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Effect of circuit arrangement on the performance of air-cooled condensers
International Journal of Refrigeration 22 (1999) 275–282
Effect of circuit arrangement on the performance of air-cooled condensers Chi-Chuan Wang a,*, Jiin-Yuh Jang b, Chien-Chang Lai b, Yu-Juei Chang a a
Energy and Resources Laboratories, Industrial Technology Research Institute, D500 ERL/ITRI, Bdlg. 64, 195-6 Section 4, Chung Hsing Road, Chutung, Hsinchu, Taiwan 310 b Department of Mechanical Engineering, National Cheng-Kung University, Tainan, Taiwan 701 Received 14 May 1998; received in revised form 20 October 1998; accepted 21 October 1998
Abstract An experimental study was carried out to investigate the effect of circuitry on the performance of wavy finned condensers. A total of eight arrangements were made and tested. The arrangements included six 1-circuit and two 2-circuit arrangements. For the one-circuit arrangement, the test results indicate that counter-cross flow would give better performance than other arrangements. However, heat conduction along the fins may offset the benefits of the counter-cross arrangement. This study has proposed two modifications to the counter-cross flow arrangement. For the two-circuit arrangement, a unique characteristic of “pressure gain” was observed when one circuit is completely condensed and the other is still in two-phase region. 䉷 1999 Elsevier Science Ltd and IIR. All rights reserved. Keywords: Air-cooled condenser; Refrigerant; Finned tube; Circuit
Effet de la disposition du circuit sur la performance des condenseurs refroidis par l’air Resume´ On a effectue´ une e´tude expe´rimentale afin d’e´tudier l’effet de la disposition du circuit sur la performance des condenseurs a` ailettes ondule´es. Huit dispositions ont e´te´ fabrique´es et e´tudie´es: six de ces dispositions comportaient un circuit et les deux autres dispositions avaient deux circuits. Les re´sultats obtenus avec les dispositions a` un circuit indiquent que le flux a` contrecourant sembler avoir une performance accrue par rapport aux autres dispositions. Cependant, la conduction de chaleur le long des ailettes pourrait re´duire l’avantage de la disposition a` contre-courant. Les auteurs propose deux modifications de la disposition a` contre-courant. Une augmentation de pression a e´te´ observe´e avec la disposition a` un circuit lorsqu’une condensation totale e´tait obtenue dans l’un des deux circuits pendant que l’autre circuit restait en re´gime diphasique. 䉷 1999 Elsevier Science Ltd and IIR. All rights reserved. Mots cle´s: Condenseur a` air; Frigorige`ne; Tube ailette´; Circuit
* Corresponding author. Tel.: ⫹886 3 5916294; fax: ⫹886 3 5820250. E-mail address: [email protected] (C.C. Wang)
Nomenclature A Di
0140-7007/99/$20.00 䉷 1999 Elsevier Science Ltd and IIR. All rights reserved. PII: S0140-700 7(98)00065-6
total surface area, m 2 inner diameter of tube, m
276
Do Fp g G H i if ifg N p Pd Pl Pt u Vfr T Tambient Tsat W x z DTm Dp Dpa Dpf Dpg Dpb a u rf rg
C.-C. Wang et al. / International Journal of Refrigeration 22 (1999) 275–282
outer diameter of tube, m fin pitch, mm gravitation constant, N m ⫺1 mass velocity, kg m ⫺2 s ⫺1 height of heat exchanger, m enthalpy at the exit of the condenser, kJ kg ⫺1 saturated enthalpy of the liquid refrigerant, kJ kg ⫺1 latent heat of the liquid refrigerant, kJ kg ⫺1 number of tube row pressure, MPa waffle height of the wavy fin pattern, mm longitudinal tube pitch, mm transverse tube pitch, mm heat transfer rate, W frontal velocity across coil, m s ⫺1 temperature, ⬚C ambient air temperature, ⬚C saturation temperature, ⬚C width of heat exchange, m vapor quality, dimensionless axial length along the direction of condenser, m mean temperature difference, ⬚C total pressure drop, Pa acceleration pressure drop, Pa frictional pressure drop, Pa gravitational pressure drop, Pa pressure drop by bends, Pa void fraction, dimensionless angle of inclination, degree liquid density of refrigerant, kg m ⫺3 vapor density of refrigerant, kg m ⫺3
1. Introduction Air-cooled heat exchangers used for air-conditioning and refrigeration applications generally consist of plate fin-andtube heat exchangers. These typically made with mechanically or hydraulically expanded round tubes into parallel continuous fins. Depending on the application, the heat exchangers can be made in one or more rows. To effectively improve the performance of air-cooled heat exchangers, passive enhancement techniques are often employed. The methods are
may be economical, but the degree of augmentation may not be very large. In practice, the most common method to increase the temperature difference is via circuitry. The main objective of the present study is focused on this effect, as applied to refrigerant condensers. The number of methods to circuit refrigerant in a heat exchanger are nearly unlimited. Te arrangements are limited by the manufacturing constraints. Usually, the inlet and outlet tubing occur on one side of the heat exchanger – denoted here as the active side of the coil. All solid bends or hairpins are on the opposite or passive side of the coil. A hairpin is the 180-degree bend created by bending a length of copper tubing into a “U” shaped configuration. The design of circuitry heavily relies on experience. There are some basic guidelines of arranging circuit as illustrated by Hogan [1]. However, detailed examination of the effect of circuitry is very rare in the open literature. In this connection, the objective of this study is to examine the effect of circuitry on the performance of a condenser.
