the mean value [7], whereas for the freezer thermal equilibrium is assumed if two conditions are met: (i) for each of the M packages the maximum (as well as the minimum) values of temperature recorded during each operating cycle through 24 h fall inside an interval 0.5  C wide [7], (ii) no marked trend of deviation from the mean temperature during a 24-h period is verified [7]. Thereafter the data from the measuring sensors are acquired for 24 h at 20-s intervals. Typical test results are energy consumption expressed in kWh/d (accuracy: 0.02 kWh/d), on and off times of the compressor (accuracy: 0.13 s), run time percentage and the time-averaged values of suction and discharge pressure during the on time. The test is repeated with different thermostat settings for determining by linear interpolation the final results according to the reference test temperatures: 5  C as the mean temperature of the refrigerator compartment and 18  C as the temperature of the warmest M package for the freezer, as stated in EN 153 [9] and UNI/EN/ISO 7371 [10]. 3.3. Doping procedure To allow injecting a controlled amount of ‘‘doping’’ air, the service tube of the compressor was modified as follows. As shown in Fig. 1c and d, a special manifold fitted with five ‘‘T’’ joints was welded to the service tube, having the end closed by a Hansen fitting. The free ends of the mentioned ‘‘T’’ joints are closed by rubber plugs, accurately sealed on the tube, so as to permit injecting air with a delicate syringe (50 cm3 total capacity), piercing one of the rubber plugs. After the injection, the plug was resealed with a suitable adhesive to prevent any unwanted air infiltration. To assure the exact fraction of non-condensable gas (air) during the tests, before each doping operation the circuit was evacuated until a pressure lower than 0.2 mbar was reached and thereafter the evacuation pump was kept running again 23 L. Cecchinato et al. / International Journal of Refrigeration 30 (2007) 19e27 for about 8 h; then the circuit was charged with the required amount of refrigerant and the appliance was put inside the climatic room at 25  C (0.5  C) and 50% relative humidity, and switched on. After thermal equilibrium was established, while the compressor was running, the desired air volume (1.0 cm3) was injected as previously described. The mass of injected air is determined from the variation in the internal volume of the syringe, being known temperature and pressure inside it. The results of the tests, performed at 32  C ambient temperature, are reported in Table 1 and Figs. 2e5. The analysis of the condenser data shows a significant effect of the injected air on the operating conditions of this component where, increasingly with the molar fraction of air, for both the refrigerator and the freezer, the following can be observed: 4.1. Continuous running tests (1) increase in pressure at the discharge side of compressor; (2) increase in the condensed subcooling; (3) increase in the wall temperature in the saturated region of the condenser, according to pressure. These tests have been very useful to get an insight into the reasons why non-condensables affect the performance of the refrigerating circuit; an important contribution for understanding the effect of non-condensables is given by the temperature profiles at condenser and evaporator that point out the transfer of refrigerant charge between these components. The change in the operating conditions of the condenser clearly is the effect of liquid flooding that increases along the fraction of injected air. The cause of this behaviour can be ascribed to a clogging action on the capillary tube carried out by non-condensable gases: if we compare the reference case (no air) with the maximum molar fraction of air (1.46% for the refrigerator and 1.69% for the freezer), we must 4. Analysis of the experimental results Table 1 Continuous running tests results Single door refrigerator Air molar fraction % Air mass Room temperature (tout) Compartment mean temperature (tin) Suction pressure Discharge pressure Evaporation sat. temperature (at compressor suction) Condensation sat. temp. (at compressor discharge) Middle condenser temperature Condenser outlet temperature Dryer temperature Subcooling of liquid at dryer g  C  C bar A bar A  C  C  C  C  C  C