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Airflow maldistribution and the performance of a packaged air conditioning unit evaporator
Applied Thermal Engineering 20 (2000) 515±528
www.elsevier.com/locate/apthermeng
Air¯ow maldistribution and the performance of a packaged air conditioning unit evaporator A.A. Aganda a, J.E.R. Coney b,*, C.G.W. Sheppard b a
Department of Mechanical Engineering, University of Nairobi, Nairobi, Kenya School of Mechanical Engineering, University of Leeds, Leeds LS2 9JT, UK
b
Received 15 September 1997; accepted 20 March 1999
Abstract The performance of an evaporator for a packaged air conditioning unit has been investigated. A heat transfer program ACOL5 validated in an earlier study, was used to predict the performance. Nonuniform velocity distribution measurements taken in a typical air conditioning unit were employed in the prediction of the evaporator performance. It was found that this maldistribution reduced the performance of an evaporator circuit, as compared to uniform ¯ow. Circuits at the edges of the evaporator, where the velocity was low, did not perform well. With the refrigerants controlled by one thermostatic valve, the worst performing circuit aected the performance of the whole evaporator, the evaporator performance being reduced by as much as 35%. The performance of the evaporator, where the circuits had dierent numbers of passes, depended on the position of the circuit in the evaporator. # 2000 Elsevier Science Ltd. All rights reserved. Keywords: Air conditioning; Evaporator; Air¯ow maldistribution
1. Introduction The design and the various assemblies of components in packaged air conditioning units often result in a considerable non-uniformity of the air ¯ow within them. These variants create diculties in developing computational procedures for predicting the performance of such units. One of the most important of the components aected by air ¯ow maldistributions is the evaporator, but there is a lack of published experimental data for non-uniform air¯ow on to * Corresponding author. Tel.: +113-233-2124; fax: +113-242-4611. 1359-4311/00/$ - see front matter # 2000 Elsevier Science Ltd. All rights reserved. PII: S 1 3 5 9 - 4 3 1 1 ( 9 9 ) 0 0 0 3 8 - 1
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such heat exchangers, although there have been contributions by Chwalowski et al. [1] and Timoney et al. [2]. The latter tests were performed on a 33 tube single circuit evaporator, in an open wind tunnel test section. A maldistribution on to the evaporator was created arti®cially, by mounting a perforated plate upstream of it, covering 50% of its area. The plate alternately covered the top and bottom of the evaporator, the air ¯ow being measured at a vertical plane in the centre of the duct. The experiments were conducted at mean air velocities of the order of 3±4 m/s and at air inlet temperatures close to 308C. Extant experimental work does not cover the range of complex ¯ow distributions present in packaged air conditioning units, which are likely to be two or even three dimensional. Further, the uneven loading of evaporator circuits, due to air maldistribution leads to maldistribution of the refrigerant. Hence, it is implied that poor performance of one circuit, due to poor air ¯ow on to it, would have an eect on others, having a good ¯ow distribution. Clearly, more data are required, both for understanding the eect of poor air ¯ow distribution and veri®cation of performance models with speci®c reference to evaporators used in packaged air-conditioning units. In a companion paper [3], a cross ¯ow heat exchanger program ACOL5, produced by the Heat Transfer and Fluid Flow Service (HTFS), [4] was validated using a single circuit multi pass heat exchanger with spatially uniform air¯ow on to the coil, for a wide range of conditions commonly prevailing in the evaporator of a packaged air conditioning unit. In this paper, a study of the eect of air ¯ow maldistribution on the performance of a multicircuit evaporator, using the ACOL5 program and a synthesis of single-circuit experimental data, is described. The EHA5 unit, produced by Airedale International Air Conditioning, was chosen for this purpose. Actual velocity measurements taken at a plane upstream of and normal to the evaporator of a typical packaged unit were used. Each evaporator circuit was considered independently, its performance being determined with respect to the air mean velocity and ¯ow distribution on to it. The performance of the evaporator was assumed to be the summation of the performances of each circuit. The eect of the air ¯ow maldistribution on the refrigerant condition is discussed and, the eect of multipass circuits on the prediction of evaporator performance is considered. 2. Apparatus 2.1. The Airedale International Air Conditioning EHA5 packaged air conditioning unit In Fig. 1, a sectional isometric view of this unit is shown. It is a typical self-contained unit in which all the necessary components for the air conditioning process are found. The main components include, proceeding from the inlet, a ®lter, an evaporator, a humidi®er, electrical heaters and a centrifugal fan, belt-driven by an electric motor. Included in the unit are small sensing devices, intended as safety controls. All such controls, including those for the electrical power supply, are housed together in a control box ®tted to the side of the unit downstream of the evaporator. Most of these items obstruct the ¯ow and aect its structure within the unit. The ®lter was a ®breglass cartridge, 50 mm thick, situated upstream of the evaporator at the
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Fig. 1. EHA5 Packaged Air Conditioning Unit Airdale. International Air Conditioning Ltd.
unit inlet section and normal to the air ¯ow. Located 400 mm downstream of the ®lter, the evaporator was 600 mm high and 610 mm wide and comprised a repeating interconnected system of copper tubes mechanically bonded to aluminium ®ns. It had a total of seven refrigerant circuits, 4±18 pass and 3±16 pass each having ®ve rows of tubes. It was manufactured from 9.52 mm (3/8 in.) inside diameter copper tubing and aluminium plate ®ns. The refrigerant ¯ow through it was controlled by a thermostatic valve and fed to the circuits through a distributer. The evaporator face did not span the whole cross-section of the unit and was oset to one side of it. Beneath the lower edge of the evaporator, there was a horizontal condensate tray positioned such that it extended upwards to cover some of the evaporator face, both upstream and downstream. Heating of the air, if necessary, could be provided by a row of U shaped, ®nned electrical heating elements, installed at the top of the rig, 50 mm downstream of the evaporator. Humidi®cation was provided by a cylindrical electrode boiler, mounted on the base of the unit and oset such that its vertical centre line coincided with that of the evaporator. The dimensions of the humidi®er were approximately 200 mm diameter and 300 mm in height; it was located some 260 mm downstream of the evaporator. The fan was a double entry centrifugal type, motor driven via a wedge belt and pulleys. The motor was positioned adjacent to and slightly forward of one of the inlet ports of the fan. The
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fan casing was approximately 740 mm downstream of the evaporator, its outlet ¯ange being connected to the outlet duct of the unit by ¯exible rubber tubing of square section. The incoming electrical power panel was situated on the side wall of the unit, just downstream of the coil. Using traversing gear, which could be positioned at any section along the unit, it was possible to obtain velocity contours, using constant-temperature, single-hotwire anenometry techniques, as described by Aganda et al. [3].