2. Test setup Experiments were performed in an environmental chamber as shown in Fig. 1. The test apparatus is based on the airenthalpy method proposed by ANSI/ASHRAE Standard 37 [2]. Cooling capacity was measured from the enthalpy difference of the air flowrate across the test sample. The air flow measuring apparatus is constructed based on ASHRAE Standard [3] 41.2. Refrigerant R-22 was used as the working fluid. The test conditions are given as follows: • Ambient inlet air temperature: 25 ^ 0.3⬚C; • Saturation refrigerant temperature at the inlet of condenser: 45 ^ 0.2⬚C; and • Inlet superheat of the refrigerant flow: 5.5 ^ 1⬚C. A total of eight samples of fin-and-tube heat exchangers were made and tested in this study. The fin pattern for the present test sample is wavy fin, and the corresponding geometry is given as follows:
1. by using enhanced fin surfaces; 2. by increasing total surface area; and 3. by increasing the effective mean temperature difference between the air flow and refrigerant flow.
• Frontal area of the heat exchanger (W × H): 595 × 305 mm; • Longitudinal tube pitch (Pl): 19.05 mm; • Transverse tube pitch (Pt): 25.4 mm; • Waffle height (Pd): 1.18 mm; • Fin pitch (Fp): 1.7 mm; • Number of tube row (N): 2; • Nominal tube diameter (Do): 9.52 mm; and • Tube diameter after expansion (Dc): 10.24 mm.
The use of enhanced fin/tube may be very effective if the enhanced fin/tube is employed on the dominant resistance side. This also leads to possible increase of initial cost and some potential pitfalls (e.g. fouling and higher fan power). Increase of surface area may be again effective but may not be cost-effective. Increase of the effective mean temperature
A total of eight circuit arrangements were tested in this study (as shown in Fig. 2). Arrangements (A)–(F) are onecircuit arrangements, while (G) and (H) are two-circuit arrangements. Arrangements (A) and (C) are counter-cross flow, arrangements (B) and (D) are parallel-cross; and arrangements (E) and (F) are z-shape cross flow. Note that
C.-C. Wang et al. / International Journal of Refrigeration 22 (1999) 275–282
277
Fig. 1. Schematic of experimental setup. Fig. 1. Sche´ma de l’installation expe´rimentale.
the inlets for arrangements (A), (B), (E), and (H) are located at the upper portion while the inlets of arrangements (C), (D), and (F) are located at the lower portion of the test samples. In order to maintain the inlet state of the refrigerant flow, a refrigerant loop, a heating water flow loop, and a cooling water loop were provided. The refrigerant flow loop consists of a variable speed gear pump which delivers subcooled refrigerant to the preheater. The refrigerant pump can provide refrigerant mass fluxes ranging from 50 to 400 kg/ m 2s ⫺1. A very accurate mass flowmeter is installed between the refrigerant pump and the preheater. The accuracy of the mass flowmeter is 0.3% of the test span. The subcooled refrigerant liquid was superheated in the preheater to obtain a 5–6⬚ superheat before entering the test section. The system pressure of the refrigerant was maintained by another
separate cooling water loop. A pressure transducer having 10 Pa resolution was installed to measure the pressure drop across the test heat exchanger. The pressure taps of the test heat exchanger were located 450 mm upstream and downstream of the heat exchangers. Two absolute pressure transducers with resolution up to 0.1 kPa were installed at the inlet and exit of the test section. All of the water and refrigerant temperatures, were measured by RTDs (Pt100 V) having a calibrated accuracy of 0.05⬚C. All of the data signals were collected and converted by a data acquisition system (Hybrid recorder). The data acquisition system then transmits the converted signals through general purpose interface bus to a host computer for further operation. In this study, the physical and transport properties for R-22 are evaluated from computer program REFPROP [4].
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Fig. 2. Schematic of the circuit arrangement for the present study. Fig. 2. Sche´ma du circuit utilise´.
3. Results and discussion 3.1. Results of one-circuit arrangement
Fig. 3. Exit quality for arrangements (A)–(F). Fig. 3. Qualite´ sortie pour les dispositions (A) a` (F).