Input power Input power/(tout  tin) W W/K Upright freezer Air molar fraction % Air mass Room temperature (tout) Compartment mean temperature (tin) Suction pressure Discharge pressure Evaporation sat. temperature (at compressor suction) Condensation sat. temp. (at compressor discharge) Middle condenser temperature Condenser outlet temperature Dryer temperature Subcooling of liquid at dryer g  C  C bar A bar A  C  C  C  C  C  C Input power Input power/(tout  tin) W W/K 0 0.29 0.59 1.17 1.46 0.00 32.0 7.4 0.71 6.4 20.7 47.4 45.5 41.4 40.8 6.6 0.06 32.1 6.6 0.66 6.5 22.0 47.9 45.6 37.4 36.9 11.0 0.12 32.1 6.9 0.69 6.9 21.0 50.3 47.2 35.8 35.3 15.0 0.24 32.0 6.1 0.68 7.4 21.3 53.3 47.3 33.2 33.2 20.1 0.29 31.9 4.1 0.65 7.6 22.4 54.2 46.8 33.3 33.3 20.9 59.5 1.51 57.7 1.49 59 1.51 59.4 1.56 58.6 1.63 0.00 0.43 0.85 1.27 1.69 0.00 32.3 30.6 0.39 5.73 33.7 42.9 42.4 41.5 40.3 2.6 0.24 32.4 31.4 0.39 6.37 33.7 47.1 43.8 39.2 32.1 15.0 0.47 32.3 31.6 0.40 7.37 33.1 53.0 44.4 35.6 29.0 24.0 0.71 32.3 31.2 0.43 8.45 31.6 58.8 43.6 34.8 28.8 30.0 0.94 32.4 30.9 0.47 9.69 29.7 64.8 43.3 34.9 29.1 35.7 67.3 1.07 68.1 1.07 70.4 1.10 74.0 1.17 79.0 1.25 24 L. Cecchinato et al. / International Journal of Refrigeration 30 (2007) 19e27 54 Wall Temperature [ºC] 51 48 45 42 39 36 0.00 0.29 0.58 1.17 1.46 33 30 0.0 1.3 2.6 5.1 3.8 7.7 6.4 8.9 10.2 Length [m] Fig. 2. Condenser wall temperatures in continuous running tests for the refrigerator. the considered system; as condensation goes on, the vapour quality decreases and the flow pattern from annular flow tends toward stratified flow, plug flow and finally bubble flow, but, unlike the case of condensation of a pure fluid, bubbles cannot collapse, because they contain an air/vapour mixture and condensation stops when the partial pressure of the refrigerant vapour inside the bubbles becomes equal to the saturation pressure of the surrounding liquid; thus, bubbles are carried inside the capillary tube and clog the flow. The strong temperature difference between the exit of the condenser and the dryer, occurring in the case of the freezer, deserves a comment; it is an effect of the heating circuit at the door, where the condensed liquid refrigerant comes into contact with a cold region. From the point of view of energy efficiency this solution is far more effective than the more usual scheme that uses, instead of the condensed liquid, the vapour coming from the compressor discharge to heat the door gasket. However, it is important to notice that the very low subcooling in the reference case of no air injected is evidence of a refrigerant charge lower than the notice that, even though subcooling is increased respectively from 6.6 to 19.9  C and from 2.6 to 35.7  C and the pressure difference at the compressor is increased from 5.7 to 7.0 and from 5.7 to 9.7 bar, the mass flow rate through the capillary tube is reduced, as can be inferred by increased refrigerating effect per mass unit, for being lower the liquid temperature at the dryer, and similar refrigerating power. As a matter of fact the refrigerating power depends primarily on the temperature difference between inside volume and outside environment and therefore is nearly constant (in the case of freezer), or slightly reduced (in the case of refrigerator). Please note that the subcooling values in Table 1 are conventional, referring to the saturation temperature at the condenser inlet. To explain how non-condensables choke the flow through a capillary tube, it is enough to recall the similar action performed by the presence of residual vapour at the capillary inlet, which results in lower mass flow rate, the more so as the vapour quality increases. Exactly the same effect occurs when non-condensable gases are present inside Wall Temperature [ºC] 18 12 0.00 0.29 6 0.58 1.17 0 1.46 -6 -12 -18 -24 -30 -0.3 1.7 3.7 5.7 7.7 9.7 11.7 13.7 15.7 Length [m] Fig. 3. Evaporator wall temperatures in continuous running tests for the refrigerator. 