3. Previous investigation Park et al. [5] attempted to predict the performance of the EHA5 evaporator using the ACOL5 code. The purpose of this code is to simulate the performance of cross ¯ow heat exchangers, usually comprised of tube bundles through which a liquid, vapour or gas termed the process stream passes; a gas stream ¯ows over these bundles, usually normal to the tubes. The process stream is either cooled or heated by the gas stream. The tubes are usually ®nned to give an increased heat transfer area, thus enhancing the heat transfer between the two streams. Cross ¯ow heat exchangers, in common use, have a wide range of geometrical arrangements, which ACOL5 is able to accommodate. The correlations used for heat transfer and pressure drop calculations are varied also to accommodate the range of conditions likely to be encountered. The program allows various options of parameters to be varied during the simulation depending on the speci®c situation. The option parameters include inlet and outlet temperatures and also ¯ow rates of both process and gas streams. On the tube side, fouling resistance may be included. The code is described fully by Aganda et al. [3]. Also, details of the options available and the methods can be found in the ACOL5 user manual [4]. In the other stage of development, ACOL5 was unable to match exactly the experimental coil con®guration. Hence, the evaporator had to be modelled as having 3±15 pass and 4±20 pass circuits, instead of the actual 3±16 pass and 4±18 pass circuits, since the program required each circuit to contain the same number of tubes per row. It was suggested that this representative model performance would give a good indication of the performance of the actual evaporator. The investigation was made in three stages: 1. The performances of the 15-pass and 20-pass circuits were analysed separately, assuming uniform air distribution onto them. 2. The evaporator performance was predicted with non-uniform air distribution but with the assumption that the evaporator was made up only of either 15 or 20 pass circuits. The ¯ow distribution shown in Fig. 2 was used but with a normalised mean velocity of 2.55 m/s [6]. This distribution was obtained 50 mm from the upstream face of the evaporator. It covers the complete section of the air conditioning unit; the position of the evaporator is shown by a dotted line. The eect of the condensate tray, ®xed to the downstream side of the evaporator, is seen clearly. 3. The performance of the evaporator was considered with 3±15 pass and 4.20 pass circuits, having a non-uniform air distribution. From this study, it was found that air maldistribution on to a circuit lowered its heat transfer
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Fig. 2. Contours of air mean velocity in the direction of ¯ow at 50 mm upstream of the evaporator. Units 2 m/s.
eectiveness. For the same air ¯ow distributions, there was a major dierence between the performance of a 15-pass and 20-pass circuit. It was deduced also that the eect of any ¯ow maldistribution could be severe when the outlet refrigerant was not superheated since, this could aect the refrigerant ¯ow control by the thermostatic valve. When the refrigerant ¯ow through all circuits is controlled by one valve, even assuming uniform refrigerant ¯ow to each circuit, the overall evaporator performance may be aected profoundly, the circuits where the air ¯ow is most distorted, performing worst. When this study was made, the ACOL5 program prediction of air conditioning evaporator performance could not be veri®ed experimentally. Hence, the results had to be considered as preliminary and detailed estimates for, the eect of maldistribution on each circuit of the evaporator were not given. In addition, the evaporator circuitry had to be remodelled due to the inability of the program to predict the performance of passes having a dierent number of tubes. The work of Aganda et al. [3], involving heat transfer testing showed a good agreement between the test's results and the ACOL5 predictions for a single circuit evaporator. In this light, a more detailed application of the program was made with a degree of con®dence, considering an evaporator with the same tube arrangement as the single circuit experimental evaporator, but of the same size as that of the EHA5 unit. The air ¯ow distributions measured in the actual EHA5 unit were used for this purpose.
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4. Performance of an EHA5-size evaporator with 16 pass, 4 row circuits As stated earlier, the evaporator of the EHA5 unit had face dimensions 600 mm high 610 mm wide. The experimental evaporator described by Aganda et al. [3] was an 8 pass 4 row unit of face dimensions, 50 mm high 305 mm wide, as shown diagrammatically in Fig. 3a. An evaporator, having the same tube arrangement as this evaporator, but with 16 passes rather than 8, would be 100 mm high, as shown diagrammatically in Fig. 3b. A full size EHA5 evaporator could be synthesised from 6 of these component circuits, each 610 mm wide.
Fig. 3. Experimental and component evaporators (all dimensions in mm): (a) experimental evaporator: 8 pass, 4 row; (b) component evaporator: 16 pass, 4 row.
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Hence, for the purpose of the study, the original seven-circuit EHA5 evaporator was replaced by 6 of the 16 pass, 4 row circuits. Although the face dimensions of the original and the proposed evaporator were the same, the number of tubes and passes and also the tube arrangements were dierent. The eect of air ¯ow maldistribution on the proposed evaporator was investigated in two stages. Firstly, the eect of air ¯ow distribution on the performance of the evaporator circuits was considered. Here, it was assumed that the mean air velocity on to each circuit was the same but the ¯ow distribution on each circuit face varied from one another. Secondly, the eect of the actual air ¯ow distributions, as experienced by the full size evaporator face, was investigated. In this case, typical of practical applications, both the mean velocity and the face velocity distribution varied from one circuit to another. In both cases, velocity distributions measured in the EHA5 unit were used. The refrigerant and air ¯ow conditions were similar to the experimental conditions adopted by Aganda et al. [3]. viz., Evaporating pressure Evaporating temperature Refrigerant mass ¯ow rate Refrigerant inlet dryness fraction Air temperature on to evaporator Air relative humidity on to evaporator
598 kPA 5.38C 0.0159 kg/s 0.14 258C 50%
5. The eect of air ¯ow maldistribution on a single circuit In Fig. 4a and b the variations in air velocity are shown, normalised by the mean air velocity on to that circuit, across each of the six circuits of the evaporator for mean upstream velocities of 2.55 and 1.88 m/s, respectively. The mean air velocity approaching each circuit, U a , is stated against the relevant plot. These normalised air velocity distributions were used to predict the performance of each of the six 16-pass, 4-row circuits of the evaporator; it is a requirement of the ACOL5 program that each circuit is analysed separately. To determine the eect of the velocity distribution only, a mean velocity of 2.5 m/s was assumed for each circuit rather than the mean values indicated in Figs. 4a and b. The predicted performances of each circuit, with the common mean velocity of 2.5 m/s but with velocity variations as measured, are given in Table 1. If there were no velocity variations and, at each velocity measurement point the velocity was uniform at 2.5 m/s, all six circuits would have identical performances and the predicted refrigerant outlet temperature would be 18.38C. It should be emphasised that Table 1 shows the performance of each circuit, having the same mean velocity of 2.5 m/s but with the air ¯ow distribution of Figs. 4a and b. The purpose of this hypothetical case was to show that even if the mean velocity for each circuit was the same, the dierence in distribution would aect the performance of each. The refrigerant outlet temperatures thus obtained may be compared with those obtained by assuming a uniform velocity at all points of 2.5 m/s. The values, taking into account velocity variation given in
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Fig. 4. Variation of normalised air velocity across evaporator width: (a) mean upstream air velocity = 2.55 m/s; (b) mean upstream air velocity = 1.88 m/s.