Fig. 3 shows the exit quality vs. frontal velocity for arrangementa (A)–(F) at Gin 200 kg m ⫺2 s ⫺1. Notice that the vapor quality is thermodynamic quality, and is given by x i ⫺ if =ifg : Therefore this value is negative for subcooled liquid. For a given arrangement (counter-cross arrangement, parallel-cross arrangement, and z-shape arrangement), the exit quality is relatively independent of the location of the refrigerant inlet position (upper or lower portion). However, it should be pointed out that the refrigerant-side pressure drops may be different owing to the effect of gravity. Table 1 presents the pressure drops for arrangements (E) and (F) at G 100 and 300 kg m ⫺2 s ⫺1. For Gin 100 kg m ⫺2 s ⫺1, it is seen that the pressure drops for arrangement (F) are considerably higher than those of arrangement (E). For a higher mass velocity of Gin 300 kg m ⫺2 s, the major contribution to the total pressure drop is by friction (approximately over 90%), therefore the pressure drops for arrangement (E) and (F) are comparable. As seen in Fig. 3, the exit quality for the counter-cross arrangements (A and C) is lower than that of parallel-cross flow arrangement (B and D). However, the z-shape
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279
Table 1 Refrigerant-side pressure drop for arrangements (E) and (F) Tableau 1 Perte de pression du coˆte´ frigorige`ne pour les dispositions (E) et (F) Frontal velocity (ms ⫺1)
Dp (kPa) G 100 kg m ⫺2 s ⫺1 (E) (F)
Dp (kPa) G 300 kg m ⫺2 s ⫺1 (E) (F)
0.7 1.0 1.5 2.0
7.1 6.5 6.1 5.9
33.3 28.9 27.9 25.8
7.7 7.8 7.8 7.9
arrangements (E) and (F) give the lowest exit quality. The results seem confusing since one would expect countercross arrangement would have a better performance than other arrangements. To explain this phenomenon, one may have to examine the detailed variations of the wall temperature along the condenser as illustrated in Fig. 4 (arrangements A and E). For G 100 kg m ⫺2 s ⫺1, one can clearly see that arrangement (A) is superior to arrangement (E) since the wall temperature drop for arrangement (A) occurs more quickly than arrangement (E). For instance, the wall temperature for arrangement (A) is approximately 7⬚C lower than that of arrangement (E) near z 4 m. The result substantiates that the counter-cross flow may have a higher heat transfer performance than the z-shape arrangement. However, the wall temperature for arrangement (A) does not consistently decrease along the condenser. Actually, the temperature may even increase when z ⬎ 10 m at G 100 kg m ⫺2 s ⫺1. Notice that the z-shape arrangement (E) does not show this phenomenon. For G 200 or
32.3 27.4 26.5 25.3
300 kg m ⫺2 s ⫺1, the test results also report similar phenomena. The wall temperatures for arrangement (A) drop more quickly than arrangement (E) but reaches an asymptotic value which is higher than that of arrangement (E) at the exit. Explanation of this phenomenon for arrangement (A) is because the temperature at the inlet portion (row 2, upper side) is much higher than that at the exit (row 1, upper side). As a result, larger temperature difference between the adjacent tubes may significantly increase the contribution of the heat conduction along the fin which is transferred back to the refrigerant. For the z-shape arrangement (E), the temperature difference between the neighboring tube is relatively small, therefore reversed heat transfer by conduction is negligible. The test results suggest that the counter-cross flow arrangement may not achieve the highest subcooling temperature. This may significantly reduce the system performance. To eliminate the reversed conduction effect and improve the performance of the counter-cross flow arrangement, possible solutions for improving the performance of the counter-cross arrangement are shown in Fig. 5. Fig. 5a has removed several redundant tubes at the exit. This may be very helpful from both heat transfer and economical point of view. Fig. 5b has used two separate one rows to form a two-row condenser. The tiny gap between these two rows may inhibit the heat conduction from the hot inlet portion back to the cold exit region. 3.2. Results of two-circuit arrangement
Fig. 4. Wall temperature variations for arrangements (A) and (E). Fig. 4. Variations des tempe´ratures des parois pour des dispositions (A) a` (E).
The previous results were for one-circuit arrangement. For two-circuit arrangement, refrigerant flow in each circuit may not be the same. Fig. 6 shows the wall temperature distributions of arrangement (G) for Gin 100, 200, and 300 kg m ⫺2 s ⫺1 at Vfr 1.0 m s ⫺1. It is interesting to see that significant differences in the wall temperature exist in circuit 1 and circuit 2. This is probably owing to the maldistribution of refrigerant flowrate. Note that the tube length for each circuit is the same, and the arrangement of these two refrigerant circuits is in a symmetrical style. Further, gravity effects would result in different pressure drops, even if the flowrates were equal. For the adiabatic test condition, it is obvious the refrigerant flowrate for both circuit should
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Fig. 5. Possible improvements for counter-cross arrangement. Fig. 5. Disposition a` contre-courant: ame´liorations possibles.
be the same. For the same axial distance and one of the circuits is fully condensed, the wall temperature for circuit 1 is higher than that of circuit 2. The test results imply that the mass flowrate for circuit 1 is higher than that of circuit 2. The difference in the mass flowrate between circuit 1 and 2 may be related to the gravity forces. As shown by Collier and Thome [5], the gravitational pressure gradient in the two-phase region is composed of vapor-phase contribution, g sin uar g, and liquid-phase contribution, g sin u (1 ⫺ a )r f,
Fig. 6. Wall temperature variations for arrangement (G). Fig. 6. Variations des tempe´ratures des parois pour la disposition (G).
i.e. ÿ
j k dp=dz g gsinu arg ⫹ 1 ⫺ arf :
1
For practical HVAC and R application, the liquid density is much higher than that of the vapor density. For an aircooled refrigerant condenser at high quality region, the flow pattern is annular at the initial stage of condensation (see the two-phase visualization of flow pattern by Wang et al. [6]). Thus the void fraction is small at this stage, therefore the gravity term is small compared to the total pressure gradient. For G 200 kg m ⫺2 s ⫺1, in the “up-portion” of circuit 1, the gravitational pressure drop is only about 4% ⬃ 5% of the total pressure drop in the “up-portion”. However, in the “down-portion” of circuit 1, the effect of gravity is reversed. Further, since the quantity of the liquid condensate in this portion is much larger than that of the “up-portion” in circuit 1. Hence the “pressure-gain” in this portion is very pronounced. An estimation of the “pressure-gain” caused by the gravitational gradient in the “down-portion” of circuit 1 is approximately 30% ⬃ 40% of the total gradient. Converse to refrigerant circuit 1, the sign of the gravitational gradient for circuit 2 in the “up-portion” is different from circuit 1 and is regarded as “pressure loss”. However, the total pressure drop is the same for circuits 1 and 2 after combination. Accordingly, circuit 1 is likely to have more mass flowrate in order to have a higher pressure drop. The test results by Kaga et al. [7] had also reported a similar characteristics for the two-circuit arrangement. This maldistribution of refrigerant flowrate may become more pronounced for lower mass flowrates. This is because the contribution of gravity force may become more significant.
