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Experimental study of capillary-assisted water evaporators for vapour-absorption systems
Applied Energy 71 (2002) 45–57 www.elsevier.com/locate/apenergy
Experimental study of capillary-assisted water evaporators for vapour-absorption systems H.M. Sabir*, A.C. Bwalya School of Engineering, Kingston University, London SW15 3DW, UK Received 23 July 2001; received in revised form 27 September 2001; accepted 29 September 2001
Abstract Direct exchange (DX) water evaporators with internal capillary structure have been developed and tested. The performances of three evaporators, two with internal open grooves (IOG) and one with a layer of sintered metal powder, are presented for various air velocities and temperatures. The internally powder coated (IPC) evaporator was found to offer the best performance, and achieved an average evaporation capacity of 700 W. The two IOG evaporators achieved average evaporation capacities of 500 and 300 W, for evaporators having deep and shallow grooves (SG), respectively. The IPC evaporator also had higher boiling and overall heat transfer coefficients compared with the two IOG evaporators. # 2002 Elsevier Science Ltd. All rights reserved.
1. Introduction Vapour absorption cycle refrigerators, powered by low-grade heat from domestic waste incinerators, combined heat-and-power systems or solar energy, potentially offer a more environmentally friendly alternative way, in terms of CO2 emissions, for cooling buildings, compared with vapour-compression systems powered by electricity from the mains supply. In addition, vapour-absorption systems are capable of using water as their refrigerant which also has a number of environmental advantages. The commercial application of absorption-cycle machines for building cooling is resisted because of their high capital cost when compared with electric vapour-compression machines. To increase the use of vapour-absorption cycle refrigeration in building Abbreviations: CA, capillary assisted; DG, deep grooved; DX, direct exchange; IOG, internally open grooved; IPC, internally powder coated; LMTD, log mean temperature difference * Corresponding author. Tel.: +44-20-8547-2000; fax: +44-20-8547-7992. E-mail address: [email protected] (H.M. Sabir). 0306-2619/02/$ - see front matter # 2002 Elsevier Science Ltd. All rights reserved. PII: S0306-2619(01)00042-3
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H.M. Sabir, A.C. Bwalya / Applied Energy 71 (2002) 45–57
Nomenclature a Cp d Di g hfg Ja K : m: e Qe q00 qmax re T t
Groove width (m) Specific heat capacity (kJ kg1 K1) Pore diameter (m) Internal tube diameter (m) Gravitational acceleration (m s2) Latent heat of vaporisation (kJ kg1) Jacob number Permeability of porous matrix (m2) Evaporation rate (kg s1) Evaporator capacity (W) Boiling heat flux (W m2) Maximum heat transfer rate (W) Effective radius (m) Temperature (K) Groove depth (m)
Greeks T "
Surface tension (N m1) Temperature difference (K) Porosity Dynamic viscosity (N s m2) Density (kg m3) Thickness of porous layer (m)
Subscripts g Vapour l Liquid sat Saturation condition
cooling systems it will be necessary to reduce their capital costs and, at the same time, improve their thermal efficiencies. In some applications, one way to do this would be to replace the conventional water evaporator with a direct heat-exchange (DX) unit. In the conventional water evaporator, used in vapour absorption cycles, refrigerant water is sprayed over the outside of a coil through which chilled water flows. This water is then circulated through fan-coil units, which, in turn, cool the air in a building. This requires two heat exchange processes, in two separate heat exchangers, and is consequently inefficient and expensive. Instead, it is proposed here to cool the air directly using a DX water evaporator. This requires only one heat-exchange process and will, therefore, be less expensive and more thermally efficient. Water has a high saturation-temperature compared with other refrigerants. An evaporator of, say 1 m in height, of the type used in conventional electric vapour-
H.M. Sabir, A.C. Bwalya / Applied Energy 71 (2002) 45–57
47
compression systems filled with refrigerant water would not be practicable for these reasons. The hydrostatic pressure created by the head of water would suppress boiling until its temperature reaches its local equivalent saturation temperature as determined by the static-head pressure. In this case, with 1 m of water head, the water contained in the bottom of the evaporator would not boil until it reached at least 46 C, which is clearly too high for air-conditioning purposes. An alternative way of producing water vapour within a DX coil is to cause the liquid to flow only over the inner surface of a tube in an evaporating film, leaving the centre of the tube open for the passage of vapour. The cooled air would flow across the outside of the evaporator coil exchanging heat directly with the refrigerant water. One way to achieve this effect is to utilise capillary action to pump a liquid film over the internal surface of the tube. The capillary action can be achieved by, for example, open channels or a layer of absorbing powder. These two capillary designs are discussed in this report.
2. Review of previous work The use of capillary effects is well established in the field of heat pipes [1]. However, the technology is just beginning to be applied to water film evaporators. In relatively recent reports [2,3], capillary effects were utilised through the use of a cloth wick, which was held against the inner surface of the evaporator tube by a metal wire mesh, rolled into a cylinder. The mesh kept the cloth pressed against the inner tube wall, while allowing the evolving vapour to escape to the centre of the tube. The researchers reported an overall heat-transfer coefficient of between 59 Wm2 K1 [2] and 54 Wm2 K1 [3] for bare evaporators tubes. Their evaporator proved the concept, but the design appeared to suffer from a number of mechanical and thermal deficiencies, including poor thermal-contact between wick and tube, non-optimised thermal properties of the wick and possible deterioration of the wick structure under long and constant exposure to moisture.