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Fig. 4 (continued)
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Table 1 Single circuit: predicted refrigerant outlet temperature with air velocity variation across the circuita Circuit
Predicted refrigerant outlet temp (8C)
(a) Experimental mean air velocity = 2.55 m/s 6 5 4 3 2 1
18.0 17.9 17.9 18.1 18.1 17.9
(b) Experimental mean air velocity = 1.88 m/s 6 5 4 3 2 1
± 17.5 17.8 17.6 17.5 15.5
a Predicted refrigerant outlet temperature for a uniform air velocity of 2.5 m/s but with no air velocity variation across the circuits = 18.38C.
Table 1a, are lower than those for a uniform velocity across the evaporator while the predictions of Table 1b show larger dierences. However, although there was a general temperature decrease, due to velocity variation, the values were small. 6. The eect of air ¯ow maldistribution on the full-size evaporator face It was now possible to predict the performance of the six circuits, using the actual mean velocities, as given in Figs. 4a and b. The resultant refrigerant outlet temperatures are set out in Table 2 Referring to Table 2a, the actual mean velocity for the whole evaporator was 2.55 m/s. If this velocity was constant over the whole of the six circuits of the evaporator, the refrigerant temperature would be 18.58C. Similarly, for Table 2b, the corresponding values were 1.88 m/s and 10.68C. Thus, it is evident that air ¯ow maldistribution over the whole evaporator face results not only in non-uniformity for each circuit but also for the whole evaporator. From Table 2a, the two greatest decreases are seen to occur for Circuits 1 and 6. These two circuits were adjacent to the lower and upper edges respectively of the evaporator, where the mean velocities were lowest. However, of the remaining circuits, three showed temperature increase, the greatest being 1.38C for Circuit 5. For the lower mean velocity (Table 2b), the outlet temperature for Circuit 1 was the saturation value of 5.38C, the dryness fraction being 0.77. Other circuits, however, showed outlet temperatures higher than the desired average of 10.68C. Although only one circuit was underperforming, the presence of liquid refrigerant in the outlet manifold could aect the
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Table 2 Full-size evaporator: predicted refrigerant outlet temperature with air velocity variation Circuit
Mean air velocity U a (m/s)
Predicted refrigerant outlet temperature (8C)
(a) Experimental mean air velocity = 2.55 m/s (predicted refrigerant outlet temperature for a uniform air velocity of 2.55 m/s across all six circuits = 18.58C) 6 2.47 17.9 5 2.73 19.8 4 2.54 18.3 3 2.48 18.8 2 2.64 19.1 1 2.41 17.1 (b) Experimental mean air velocity =1.88 m/s (predicted refrigerant outlet temperature for a uniform air velocity of 1.88 m/s across all six circuits = 10.68C) 6 ± ± 5 2.06 13.4 4 2.09 14.0 3 1.94 11.2 2 2.11 13.8 1 1.20 5.3( = 0.77)
control, by the thermostatic valve, of the refrigerant ¯ow and so aect the performance of the whole evaporator. Such air ¯ow maldistribution, as discussed here, is often exacerbated by the additional maldistributions inherent in the design of packaged air conditioning units. These eects taken together lead to underperformance of evaporator edge circuits, with negative eects on the performance of the whole evaporator.
7. The eect of air ¯ow maldistribution on refrigerant ¯ow rate An evaporator thermostatic valve regulates the refrigerant mass ¯ow rate by sensing the temperature of the refrigerant at the outlet manifold. At a high outlet temperature, the valve increases the refrigerant ¯ow, and at a lower temperature, it reduces it. In a multi-circuit evaporator, each circuit having a dierent performance owing to varying air mean velocities and distributions, the circuit, with the lowest mean air velocity and in which the outlet refrigerant is in two-phase, will tend to ¯ood out ®rst [7]. This could cause the expansion valve to throttle the ¯ow to ensure that the refrigerant is superheated from the lowest performing circuit, thereby reducing the ¯ow through all the circuits and thus drastically lowering the evaporator performance. However, this assumes that the expansion valve is so placed that it will sense preferentially, the ¯ow from this circuit and also that the liquid refrigerant from it, has not achieved
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equilibrium with the superheated vapour from higher performing circuits. In practice, the valve might sense the conditions after equilibrium has been achieved, resulting from the mixture of liquid from the lower performing circuits and superheated vapour from the higher performing circuits. It is unlikely, therefore, that the valve would operate in such a way as to ensure superheated vapour leaving each circuit, which would cause a greater valve closure than should happen in reality. Nevertheless, it may be of interest to consider the former mode of operation, in order to give an indication of the worst possible eect of maldistribution. In Table 2b, the predicted outlet condition of the refrigerant from Circuit 1 is at the saturation temperature of 5.38C with a dryness fraction of 0.77. Here, the thermostatic valve would sense the outlet temperature as 5.38C and reduce the refrigerant mass ¯ow rate such that the outlet temperature from this circuit could be, for instance, 10.38C giving a superheat of 5 K. The refrigerant mass ¯ow would be reduced from 0.0159 to 0.011 kg/s, which might be applied to all circuits. The estimated capacity loss per circuit resulting from the reduction would be 0.86 kW with a total of 5.16 kW for the whole evaporator, causing a 35% loss in heat transfer performance. Therefore, although on an individual basis, other circuits would have a higher load, the poor performance of Circuit 1 would aect the whole evaporator. If Circuit 1, or any other circuit with a low air velocity had more passes, it would have a better performance. In support of this, Park et al. [5] found a major dierence between the performances of a 15 pass and 20 pass circuit.