C.-C. Wang et al. / International Journal of Refrigeration 22 (1999) 275–282
281 ⫺1
hereafter to 14 kPa for Vfr 0.4 m s . Analogous phenomenon was observed for Gin 200 kg m ⫺2 s ⫺1 near Vfr 0.4 m s ⫺1. Unlike the test results of the two-circuit arrangement, the one-circuit arrangements (A–F) do not reveal such kinds of characteristics as seen in Fig. 8. Possible explanation of the sudden drop of pressure drop is as follows. As known, the total pressure gradient within the condenser comprises acceleration loss, friction loss, gravity loss, and bend loss. i.e. Dp Dp Dp Dp Dp ⫹ ⫹ ⫹ : 2 Dz Dz a Dz f Dz g Dz b
Fig. 7. Pressure drops for arrangements (G) and (H). Fig. 7. Pertes de pression pour les dispositions (G) et (H).
3.3. Results of pressure drops Fig. 7 shows the pressure drops on the refrigerant-side vs. frontal velocity for arrangements G and H. A unique characteristics of this graph is that refrigerant pressure drop experiences a minimum value near the complete condensation point. For example, Gin 300 kg m ⫺2 s ⫺1, the pressure drop is approximately 12.8 kPa at Vfr 2.0 m s ⫺1 for arrangement (H). The pressure drop slightly increases to 13.1 kPa for Vfr 1.5 m s ⫺1. Notice that this value drops significantly to 9.2 kPa for Vfr 1.1 m s ⫺1 and increases
Of these four terms, the wall friction and the return bend loss Dpf and Dpb are always negative along the direction of refrigerant flow. The gravity loss, Dpg, is positive when refrigerant flow in the hairpin is moving downwards. The acceleration loss, Dpa, is negative for evaporation or boiling. However, for two-phase condensing flow, Dpa is negative since the quality is decreasing along the tube length. As a result, Dpa is in fact a “pressure gain” term. In practical airconditioning and refrigeration application, the frictional pressure drop Dpf would constitute most of the two-phase pressure drop. Jung and Radermacher [8] estimated the contribution of the acceleration term Dpa is always less than 10%. For the one-circuit arrangement (A–F), the contribution of this term is very small. For the present two-circuit arrangement, mal-distribution in the circuitry is likely as mentioned in Section 3.2. For some circumstances, it is very likely that one refrigerant circuit is completely condensed (subcooled) while the other circuit is still in the two-phase region. When these two circuits are combined, it is expected that direct-contact condensation would occur. This would significantly improve the heat transfer characteristics and may significantly increase the pressure gain of Dpa. Consequently, the pressure drop is suddenly decreased. Examinations of the present data indicate that the sudden decrease of pressure drop occurs near the point where average vapor quality is near zero. This further suggests that this phenomenon occurs only for one circuit in the saturated region while the other one is in the subcooled region. Note that this phenomenon is seen only for the two-circuit arrangement. The one-circuit design does not experience the sudden drop of pressure drop as seen in Fig. 8. 4. Conclusions An experimental study was carried out to investigate the effect of circuitry on the performance of a wavy finned condenser. A total of eight arrangements were made and tested. The arrangements included six 1-circuit and two 2circuit arrangements. On the basis of previous discussions, the following conclusions are made:
Fig. 8. Pressure drops for arrangements (A), (C), and (E). Fig. 8. Pertes de pression pour les dispositions (A), (C) et (E).
• The counter-cross arrangement gives better performance. However, the reversed heat conduction from
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the inlet portion to the exit portion may offset the benefit of counter-cross arrangement. Two modifications to improve the counter-cross arrangement are proposed in this study. • For the two-circuit arrangement, mal-distribution of the refrigerant flow is observed. The mal-distribution of the refrigerant flow is strongly influenced by gravitational force. • For the two-circuit arrangement, an unusual characteristics of the refrigerant pressure drop is observed. The refrigerant pressure drop experiences a sudden drop near x 0. This is because direct condensation occurs when these two circuits combine.
Acknowledgements The authors would like to express gratitude for the Energy R and D foundation funding from the Energy Commission of the Ministry of Economic Affairs, which provides financial supports of the current study. Suggestions and corrections by Prof. Ralph Webb is highly appreciated.
References [1] Hogan MR. The development of a low-temperature heat pump grain dryer. Ph.D. Thesis, Purdue University, 1980:116–127. [2] ANSI/ASHRAE standard 37-1988, Methods of testing for rating unitary air-conditioning and heat pump equipment. [3] ASHRAE, ASHRAE standard 41.2-1987 standard methods for laboratory air-flow measurement, 1987. [4] REFPROP 6.0, National Institute of Standards and Technology, Gaithersburg, MD, 1998. [5] Collier JG, Thome JR. Convective boiling and condensation. 3rd ed.. Oxford Science Publications, 1994:48. [6] Wang CC, Chiang CS, Lu DC. Visual observation of flow pattern of R-22, R-134a, and R-407C in a 6.5 mm smooth tube. Experimental Thermal and Fluid Science 1997; 15(4):395–405. [7] Kaga K, Yamada K, Takeshita S, Yamanaka MG. Improvement of a capacity of plate fin tubed heat exchanger by thermal analysis considering thermal interaction between pipes by heat conduction in fins. Proc. 10th Int. Heat Transfer Conf., Brayton, UK, The industrial Sections Paper (1994) paper no. I/2-CHE-8 99-104. [8] Jung DS, Radermacher R. Prediction of pressure drop during horizontal annular flow boiling of pure and mixed refrigerants. Int Heat Mass Transfer 1989;32(12):2435–2446.