17.7 25 Wall Temperature [ºC] L. Cecchinato et al. / International Journal of Refrigeration 30 (2007) 19e27 80 76 72 68 64 60 56 52 48 44 40 36 32 28 24 0.00 0.43 0.85 1.27 1.69 0 1 2 3 4 5 6 7 8 9 10 11 12 Length [m] Fig. 4. Condenser wall temperatures in continuous running tests for the freezer. Instead, in the freezer case (see Fig. 5), the operating conditions at the evaporator are subjected to minor changes, resulting, increasingly with the molar fraction of air, in a slight increase of the pressure at the suction of the compressor (the opposite with respect to the freezer), while the superheating at the exit of the evaporator and therefore the dry area fraction seem essentially unaffected by the injected air. A possible explanation is the presence of the vessel at the outlet of the evaporator in which a reserve of refrigerant charge is present, contained inside a small cylindrical vessel, thus preventing starvation of the evaporator when refrigerant is transferred from the low pressure side of the circuit to the high pressure side. The small increase in evaporation pressure as the air fraction is increased can be explained by the variation in refrigerating capacity of the compressor as a function of the operating conditions. Suction pressure increases if the optimised value (or of a capillary tube too short): as a matter of fact, in this situation a partial condensation is occurring inside the line near the door frame. In general, flooding the condenser involves starving the evaporator and the change in operation of the evaporator during the refrigerator tests can be explained as a direct consequence of the decrease in wet area. In this case indeed, increasingly with the molar fraction of air, the following can be observed (see Fig. 2): (1) slight lowering of pressure at suction side of the compressor (except for the test with 0.29% air molar fraction); (2) increase in vapour superheating at the evaporator outlet; (3) lowering of the wall temperature in the saturated region of the evaporator, according to pressure. -16 -18 0.00 0.43 0.85 1.27 1.69 Wall Temperature [ºC] -20 -22 -24 -26 -28 -30 -32 -34 -36 -38 -40 0 2 4 6 8 10 12 14 16 18 20 22 24 Length [m] Fig. 5. Evaporator wall temperatures in continuous running tests for the freezer. 26 28 26 L. Cecchinato et al. / International Journal of Refrigeration 30 (2007) 19e27 refrigerating capacity of the compressor decreases; on the other hand the refrigerating capacity decreases as the condensation pressure increases, whereas it increases as the liquid temperature at the dryer lowers; so the experimental result derives from the combination of two opposite factors, but it is to be noticed that the variation in liquid temperature is small, being the state of the liquid determined by heat transfer at the door gasket. Energy efficiency is not directly connected only to the electrical power spent, since the mean temperature of the internal volume is not constant. For this reason the ratio of the electrical power to the temperature difference between outside and inside is assumed to be a representative value of energy efficiency. The results in Table 1 attest that energy efficiency decreases, during continuous running tests, as the fraction of non-condensable increases, but for molar fraction lower than 1.17 and 1.27 for the refrigerator and freezer respectively, the penalisation is scarcely detectable. 4.2. Energy consumption tests These tests were performed at 25  C ambient temperature. Here the energy consumption data obtained at different air fractions are directly comparable, since the internal temperature for the refrigerator and the maximum M-package temperature for the freezer are the same, being automatically adjusted by the thermostat through different running times of the compressor. For the definition of the time depending variables reference is made to Section 3.2. Table 2 offers a clear and direct representation of the effect of non-condensable gases on compressor run time, energy consumption and operating pressures. As far as the household refrigerator is concerned the results, in terms of penalisation in energy efficiency due to the presence of non-condensable gas, are in good agreement with the theoretical analysis shown in Section 4.1. It is worth noting that the effect of the injected air appears more detrimental during cyclic operation than in continuous running. For the freezer the penalisation is far lower and in some circumstances indeed a benefit appears, as for example for air fractions of 0.21% and 0.43%. At first glace this fact appears inexplicable, but a possible explanation exists and consists in bad design of