8. The eect of multipass circuits on the prediction of evaporator performance with air ¯ow maldistribution Previously, the eect of poor air distribution was considered where all circuits had 16 passes. However, in practice, evaporators may consist of circuits with a variety of passes. Such con®gurations are bound to complicate the prediction of the eect of particular ¯ow distribution. The performance of Circuit 1, having 16 passes and 4 rows may be taken as an example (Table 2b). For a uniform mean air velocity of 1.3 m/s and refrigerant mass ¯ow rate of 0.011 kg/s, the refrigerant outlet temperature would be 12.48C. For a 20 pass 4-row circuit with the same tube geometry, and similar mass ¯ow rates of both refrigerant and air, the outlet temperature would be 14.78C. If the mean air velocity of 1.2 m/s were to be maintained on to the 20 pass circuit of a larger face area, the performance would be much enhanced. The circuit position within a maldistributed ¯ow would have an eect on the overall evaporator performance. Whereas a 16 pass circuit may have two phase refrigerant at outlet, a 20 pass circuit in the same position may deliver a superheat vapour. Therefore, it would be possible to have a 16 pass circuit with a relatively higher velocity performing less eectively than a 20 pass circuit with a lower velocity. Hence, the eect of poor air ¯ow distribution becomes less obvious and requires detailed analysis of every circuit. A particular air ¯ow distribution may have quite dierent eects on two multi-pass evaporators of same size but of dierent circuit arrangements. Maldistribution has a marked eect in systems where there is two-phase ¯ow. The heat
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transfer coecient between a two-phase refrigerant mixture and a tube wall can be shown to rise with dryness fraction, until a maximum is reached at a high dryness fraction (10.85). There ensues a rapid fall in heat transfer coecient until a very low value is reached at the dry saturated condition, which continues into the superheat region. Maldistribution tends to increase the tube area devoted to superheating the vapour in some circuits with a consequent reduction of heat transfer to the refrigerant.
9. Conclusions The present study has investigated the eect of air ¯ow maldistribution on to an evaporator using the ACOL5 program, the following conclusions may be drawn. . Both the present study and that of Park et al. [5], showed that velocity maldistribution had the eect of lowering the performance of an evaporator. However, the reduction for an individual single circuit, resulting only from air maldistribution was small. For the velocity distributions considered in this study, the superheat temperature drop in most of the circuits was less than 18C. This was in accord with the analytical study of Fagan [7] and the experimental measurements of Timoney et al. [2], where small changes were shown to occur in the single circuit evaporator performance due to air maldistribution. . When the analysis was extended to the full size multi-circuit evaporator where ¯ow maldistribution resulted in dierent mean velocities in addition to maldistribution for each circuit, there was a wide variation in predicted performance between the circuits. Those with higher mean velocity had a better performance compared with when ¯ow was uniform, while those with low mean velocity performed less eectively. . In multi-circuit evaporators where the refrigerant ¯ow control was by one thermostatic valve, if the outlet refrigerant from the less eective circuits were two-phase, and the throttle valve control sensor so placed as to react to this circuit, the valve could reduce the refrigerant mass ¯ow rate such that the outlet conditions of the refrigerant from the circuit with the poorest performance correspond to the set value (5±7 K superheat). The resultant reduced mass ¯ow rate of the refrigerant would lead to low performance of the high load circuits and therefore of the whole evaporator. It was shown that the eect of air maldistribution, which could lead to reduction in refrigerant mass ¯ow, caused a loss in the worst case of up to 38% in evaporator heat transfer performance. In practice, the throttle valve control sensor might be placed in a position where it would respond to the mixed output of all circuits, mitigating this eect to some extent. . The analysis of the eect of the air maldistribution is complicated further when the circuits have dierent numbers of passes. In the present case, it was found that the eect of a certain ¯ow maldistribution may depend on the circuit arrangement. . In the predictions, uniform distribution of the refrigerant between circuits was assumed. In practice, subdivision of the refrigerant between circuits might be imperfect and, even if it were not, since the pressure drop across each circuit must be the same, refrigerant ¯ow rates must vary if heat transfer diers from circuit to circuit.
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Acknowledgements The authors wish to acknowledge the contribution to this work by Airedale International Air Conditioning Ltd., Leeds and the Heat Transfer and Fluid Flow Service, National Engineering Laboratory, East Kilbride, Glasgow through the DTI/EPSCR Link Programme. HEPRACE. References [1] M. Chwalowski, D.A. Didion, P.A. Domanski, Veri®cation of evaporator computer models and analysis of performance of an evaporator coil, Trans. ASHRAE 95(1) (1989) 1229±1237. [2] D.J. Timoney, P.J. Foley, Some eects of air ¯ow maldistribution on the air performance of a compact evaporator with R134a. Eurotherm Seminar No. 26 (1993). [3] A.A. Aganda, J.E.R. Coney, P.E. Farrant, C.G.W., Sheppard, T. Wongwuttanasatian, A comparison of the experimental and predicted heat transfer performance of an evaporator coil circuit, Applied Thermal Engineering 20 (6) (2000) 499±513. [4] ACOL5 Program User Manual, Heat Transfer and Fluid Flow Service, National Engineering Laboratory, East Kilbride, Glasgow G75 OQU. [5] M. Park, P.E. Farrant, G. Hewitt, The eect of air maldistribution on heat transfer performance in air conditioning units. Heat Transfer and Fluid Flow Service, National Engineering Laboratory, East Kilbride, Glasgow G75 OQU, 1993. [6] A.A. Aganda, The eect of air-¯ow non uniformity on heat exchanger performance, with special reference to air conditioning units. Ph.D. thesis, Department of Mechanical Engineering, The University of Leeds, 1995. [7] T.J. Fagan, The eects of air ¯ow maldistributions on air-to-refrigerant heat exchanger performance, Trans. ASHRAE 86 (2) (1980) 699±713.