Effect of circuit arrangement on the performance of air-cooled condensers Chi-Chuan Wang a,*, Jiin-Yuh Jang b, Chien-Chang Lai b, Yu-Juei Chang a a
Energy and Resources Laboratories, Industrial Technology Research Institute, D500 ERL/ITRI, Bdlg. 64, 195-6 Section 4, Chung Hsing Road, Chutung, Hsinchu, Taiwan 310 b Department of Mechanical Engineering, National Cheng-Kung University, Tainan, Taiwan 701 Received 14 May 1998; received in revised form 20 October 1998; accepted 21 October 1998
Abstract An experimental study was carried out to investigate the effect of circuitry on the performance of wavy finned condensers. A total of eight arrangements were made and tested. The arrangements included six 1-circuit and two 2-circuit arrangements. For the one-circuit arrangement, the test results indicate that counter-cross flow would give better performance than other arrangements. However, heat conduction along the fins may offset the benefits of the counter-cross arrangement. This study has proposed two modifications to the counter-cross flow arrangement. For the two-circuit arrangement, a unique characteristic of “pressure gain” was observed when one circuit is completely condensed and the other is still in two-phase region. 䉷 1999 Elsevier Science Ltd and IIR. All rights reserved. Keywords: Air-cooled condenser; Refrigerant; Finned tube; Circuit
Effet de la disposition du circuit sur la performance des condenseurs refroidis par l’air Resume´ On a effectue´ une e´tude expe´rimentale afin d’e´tudier l’effet de la disposition du circuit sur la performance des condenseurs a` ailettes ondule´es. Huit dispositions ont e´te´ fabrique´es et e´tudie´es: six de ces dispositions comportaient un circuit et les deux autres dispositions avaient deux circuits. Les re´sultats obtenus avec les dispositions a` un circuit indiquent que le flux a` contrecourant sembler avoir une performance accrue par rapport aux autres dispositions. Cependant, la conduction de chaleur le long des ailettes pourrait re´duire l’avantage de la disposition a` contre-courant. Les auteurs propose deux modifications de la disposition a` contre-courant. Une augmentation de pression a e´te´ observe´e avec la disposition a` un circuit lorsqu’une condensation totale e´tait obtenue dans l’un des deux circuits pendant que l’autre circuit restait en re´gime diphasique. 䉷 1999 Elsevier Science Ltd and IIR. All rights reserved. Mots cle´s: Condenseur a` air; Frigorige`ne; Tube ailette´; Circuit
* Corresponding author. Tel.: ⫹886 3 5916294; fax: ⫹886 3 5820250. E-mail address: [email protected] (C.C. Wang)
Nomenclature A Di
0140-7007/99/$20.00 䉷 1999 Elsevier Science Ltd and IIR. All rights reserved. PII: S0140-700 7(98)00065-6
total surface area, m 2 inner diameter of tube, m
276
Do Fp g G H i if ifg N p Pd Pl Pt u Vfr T Tambient Tsat W x z DTm Dp Dpa Dpf Dpg Dpb a u rf rg
C.-C. Wang et al. / International Journal of Refrigeration 22 (1999) 275–282
outer diameter of tube, m fin pitch, mm gravitation constant, N m ⫺1 mass velocity, kg m ⫺2 s ⫺1 height of heat exchanger, m enthalpy at the exit of the condenser, kJ kg ⫺1 saturated enthalpy of the liquid refrigerant, kJ kg ⫺1 latent heat of the liquid refrigerant, kJ kg ⫺1 number of tube row pressure, MPa waffle height of the wavy fin pattern, mm longitudinal tube pitch, mm transverse tube pitch, mm heat transfer rate, W frontal velocity across coil, m s ⫺1 temperature, ⬚C ambient air temperature, ⬚C saturation temperature, ⬚C width of heat exchange, m vapor quality, dimensionless axial length along the direction of condenser, m mean temperature difference, ⬚C total pressure drop, Pa acceleration pressure drop, Pa frictional pressure drop, Pa gravitational pressure drop, Pa pressure drop by bends, Pa void fraction, dimensionless angle of inclination, degree liquid density of refrigerant, kg m ⫺3 vapor density of refrigerant, kg m ⫺3
1. Introduction Air-cooled heat exchangers used for air-conditioning and refrigeration applications generally consist of plate fin-andtube heat exchangers. These typically made with mechanically or hydraulically expanded round tubes into parallel continuous fins. Depending on the application, the heat exchangers can be made in one or more rows. To effectively improve the performance of air-cooled heat exchangers, passive enhancement techniques are often employed. The methods are
may be economical, but the degree of augmentation may not be very large. In practice, the most common method to increase the temperature difference is via circuitry. The main objective of the present study is focused on this effect, as applied to refrigerant condensers. The number of methods to circuit refrigerant in a heat exchanger are nearly unlimited. Te arrangements are limited by the manufacturing constraints. Usually, the inlet and outlet tubing occur on one side of the heat exchanger – denoted here as the active side of the coil. All solid bends or hairpins are on the opposite or passive side of the coil. A hairpin is the 180-degree bend created by bending a length of copper tubing into a “U” shaped configuration. The design of circuitry heavily relies on experience. There are some basic guidelines of arranging circuit as illustrated by Hogan [1]. However, detailed examination of the effect of circuitry is very rare in the open literature. In this connection, the objective of this study is to examine the effect of circuitry on the performance of a condenser.