the capillary tube. When the capillary tube is oversized, as it seems to be in this case, a penalisation arises for being the high pressure side starved of refrigerant so that the heating circuit at the door gasket acts as condenser and not as a subcooler Table 2 Energy consumption tests results Single door refrigerator Air molar fraction % Air mass Room temperature Compartment mean temperature Suction pressure Discharge pressure Evaporation temp. (comp. suction) Cond. temp. (comp. discharge) Energy consumption Energy consumption variation Run time Run time variation g  C  C bar A bar A  C  C kWh/day % % % Upright freezer Air molar fraction % Air mass Room temperature M-package maximum temperature Suction pressure Discharge pressure Evaporation temp. (comp. suction) Cond. temp. (comp. discharge) Energy consumption Energy consumption variation Run time Run time variation g  C  C bar A bar A  C  C kWh/day % % % 0.00 0.29 0.59 1.46 0.00 25.0 5.0 0.72 5.2 20.0 38.9 0.57 e 38.6 e 0.06 0.12 0.29 0.66 5.4 22.0 40.7 0.59 3.5 41.7 8.0 0.63 5.5 23.1 41.4 0.64 12.3 47.4 22.8 0.62 6.4 23.5 47.3 0.68 19.3 49.5 28.2 0.00 0.21 0.43 0.64 0.85 1.06 0.12 0.24 0.35 0.47 0.59 0.00 25.0 18.0 0.52 5.0 27.4 38.0 0.853 e 40.1 e 0.52 5.4 27.8 40.5 0.845 0.9 39.5 1.5 0.51 5.8 28.3 43.3 0.827 3.1 39.4 1.8 0.50 6.2 28.3 46.1 0.873 2.3 41.7 4.0 0.51 6.7 28.3 49.2 0.898 5.2 42.3 5.3 0.52 7.5 27.4 53.8 0.931 9.2 42.9 6.9 L. Cecchinato et al. / International Journal of Refrigeration 30 (2007) 19e27 (see the analysis in Section 4.1). When air is injected, the main effect is to lower the capacity of the capillary tube, so as to remedy the sizing error. 5. Conclusions Tests conducted on a household refrigerator and on an upright freezer have demonstrated that the effect of noncondensable gases on energy performance results not so much in worse heat transfer coefficient during condensation and evaporation, but in clogging action on the capillary tube. Unfortunately a capillary tube does not possess an intrinsic capability of controlling the mass flow rate and therefore, in general, the consequence is flooding of the condenser and starving of the evaporator, which are responsible for a degraded performance. The detrimental effect of non-condensable is ineluctable when the air fraction exceeds certain limits but, if the capillary tube is oversized, a limited amount of noncondensable can even improve the performance, in terms of refrigerating capacity and energy efficiency, because it can compensate for the excessive capacity of the throttling device. [2] [3] [4] [5] [6] [7] [8] [9] References [10] [1] M.K. Jensen, Condensation with noncondensables and in multicomponent mixtures, in: R.K. Shah, E.C. Subbarao, 27 R.A. Mashelkar (Eds.), Heat Transfer Equipment Design, Hemisphere Publishing Corp, 1988, pp. 497e512. A. Burghardt, Condensation of multicomponent mixture, VDI Heat Atlas, VDI Verlag, Dusseldorf, 1993, Sect. Jbb. D.R. Webb, Condensation of vapour mixtures, Heat Exchanger Design Handbook, Hemisphere Publishing Corp, 1995, Section 2.6.3. L. Cecchinato, M. Dell’Eva, E. Fornasieri, M. Marcer, C. Zilio, The effects of non-condensable gases in household refrigerators, in: Proc. 10th International Refrigeration and Air Conditioning Conference at Purdue, 2004. L. Cecchinato, M. Dell’Eva, E. Fornasieri, M. Marcer, O. Monego, C. Zilio, The effects of non-condensable gases in vertical freezers. IIR International Conference on Commercial Refrigeration, Vicenza, Italy, 30e31 August, 2005. UNI/EN/ISO 5155, Household refrigerating appliancesd Frozen food storage cabinets and food freezersdCharacteristics and test methods, Milan (Italy), 1999. CECED (European Committee of Domestic Equipment Manufacturers), Operational code for appliance testingdRefrigerators and Freezers, Brussels (Belgium), 2000. ANSI/AHAM, HRF-1e1988, American National Standard, Household refrigerators/Household Freezers, Chicago (USA), 1988. EN 153, Methods of measuring the energy consumption of electrics mains operated household refrigerators, frozen food storage cabinets, food freezers and their combination, together with associated characteristics, Brussels, 1997. UNI/EN/ISO 7371, Household refrigerating appliancesd Refrigerators with or without low-temperature compartmentd Characteristics and test methods, Milan (Italy), 2000.