www.elsevier.com/locate/apthermeng
Air¯ow maldistribution and the performance of a packaged air conditioning unit evaporator A.A. Aganda a, J.E.R. Coney b,*, C.G.W. Sheppard b a
Department of Mechanical Engineering, University of Nairobi, Nairobi, Kenya School of Mechanical Engineering, University of Leeds, Leeds LS2 9JT, UK
b
Received 15 September 1997; accepted 20 March 1999
Abstract The performance of an evaporator for a packaged air conditioning unit has been investigated. A heat transfer program ACOL5 validated in an earlier study, was used to predict the performance. Nonuniform velocity distribution measurements taken in a typical air conditioning unit were employed in the prediction of the evaporator performance. It was found that this maldistribution reduced the performance of an evaporator circuit, as compared to uniform ¯ow. Circuits at the edges of the evaporator, where the velocity was low, did not perform well. With the refrigerants controlled by one thermostatic valve, the worst performing circuit aected the performance of the whole evaporator, the evaporator performance being reduced by as much as 35%. The performance of the evaporator, where the circuits had dierent numbers of passes, depended on the position of the circuit in the evaporator. # 2000 Elsevier Science Ltd. All rights reserved. Keywords: Air conditioning; Evaporator; Air¯ow maldistribution
1. Introduction The design and the various assemblies of components in packaged air conditioning units often result in a considerable non-uniformity of the air ¯ow within them. These variants create diculties in developing computational procedures for predicting the performance of such units. One of the most important of the components aected by air ¯ow maldistributions is the evaporator, but there is a lack of published experimental data for non-uniform air¯ow on to * Corresponding author. Tel.: +113-233-2124; fax: +113-242-4611. 1359-4311/00/$ - see front matter # 2000 Elsevier Science Ltd. All rights reserved. PII: S 1 3 5 9 - 4 3 1 1 ( 9 9 ) 0 0 0 3 8 - 1
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such heat exchangers, although there have been contributions by Chwalowski et al. [1] and Timoney et al. [2]. The latter tests were performed on a 33 tube single circuit evaporator, in an open wind tunnel test section. A maldistribution on to the evaporator was created arti®cially, by mounting a perforated plate upstream of it, covering 50% of its area. The plate alternately covered the top and bottom of the evaporator, the air ¯ow being measured at a vertical plane in the centre of the duct. The experiments were conducted at mean air velocities of the order of 3±4 m/s and at air inlet temperatures close to 308C. Extant experimental work does not cover the range of complex ¯ow distributions present in packaged air conditioning units, which are likely to be two or even three dimensional. Further, the uneven loading of evaporator circuits, due to air maldistribution leads to maldistribution of the refrigerant. Hence, it is implied that poor performance of one circuit, due to poor air ¯ow on to it, would have an eect on others, having a good ¯ow distribution. Clearly, more data are required, both for understanding the eect of poor air ¯ow distribution and veri®cation of performance models with speci®c reference to evaporators used in packaged air-conditioning units. In a companion paper [3], a cross ¯ow heat exchanger program ACOL5, produced by the Heat Transfer and Fluid Flow Service (HTFS), [4] was validated using a single circuit multi pass heat exchanger with spatially uniform air¯ow on to the coil, for a wide range of conditions commonly prevailing in the evaporator of a packaged air conditioning unit. In this paper, a study of the eect of air ¯ow maldistribution on the performance of a multicircuit evaporator, using the ACOL5 program and a synthesis of single-circuit experimental data, is described. The EHA5 unit, produced by Airedale International Air Conditioning, was chosen for this purpose. Actual velocity measurements taken at a plane upstream of and normal to the evaporator of a typical packaged unit were used. Each evaporator circuit was considered independently, its performance being determined with respect to the air mean velocity and ¯ow distribution on to it. The performance of the evaporator was assumed to be the summation of the performances of each circuit. The eect of the air ¯ow maldistribution on the refrigerant condition is discussed and, the eect of multipass circuits on the prediction of evaporator performance is considered. 2. Apparatus 2.1. The Airedale International Air Conditioning EHA5 packaged air conditioning unit In Fig. 1, a sectional isometric view of this unit is shown. It is a typical self-contained unit in which all the necessary components for the air conditioning process are found. The main components include, proceeding from the inlet, a ®lter, an evaporator, a humidi®er, electrical heaters and a centrifugal fan, belt-driven by an electric motor. Included in the unit are small sensing devices, intended as safety controls. All such controls, including those for the electrical power supply, are housed together in a control box ®tted to the side of the unit downstream of the evaporator. Most of these items obstruct the ¯ow and aect its structure within the unit. The ®lter was a ®breglass cartridge, 50 mm thick, situated upstream of the evaporator at the
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Fig. 1. EHA5 Packaged Air Conditioning Unit Airdale. International Air Conditioning Ltd.
unit inlet section and normal to the air ¯ow. Located 400 mm downstream of the ®lter, the evaporator was 600 mm high and 610 mm wide and comprised a repeating interconnected system of copper tubes mechanically bonded to aluminium ®ns. It had a total of seven refrigerant circuits, 4±18 pass and 3±16 pass each having ®ve rows of tubes. It was manufactured from 9.52 mm (3/8 in.) inside diameter copper tubing and aluminium plate ®ns. The refrigerant ¯ow through it was controlled by a thermostatic valve and fed to the circuits through a distributer. The evaporator face did not span the whole cross-section of the unit and was oset to one side of it. Beneath the lower edge of the evaporator, there was a horizontal condensate tray positioned such that it extended upwards to cover some of the evaporator face, both upstream and downstream. Heating of the air, if necessary, could be provided by a row of U shaped, ®nned electrical heating elements, installed at the top of the rig, 50 mm downstream of the evaporator. Humidi®cation was provided by a cylindrical electrode boiler, mounted on the base of the unit and oset such that its vertical centre line coincided with that of the evaporator. The dimensions of the humidi®er were approximately 200 mm diameter and 300 mm in height; it was located some 260 mm downstream of the evaporator. The fan was a double entry centrifugal type, motor driven via a wedge belt and pulleys. The motor was positioned adjacent to and slightly forward of one of the inlet ports of the fan. The
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fan casing was approximately 740 mm downstream of the evaporator, its outlet ¯ange being connected to the outlet duct of the unit by ¯exible rubber tubing of square section. The incoming electrical power panel was situated on the side wall of the unit, just downstream of the coil. Using traversing gear, which could be positioned at any section along the unit, it was possible to obtain velocity contours, using constant-temperature, single-hotwire anenometry techniques, as described by Aganda et al. [3].