2. Test setup Experiments were performed in an environmental chamber as shown in Fig. 1. The test apparatus is based on the airenthalpy method proposed by ANSI/ASHRAE Standard 37 [2]. Cooling capacity was measured from the enthalpy difference of the air flowrate across the test sample. The air flow measuring apparatus is constructed based on ASHRAE Standard [3] 41.2. Refrigerant R-22 was used as the working fluid. The test conditions are given as follows: • Ambient inlet air temperature: 25 ^ 0.3⬚C; • Saturation refrigerant temperature at the inlet of condenser: 45 ^ 0.2⬚C; and • Inlet superheat of the refrigerant flow: 5.5 ^ 1⬚C. A total of eight samples of fin-and-tube heat exchangers were made and tested in this study. The fin pattern for the present test sample is wavy fin, and the corresponding geometry is given as follows:
1. by using enhanced fin surfaces; 2. by increasing total surface area; and 3. by increasing the effective mean temperature difference between the air flow and refrigerant flow.
• Frontal area of the heat exchanger (W × H): 595 × 305 mm; • Longitudinal tube pitch (Pl): 19.05 mm; • Transverse tube pitch (Pt): 25.4 mm; • Waffle height (Pd): 1.18 mm; • Fin pitch (Fp): 1.7 mm; • Number of tube row (N): 2; • Nominal tube diameter (Do): 9.52 mm; and • Tube diameter after expansion (Dc): 10.24 mm.
The use of enhanced fin/tube may be very effective if the enhanced fin/tube is employed on the dominant resistance side. This also leads to possible increase of initial cost and some potential pitfalls (e.g. fouling and higher fan power). Increase of surface area may be again effective but may not be cost-effective. Increase of the effective mean temperature
A total of eight circuit arrangements were tested in this study (as shown in Fig. 2). Arrangements (A)–(F) are onecircuit arrangements, while (G) and (H) are two-circuit arrangements. Arrangements (A) and (C) are counter-cross flow, arrangements (B) and (D) are parallel-cross; and arrangements (E) and (F) are z-shape cross flow. Note that
C.-C. Wang et al. / International Journal of Refrigeration 22 (1999) 275–282
277
Fig. 1. Schematic of experimental setup. Fig. 1. Sche´ma de l’installation expe´rimentale.
the inlets for arrangements (A), (B), (E), and (H) are located at the upper portion while the inlets of arrangements (C), (D), and (F) are located at the lower portion of the test samples. In order to maintain the inlet state of the refrigerant flow, a refrigerant loop, a heating water flow loop, and a cooling water loop were provided. The refrigerant flow loop consists of a variable speed gear pump which delivers subcooled refrigerant to the preheater. The refrigerant pump can provide refrigerant mass fluxes ranging from 50 to 400 kg/ m 2s ⫺1. A very accurate mass flowmeter is installed between the refrigerant pump and the preheater. The accuracy of the mass flowmeter is 0.3% of the test span. The subcooled refrigerant liquid was superheated in the preheater to obtain a 5–6⬚ superheat before entering the test section. The system pressure of the refrigerant was maintained by another
separate cooling water loop. A pressure transducer having 10 Pa resolution was installed to measure the pressure drop across the test heat exchanger. The pressure taps of the test heat exchanger were located 450 mm upstream and downstream of the heat exchangers. Two absolute pressure transducers with resolution up to 0.1 kPa were installed at the inlet and exit of the test section. All of the water and refrigerant temperatures, were measured by RTDs (Pt100 V) having a calibrated accuracy of 0.05⬚C. All of the data signals were collected and converted by a data acquisition system (Hybrid recorder). The data acquisition system then transmits the converted signals through general purpose interface bus to a host computer for further operation. In this study, the physical and transport properties for R-22 are evaluated from computer program REFPROP [4].
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Fig. 2. Schematic of the circuit arrangement for the present study. Fig. 2. Sche´ma du circuit utilise´.
3. Results and discussion 3.1. Results of one-circuit arrangement
Fig. 3. Exit quality for arrangements (A)–(F). Fig. 3. Qualite´ sortie pour les dispositions (A) a` (F).