3. Previous investigation Park et al. [5] attempted to predict the performance of the EHA5 evaporator using the ACOL5 code. The purpose of this code is to simulate the performance of cross ¯ow heat exchangers, usually comprised of tube bundles through which a liquid, vapour or gas termed the process stream passes; a gas stream ¯ows over these bundles, usually normal to the tubes. The process stream is either cooled or heated by the gas stream. The tubes are usually ®nned to give an increased heat transfer area, thus enhancing the heat transfer between the two streams. Cross ¯ow heat exchangers, in common use, have a wide range of geometrical arrangements, which ACOL5 is able to accommodate. The correlations used for heat transfer and pressure drop calculations are varied also to accommodate the range of conditions likely to be encountered. The program allows various options of parameters to be varied during the simulation depending on the speci®c situation. The option parameters include inlet and outlet temperatures and also ¯ow rates of both process and gas streams. On the tube side, fouling resistance may be included. The code is described fully by Aganda et al. [3]. Also, details of the options available and the methods can be found in the ACOL5 user manual [4]. In the other stage of development, ACOL5 was unable to match exactly the experimental coil con®guration. Hence, the evaporator had to be modelled as having 3±15 pass and 4±20 pass circuits, instead of the actual 3±16 pass and 4±18 pass circuits, since the program required each circuit to contain the same number of tubes per row. It was suggested that this representative model performance would give a good indication of the performance of the actual evaporator. The investigation was made in three stages: 1. The performances of the 15-pass and 20-pass circuits were analysed separately, assuming uniform air distribution onto them. 2. The evaporator performance was predicted with non-uniform air distribution but with the assumption that the evaporator was made up only of either 15 or 20 pass circuits. The ¯ow distribution shown in Fig. 2 was used but with a normalised mean velocity of 2.55 m/s [6]. This distribution was obtained 50 mm from the upstream face of the evaporator. It covers the complete section of the air conditioning unit; the position of the evaporator is shown by a dotted line. The eect of the condensate tray, ®xed to the downstream side of the evaporator, is seen clearly. 3. The performance of the evaporator was considered with 3±15 pass and 4.20 pass circuits, having a non-uniform air distribution. From this study, it was found that air maldistribution on to a circuit lowered its heat transfer
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Fig. 2. Contours of air mean velocity in the direction of ¯ow at 50 mm upstream of the evaporator. Units 2 m/s.
eectiveness. For the same air ¯ow distributions, there was a major dierence between the performance of a 15-pass and 20-pass circuit. It was deduced also that the eect of any ¯ow maldistribution could be severe when the outlet refrigerant was not superheated since, this could aect the refrigerant ¯ow control by the thermostatic valve. When the refrigerant ¯ow through all circuits is controlled by one valve, even assuming uniform refrigerant ¯ow to each circuit, the overall evaporator performance may be aected profoundly, the circuits where the air ¯ow is most distorted, performing worst. When this study was made, the ACOL5 program prediction of air conditioning evaporator performance could not be veri®ed experimentally. Hence, the results had to be considered as preliminary and detailed estimates for, the eect of maldistribution on each circuit of the evaporator were not given. In addition, the evaporator circuitry had to be remodelled due to the inability of the program to predict the performance of passes having a dierent number of tubes. The work of Aganda et al. [3], involving heat transfer testing showed a good agreement between the test's results and the ACOL5 predictions for a single circuit evaporator. In this light, a more detailed application of the program was made with a degree of con®dence, considering an evaporator with the same tube arrangement as the single circuit experimental evaporator, but of the same size as that of the EHA5 unit. The air ¯ow distributions measured in the actual EHA5 unit were used for this purpose.
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4. Performance of an EHA5-size evaporator with 16 pass, 4 row circuits As stated earlier, the evaporator of the EHA5 unit had face dimensions 600 mm high 610 mm wide. The experimental evaporator described by Aganda et al. [3] was an 8 pass 4 row unit of face dimensions, 50 mm high 305 mm wide, as shown diagrammatically in Fig. 3a. An evaporator, having the same tube arrangement as this evaporator, but with 16 passes rather than 8, would be 100 mm high, as shown diagrammatically in Fig. 3b. A full size EHA5 evaporator could be synthesised from 6 of these component circuits, each 610 mm wide.
Fig. 3. Experimental and component evaporators (all dimensions in mm): (a) experimental evaporator: 8 pass, 4 row; (b) component evaporator: 16 pass, 4 row.
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Hence, for the purpose of the study, the original seven-circuit EHA5 evaporator was replaced by 6 of the 16 pass, 4 row circuits. Although the face dimensions of the original and the proposed evaporator were the same, the number of tubes and passes and also the tube arrangements were dierent. The eect of air ¯ow maldistribution on the proposed evaporator was investigated in two stages. Firstly, the eect of air ¯ow distribution on the performance of the evaporator circuits was considered. Here, it was assumed that the mean air velocity on to each circuit was the same but the ¯ow distribution on each circuit face varied from one another. Secondly, the eect of the actual air ¯ow distributions, as experienced by the full size evaporator face, was investigated. In this case, typical of practical applications, both the mean velocity and the face velocity distribution varied from one circuit to another. In both cases, velocity distributions measured in the EHA5 unit were used. The refrigerant and air ¯ow conditions were similar to the experimental conditions adopted by Aganda et al. [3]. viz., Evaporating pressure Evaporating temperature Refrigerant mass ¯ow rate Refrigerant inlet dryness fraction Air temperature on to evaporator Air relative humidity on to evaporator
598 kPA 5.38C 0.0159 kg/s 0.14 258C 50%
5. The eect of air ¯ow maldistribution on a single circuit In Fig. 4a and b the variations in air velocity are shown, normalised by the mean air velocity on to that circuit, across each of the six circuits of the evaporator for mean upstream velocities of 2.55 and 1.88 m/s, respectively. The mean air velocity approaching each circuit, U a , is stated against the relevant plot. These normalised air velocity distributions were used to predict the performance of each of the six 16-pass, 4-row circuits of the evaporator; it is a requirement of the ACOL5 program that each circuit is analysed separately. To determine the eect of the velocity distribution only, a mean velocity of 2.5 m/s was assumed for each circuit rather than the mean values indicated in Figs. 4a and b. The predicted performances of each circuit, with the common mean velocity of 2.5 m/s but with velocity variations as measured, are given in Table 1. If there were no velocity variations and, at each velocity measurement point the velocity was uniform at 2.5 m/s, all six circuits would have identical performances and the predicted refrigerant outlet temperature would be 18.38C. It should be emphasised that Table 1 shows the performance of each circuit, having the same mean velocity of 2.5 m/s but with the air ¯ow distribution of Figs. 4a and b. The purpose of this hypothetical case was to show that even if the mean velocity for each circuit was the same, the dierence in distribution would aect the performance of each. The refrigerant outlet temperatures thus obtained may be compared with those obtained by assuming a uniform velocity at all points of 2.5 m/s. The values, taking into account velocity variation given in
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Fig. 4. Variation of normalised air velocity across evaporator width: (a) mean upstream air velocity = 2.55 m/s; (b) mean upstream air velocity = 1.88 m/s.