Fig. 3 shows the exit quality vs. frontal velocity for arrangementa (A)–(F) at Gin 200 kg m ⫺2 s ⫺1. Notice that the vapor quality is thermodynamic quality, and is given by x i ⫺ if =ifg : Therefore this value is negative for subcooled liquid. For a given arrangement (counter-cross arrangement, parallel-cross arrangement, and z-shape arrangement), the exit quality is relatively independent of the location of the refrigerant inlet position (upper or lower portion). However, it should be pointed out that the refrigerant-side pressure drops may be different owing to the effect of gravity. Table 1 presents the pressure drops for arrangements (E) and (F) at G 100 and 300 kg m ⫺2 s ⫺1. For Gin 100 kg m ⫺2 s ⫺1, it is seen that the pressure drops for arrangement (F) are considerably higher than those of arrangement (E). For a higher mass velocity of Gin 300 kg m ⫺2 s, the major contribution to the total pressure drop is by friction (approximately over 90%), therefore the pressure drops for arrangement (E) and (F) are comparable. As seen in Fig. 3, the exit quality for the counter-cross arrangements (A and C) is lower than that of parallel-cross flow arrangement (B and D). However, the z-shape
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Table 1 Refrigerant-side pressure drop for arrangements (E) and (F) Tableau 1 Perte de pression du coˆte´ frigorige`ne pour les dispositions (E) et (F) Frontal velocity (ms ⫺1)
Dp (kPa) G 100 kg m ⫺2 s ⫺1 (E) (F)
Dp (kPa) G 300 kg m ⫺2 s ⫺1 (E) (F)
0.7 1.0 1.5 2.0
7.1 6.5 6.1 5.9
33.3 28.9 27.9 25.8
7.7 7.8 7.8 7.9
arrangements (E) and (F) give the lowest exit quality. The results seem confusing since one would expect countercross arrangement would have a better performance than other arrangements. To explain this phenomenon, one may have to examine the detailed variations of the wall temperature along the condenser as illustrated in Fig. 4 (arrangements A and E). For G 100 kg m ⫺2 s ⫺1, one can clearly see that arrangement (A) is superior to arrangement (E) since the wall temperature drop for arrangement (A) occurs more quickly than arrangement (E). For instance, the wall temperature for arrangement (A) is approximately 7⬚C lower than that of arrangement (E) near z 4 m. The result substantiates that the counter-cross flow may have a higher heat transfer performance than the z-shape arrangement. However, the wall temperature for arrangement (A) does not consistently decrease along the condenser. Actually, the temperature may even increase when z ⬎ 10 m at G 100 kg m ⫺2 s ⫺1. Notice that the z-shape arrangement (E) does not show this phenomenon. For G 200 or
32.3 27.4 26.5 25.3
300 kg m ⫺2 s ⫺1, the test results also report similar phenomena. The wall temperatures for arrangement (A) drop more quickly than arrangement (E) but reaches an asymptotic value which is higher than that of arrangement (E) at the exit. Explanation of this phenomenon for arrangement (A) is because the temperature at the inlet portion (row 2, upper side) is much higher than that at the exit (row 1, upper side). As a result, larger temperature difference between the adjacent tubes may significantly increase the contribution of the heat conduction along the fin which is transferred back to the refrigerant. For the z-shape arrangement (E), the temperature difference between the neighboring tube is relatively small, therefore reversed heat transfer by conduction is negligible. The test results suggest that the counter-cross flow arrangement may not achieve the highest subcooling temperature. This may significantly reduce the system performance. To eliminate the reversed conduction effect and improve the performance of the counter-cross flow arrangement, possible solutions for improving the performance of the counter-cross arrangement are shown in Fig. 5. Fig. 5a has removed several redundant tubes at the exit. This may be very helpful from both heat transfer and economical point of view. Fig. 5b has used two separate one rows to form a two-row condenser. The tiny gap between these two rows may inhibit the heat conduction from the hot inlet portion back to the cold exit region. 3.2. Results of two-circuit arrangement
Fig. 4. Wall temperature variations for arrangements (A) and (E). Fig. 4. Variations des tempe´ratures des parois pour des dispositions (A) a` (E).
The previous results were for one-circuit arrangement. For two-circuit arrangement, refrigerant flow in each circuit may not be the same. Fig. 6 shows the wall temperature distributions of arrangement (G) for Gin 100, 200, and 300 kg m ⫺2 s ⫺1 at Vfr 1.0 m s ⫺1. It is interesting to see that significant differences in the wall temperature exist in circuit 1 and circuit 2. This is probably owing to the maldistribution of refrigerant flowrate. Note that the tube length for each circuit is the same, and the arrangement of these two refrigerant circuits is in a symmetrical style. Further, gravity effects would result in different pressure drops, even if the flowrates were equal. For the adiabatic test condition, it is obvious the refrigerant flowrate for both circuit should
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Fig. 5. Possible improvements for counter-cross arrangement. Fig. 5. Disposition a` contre-courant: ame´liorations possibles.
be the same. For the same axial distance and one of the circuits is fully condensed, the wall temperature for circuit 1 is higher than that of circuit 2. The test results imply that the mass flowrate for circuit 1 is higher than that of circuit 2. The difference in the mass flowrate between circuit 1 and 2 may be related to the gravity forces. As shown by Collier and Thome [5], the gravitational pressure gradient in the two-phase region is composed of vapor-phase contribution, g sin uar g, and liquid-phase contribution, g sin u (1 ⫺ a )r f,
Fig. 6. Wall temperature variations for arrangement (G). Fig. 6. Variations des tempe´ratures des parois pour la disposition (G).
i.e. ÿ
j k dp=dz g gsinu arg ⫹ 1 ⫺ arf :
1
For practical HVAC and R application, the liquid density is much higher than that of the vapor density. For an aircooled refrigerant condenser at high quality region, the flow pattern is annular at the initial stage of condensation (see the two-phase visualization of flow pattern by Wang et al. [6]). Thus the void fraction is small at this stage, therefore the gravity term is small compared to the total pressure gradient. For G 200 kg m ⫺2 s ⫺1, in the “up-portion” of circuit 1, the gravitational pressure drop is only about 4% ⬃ 5% of the total pressure drop in the “up-portion”. However, in the “down-portion” of circuit 1, the effect of gravity is reversed. Further, since the quantity of the liquid condensate in this portion is much larger than that of the “up-portion” in circuit 1. Hence the “pressure-gain” in this portion is very pronounced. An estimation of the “pressure-gain” caused by the gravitational gradient in the “down-portion” of circuit 1 is approximately 30% ⬃ 40% of the total gradient. Converse to refrigerant circuit 1, the sign of the gravitational gradient for circuit 2 in the “up-portion” is different from circuit 1 and is regarded as “pressure loss”. However, the total pressure drop is the same for circuits 1 and 2 after combination. Accordingly, circuit 1 is likely to have more mass flowrate in order to have a higher pressure drop. The test results by Kaga et al. [7] had also reported a similar characteristics for the two-circuit arrangement. This maldistribution of refrigerant flowrate may become more pronounced for lower mass flowrates. This is because the contribution of gravity force may become more significant.