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Fig. 4 (continued)
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Table 1 Single circuit: predicted refrigerant outlet temperature with air velocity variation across the circuita Circuit
Predicted refrigerant outlet temp (8C)
(a) Experimental mean air velocity = 2.55 m/s 6 5 4 3 2 1
18.0 17.9 17.9 18.1 18.1 17.9
(b) Experimental mean air velocity = 1.88 m/s 6 5 4 3 2 1
± 17.5 17.8 17.6 17.5 15.5
a Predicted refrigerant outlet temperature for a uniform air velocity of 2.5 m/s but with no air velocity variation across the circuits = 18.38C.
Table 1a, are lower than those for a uniform velocity across the evaporator while the predictions of Table 1b show larger dierences. However, although there was a general temperature decrease, due to velocity variation, the values were small. 6. The eect of air ¯ow maldistribution on the full-size evaporator face It was now possible to predict the performance of the six circuits, using the actual mean velocities, as given in Figs. 4a and b. The resultant refrigerant outlet temperatures are set out in Table 2 Referring to Table 2a, the actual mean velocity for the whole evaporator was 2.55 m/s. If this velocity was constant over the whole of the six circuits of the evaporator, the refrigerant temperature would be 18.58C. Similarly, for Table 2b, the corresponding values were 1.88 m/s and 10.68C. Thus, it is evident that air ¯ow maldistribution over the whole evaporator face results not only in non-uniformity for each circuit but also for the whole evaporator. From Table 2a, the two greatest decreases are seen to occur for Circuits 1 and 6. These two circuits were adjacent to the lower and upper edges respectively of the evaporator, where the mean velocities were lowest. However, of the remaining circuits, three showed temperature increase, the greatest being 1.38C for Circuit 5. For the lower mean velocity (Table 2b), the outlet temperature for Circuit 1 was the saturation value of 5.38C, the dryness fraction being 0.77. Other circuits, however, showed outlet temperatures higher than the desired average of 10.68C. Although only one circuit was underperforming, the presence of liquid refrigerant in the outlet manifold could aect the
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Table 2 Full-size evaporator: predicted refrigerant outlet temperature with air velocity variation Circuit
Mean air velocity U a (m/s)
Predicted refrigerant outlet temperature (8C)
(a) Experimental mean air velocity = 2.55 m/s (predicted refrigerant outlet temperature for a uniform air velocity of 2.55 m/s across all six circuits = 18.58C) 6 2.47 17.9 5 2.73 19.8 4 2.54 18.3 3 2.48 18.8 2 2.64 19.1 1 2.41 17.1 (b) Experimental mean air velocity =1.88 m/s (predicted refrigerant outlet temperature for a uniform air velocity of 1.88 m/s across all six circuits = 10.68C) 6 ± ± 5 2.06 13.4 4 2.09 14.0 3 1.94 11.2 2 2.11 13.8 1 1.20 5.3( = 0.77)
control, by the thermostatic valve, of the refrigerant ¯ow and so aect the performance of the whole evaporator. Such air ¯ow maldistribution, as discussed here, is often exacerbated by the additional maldistributions inherent in the design of packaged air conditioning units. These eects taken together lead to underperformance of evaporator edge circuits, with negative eects on the performance of the whole evaporator.
7. The eect of air ¯ow maldistribution on refrigerant ¯ow rate An evaporator thermostatic valve regulates the refrigerant mass ¯ow rate by sensing the temperature of the refrigerant at the outlet manifold. At a high outlet temperature, the valve increases the refrigerant ¯ow, and at a lower temperature, it reduces it. In a multi-circuit evaporator, each circuit having a dierent performance owing to varying air mean velocities and distributions, the circuit, with the lowest mean air velocity and in which the outlet refrigerant is in two-phase, will tend to ¯ood out ®rst [7]. This could cause the expansion valve to throttle the ¯ow to ensure that the refrigerant is superheated from the lowest performing circuit, thereby reducing the ¯ow through all the circuits and thus drastically lowering the evaporator performance. However, this assumes that the expansion valve is so placed that it will sense preferentially, the ¯ow from this circuit and also that the liquid refrigerant from it, has not achieved
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equilibrium with the superheated vapour from higher performing circuits. In practice, the valve might sense the conditions after equilibrium has been achieved, resulting from the mixture of liquid from the lower performing circuits and superheated vapour from the higher performing circuits. It is unlikely, therefore, that the valve would operate in such a way as to ensure superheated vapour leaving each circuit, which would cause a greater valve closure than should happen in reality. Nevertheless, it may be of interest to consider the former mode of operation, in order to give an indication of the worst possible eect of maldistribution. In Table 2b, the predicted outlet condition of the refrigerant from Circuit 1 is at the saturation temperature of 5.38C with a dryness fraction of 0.77. Here, the thermostatic valve would sense the outlet temperature as 5.38C and reduce the refrigerant mass ¯ow rate such that the outlet temperature from this circuit could be, for instance, 10.38C giving a superheat of 5 K. The refrigerant mass ¯ow would be reduced from 0.0159 to 0.011 kg/s, which might be applied to all circuits. The estimated capacity loss per circuit resulting from the reduction would be 0.86 kW with a total of 5.16 kW for the whole evaporator, causing a 35% loss in heat transfer performance. Therefore, although on an individual basis, other circuits would have a higher load, the poor performance of Circuit 1 would aect the whole evaporator. If Circuit 1, or any other circuit with a low air velocity had more passes, it would have a better performance. In support of this, Park et al. [5] found a major dierence between the performances of a 15 pass and 20 pass circuit.
8. The eect of multipass circuits on the prediction of evaporator performance with air ¯ow maldistribution Previously, the eect of poor air distribution was considered where all circuits had 16 passes. However, in practice, evaporators may consist of circuits with a variety of passes. Such con®gurations are bound to complicate the prediction of the eect of particular ¯ow distribution. The performance of Circuit 1, having 16 passes and 4 rows may be taken as an example (Table 2b). For a uniform mean air velocity of 1.3 m/s and refrigerant mass ¯ow rate of 0.011 kg/s, the refrigerant outlet temperature would be 12.48C. For a 20 pass 4-row circuit with the same tube geometry, and similar mass ¯ow rates of both refrigerant and air, the outlet temperature would be 14.78C. If the mean air velocity of 1.2 m/s were to be maintained on to the 20 pass circuit of a larger face area, the performance would be much enhanced. The circuit position within a maldistributed ¯ow would have an eect on the overall evaporator performance. Whereas a 16 pass circuit may have two phase refrigerant at outlet, a 20 pass circuit in the same position may deliver a superheat vapour. Therefore, it would be possible to have a 16 pass circuit with a relatively higher velocity performing less eectively than a 20 pass circuit with a lower velocity. Hence, the eect of poor air ¯ow distribution becomes less obvious and requires detailed analysis of every circuit. A particular air ¯ow distribution may have quite dierent eects on two multi-pass evaporators of same size but of dierent circuit arrangements. Maldistribution has a marked eect in systems where there is two-phase ¯ow. The heat
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transfer coecient between a two-phase refrigerant mixture and a tube wall can be shown to rise with dryness fraction, until a maximum is reached at a high dryness fraction (10.85). There ensues a rapid fall in heat transfer coecient until a very low value is reached at the dry saturated condition, which continues into the superheat region. Maldistribution tends to increase the tube area devoted to superheating the vapour in some circuits with a consequent reduction of heat transfer to the refrigerant.
9. Conclusions The present study has investigated the eect of air ¯ow maldistribution on to an evaporator using the ACOL5 program, the following conclusions may be drawn. . Both the present study and that of Park et al. [5], showed that velocity maldistribution had the eect of lowering the performance of an evaporator. However, the reduction for an individual single circuit, resulting only from air maldistribution was small. For the velocity distributions considered in this study, the superheat temperature drop in most of the circuits was less than 18C. This was in accord with the analytical study of Fagan [7] and the experimental measurements of Timoney et al. [2], where small changes were shown to occur in the single circuit evaporator performance due to air maldistribution. . When the analysis was extended to the full size multi-circuit evaporator where ¯ow maldistribution resulted in dierent mean velocities in addition to maldistribution for each circuit, there was a wide variation in predicted performance between the circuits. Those with higher mean velocity had a better performance compared with when ¯ow was uniform, while those with low mean velocity performed less eectively. . In multi-circuit evaporators where the refrigerant ¯ow control was by one thermostatic valve, if the outlet refrigerant from the less eective circuits were two-phase, and the throttle valve control sensor so placed as to react to this circuit, the valve could reduce the refrigerant mass ¯ow rate such that the outlet conditions of the refrigerant from the circuit with the poorest performance correspond to the set value (5±7 K superheat). The resultant reduced mass ¯ow rate of the refrigerant would lead to low performance of the high load circuits and therefore of the whole evaporator. It was shown that the eect of air maldistribution, which could lead to reduction in refrigerant mass ¯ow, caused a loss in the worst case of up to 38% in evaporator heat transfer performance. In practice, the throttle valve control sensor might be placed in a position where it would respond to the mixed output of all circuits, mitigating this eect to some extent. . The analysis of the eect of the air maldistribution is complicated further when the circuits have dierent numbers of passes. In the present case, it was found that the eect of a certain ¯ow maldistribution may depend on the circuit arrangement. . In the predictions, uniform distribution of the refrigerant between circuits was assumed. In practice, subdivision of the refrigerant between circuits might be imperfect and, even if it were not, since the pressure drop across each circuit must be the same, refrigerant ¯ow rates must vary if heat transfer diers from circuit to circuit.
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Acknowledgements The authors wish to acknowledge the contribution to this work by Airedale International Air Conditioning Ltd., Leeds and the Heat Transfer and Fluid Flow Service, National Engineering Laboratory, East Kilbride, Glasgow through the DTI/EPSCR Link Programme. HEPRACE. References [1] M. Chwalowski, D.A. Didion, P.A. Domanski, Veri®cation of evaporator computer models and analysis of performance of an evaporator coil, Trans. ASHRAE 95(1) (1989) 1229±1237. [2] D.J. Timoney, P.J. Foley, Some eects of air ¯ow maldistribution on the air performance of a compact evaporator with R134a. Eurotherm Seminar No. 26 (1993). [3] A.A. Aganda, J.E.R. Coney, P.E. Farrant, C.G.W., Sheppard, T. Wongwuttanasatian, A comparison of the experimental and predicted heat transfer performance of an evaporator coil circuit, Applied Thermal Engineering 20 (6) (2000) 499±513. [4] ACOL5 Program User Manual, Heat Transfer and Fluid Flow Service, National Engineering Laboratory, East Kilbride, Glasgow G75 OQU. [5] M. Park, P.E. Farrant, G. Hewitt, The eect of air maldistribution on heat transfer performance in air conditioning units. Heat Transfer and Fluid Flow Service, National Engineering Laboratory, East Kilbride, Glasgow G75 OQU, 1993. [6] A.A. Aganda, The eect of air-¯ow non uniformity on heat exchanger performance, with special reference to air conditioning units. Ph.D. thesis, Department of Mechanical Engineering, The University of Leeds, 1995. [7] T.J. Fagan, The eects of air ¯ow maldistributions on air-to-refrigerant heat exchanger performance, Trans. ASHRAE 86 (2) (1980) 699±713.