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281 ⫺1
hereafter to 14 kPa for Vfr 0.4 m s . Analogous phenomenon was observed for Gin 200 kg m ⫺2 s ⫺1 near Vfr 0.4 m s ⫺1. Unlike the test results of the two-circuit arrangement, the one-circuit arrangements (A–F) do not reveal such kinds of characteristics as seen in Fig. 8. Possible explanation of the sudden drop of pressure drop is as follows. As known, the total pressure gradient within the condenser comprises acceleration loss, friction loss, gravity loss, and bend loss. i.e. Dp Dp Dp Dp Dp ⫹ ⫹ ⫹ : 2 Dz Dz a Dz f Dz g Dz b
Fig. 7. Pressure drops for arrangements (G) and (H). Fig. 7. Pertes de pression pour les dispositions (G) et (H).
3.3. Results of pressure drops Fig. 7 shows the pressure drops on the refrigerant-side vs. frontal velocity for arrangements G and H. A unique characteristics of this graph is that refrigerant pressure drop experiences a minimum value near the complete condensation point. For example, Gin 300 kg m ⫺2 s ⫺1, the pressure drop is approximately 12.8 kPa at Vfr 2.0 m s ⫺1 for arrangement (H). The pressure drop slightly increases to 13.1 kPa for Vfr 1.5 m s ⫺1. Notice that this value drops significantly to 9.2 kPa for Vfr 1.1 m s ⫺1 and increases
Of these four terms, the wall friction and the return bend loss Dpf and Dpb are always negative along the direction of refrigerant flow. The gravity loss, Dpg, is positive when refrigerant flow in the hairpin is moving downwards. The acceleration loss, Dpa, is negative for evaporation or boiling. However, for two-phase condensing flow, Dpa is negative since the quality is decreasing along the tube length. As a result, Dpa is in fact a “pressure gain” term. In practical airconditioning and refrigeration application, the frictional pressure drop Dpf would constitute most of the two-phase pressure drop. Jung and Radermacher [8] estimated the contribution of the acceleration term Dpa is always less than 10%. For the one-circuit arrangement (A–F), the contribution of this term is very small. For the present two-circuit arrangement, mal-distribution in the circuitry is likely as mentioned in Section 3.2. For some circumstances, it is very likely that one refrigerant circuit is completely condensed (subcooled) while the other circuit is still in the two-phase region. When these two circuits are combined, it is expected that direct-contact condensation would occur. This would significantly improve the heat transfer characteristics and may significantly increase the pressure gain of Dpa. Consequently, the pressure drop is suddenly decreased. Examinations of the present data indicate that the sudden decrease of pressure drop occurs near the point where average vapor quality is near zero. This further suggests that this phenomenon occurs only for one circuit in the saturated region while the other one is in the subcooled region. Note that this phenomenon is seen only for the two-circuit arrangement. The one-circuit design does not experience the sudden drop of pressure drop as seen in Fig. 8. 4. Conclusions An experimental study was carried out to investigate the effect of circuitry on the performance of a wavy finned condenser. A total of eight arrangements were made and tested. The arrangements included six 1-circuit and two 2circuit arrangements. On the basis of previous discussions, the following conclusions are made:
Fig. 8. Pressure drops for arrangements (A), (C), and (E). Fig. 8. Pertes de pression pour les dispositions (A), (C) et (E).
• The counter-cross arrangement gives better performance. However, the reversed heat conduction from
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the inlet portion to the exit portion may offset the benefit of counter-cross arrangement. Two modifications to improve the counter-cross arrangement are proposed in this study. • For the two-circuit arrangement, mal-distribution of the refrigerant flow is observed. The mal-distribution of the refrigerant flow is strongly influenced by gravitational force. • For the two-circuit arrangement, an unusual characteristics of the refrigerant pressure drop is observed. The refrigerant pressure drop experiences a sudden drop near x 0. This is because direct condensation occurs when these two circuits combine.
Acknowledgements The authors would like to express gratitude for the Energy R and D foundation funding from the Energy Commission of the Ministry of Economic Affairs, which provides financial supports of the current study. Suggestions and corrections by Prof. Ralph Webb is highly appreciated.
References [1] Hogan MR. The development of a low-temperature heat pump grain dryer. Ph.D. Thesis, Purdue University, 1980:116–127. [2] ANSI/ASHRAE standard 37-1988, Methods of testing for rating unitary air-conditioning and heat pump equipment. [3] ASHRAE, ASHRAE standard 41.2-1987 standard methods for laboratory air-flow measurement, 1987. [4] REFPROP 6.0, National Institute of Standards and Technology, Gaithersburg, MD, 1998. [5] Collier JG, Thome JR. Convective boiling and condensation. 3rd ed.. Oxford Science Publications, 1994:48. [6] Wang CC, Chiang CS, Lu DC. Visual observation of flow pattern of R-22, R-134a, and R-407C in a 6.5 mm smooth tube. Experimental Thermal and Fluid Science 1997; 15(4):395–405. [7] Kaga K, Yamada K, Takeshita S, Yamanaka MG. Improvement of a capacity of plate fin tubed heat exchanger by thermal analysis considering thermal interaction between pipes by heat conduction in fins. Proc. 10th Int. Heat Transfer Conf., Brayton, UK, The industrial Sections Paper (1994) paper no. I/2-CHE-8 99-104. [8] Jung DS, Radermacher R. Prediction of pressure drop during horizontal annular flow boiling of pure and mixed refrigerants. Int Heat Mass Transfer 1989;32(12):2435–2446.