3. Evaporators with internal capillary structure This report presents the experimental performance data of three capillary-assisted (CA) evaporators, two with open grooves and one with a sintered layer of metal powder. The grooves and the powder layer utilise capillary effects to create a liquid film over the internal surface of the evaporator tube. The evaporators differ from other CA evaporators in that their capillary structure forms an integral part of the internal surface of their tubes. 3.1. Open grooved evaporators The two internally open-grooved (IOG) evaporators had open, circumferential, Vshaped grooves engraved in the internal surface of their tubes. Both were produced
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from copper tubes of external diameter of 28.6 mm (1 18 inch) using a specially designed tap. The first, deep grooved (DG) evaporator was made out of thick walled copper tube (of smaller internal diameter), which resulted in grooves 0.42 mm deep and 0.31 mm wide. The second, shallow grooved (SG) evaporator was made out of thinner walled tubes (of larger internal diameter), resulting in shallower and narrower grooves of 0.32 mm depth and 0.23 mm width. The two evaporators had the same number of grooves, i.e. 2000 grooves per metre (0.5 mm pitch). However, the wider channels of DG meant that more of its internal surface was covered by liquid compared with SG. The capillary action (maximum head, flow rate, etc.) depends strongly on the effective radius re, which for open channels is given by re ¼ 2
flow area wetted perimeter
ð1Þ
Applying Eq. (1) to the geometry of V-shaped channels, one can derive re as: re ¼
at qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi a þ 2 t2 þ ða=2Þ2
ð2Þ
where a and t are the groove’s width and depth, respectively. Calculated from Eq. (2), the effective radii of DG and SG are 0.107 and 0.082 mm, respectively. The maximum evaporative power is achieved when the evaporation rate equals the maximum flow rate of liquid that can be supported by the capillary action. This rate is given by: r4e hfg 2 qmax ¼ Di g ð3Þ 4Di re which is a form of the Hagen–Poiseulle equation [4]. The term in brackets is the difference between the capillary lift and the static liquid-head. 3.2. Powder coated evaporator The internally powder coated (IPC) evaporator had a 0.5 mm thick layer of porous copper powder sintered to the internal surface of its tube. The tube had the same external diameter as the IOG evaporators, i.e. 28.6 mm. The effective radius here is equal to the average pore size, which is a function of the particle size of the powder, its material and the sintering process. The average pore (= re) and particle sizes of the powder were 20 and 50 mm, respectively. The boiling heat flux for the powder layer can be calculated from the expression derived by Madhusudana Rao and Balakrishnan [5], which assumes a close-packed arrangement of spherical metal powder particles of uniform diameter
H.M. Sabir, A.C. Bwalya / Applied Energy 71 (2002) 45–57
00
q ¼ 2:4 10
4
l g 2 Kh2fg ð=dÞ0:8 Tsat Ja0:22 g Tsat ðl g Þ"1:23
where Ja is the Jacob number=
49
ð4Þ
l CpTsat g hfg
4. Effective radius significance The performance of CA evaporators is a strong function of re. The smaller the effective radius, the higher the capillary lift and the more the coverage of the surface. However, for a given tube diameter (i.e. the required capillary lift), the associated effective radius should not be decreased further than is necessary to affect such a lift, as this, apart from unnecessarily increasing the lift, will increase resistance to liquid flow and decrease the flow rate. The latter effect limits the cooling capacity of the evaporator.
5. Evaporator design Each of the three evaporators was made of nine tubes connected in series and in such a way to allow the vapour to escape from each tube to two risers (one on each side), while allowing the un-evaporated liquid to flow through the entire length, (Fig. 1). The effective length of each evaporator was 2.79 m, giving an external surface area of 0.25 m2.
6. Experimental rig The test evaporator was fitted to a single effect lithium bromide/water (LiBr/H2O) refrigeration system. The evaporator load was provided by air passing across the evaporator tubes, which were placed inside a wind tunnel, through which the air flowed. The temperature and velocity of the air were controlled by a heat-exchanger/ water-boiler combination. The combination provided heating and circulation of the load air in the following manner. Hot water, produced in the boiler, was circulated to the heat exchanger, where its heat is transferred to the air. The air temperature was controlled by varying the input heat to the water boiler and by controlling the circulation rate of the water. The velocity of the air was controlled through varying the setting of the air fan of the heat exchanger. The rig is shown schematically in Fig. 2. The refrigerant (liquid water) entered the evaporator through a throttle valve. Some of the water evaporated and the remainder passed to a holding tank before it : was circulated to the absorber. The evaporation rate, me , was obtained by measuring the difference in mass flow rates of those for the liquid water at inlet and outlet from the evaporator. The evaporator capacity was calculated from:
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Fig. 1. Schematic representation of the evaporator and wind tunnel.
: : Q e ¼ me hfg
ð5Þ
where hfg is the latent heat of vaporisation at the evaporator pressure. Eq. (5) implies that the liquid enters the evaporator at saturated conditions. This might not be entirely accurate. However, the experimental conditions were such that the difference between the condensing and evaporation temperatures was of the order of 10– 12 C. Under these conditions, the error resulting from using Eq. (5) was of the order of 2%.
7. Results and discussion : The performance is presented in terms of the evaporation capacity, Qe , boiling heat-transfer coefficient, and overall heat-transfer coefficient. The boiling heat-
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Fig. 2. Schematic diagram of the complete test rig.
: transfer coefficient is calculated from Qe and the difference between the evaporator temperature (taken to be that of the liquid refrigerant leaving the evaporator) and the temperature of the outside surface of the evaporator tube. The latter was uniform longitudinally and circumferentially for the IPC evaporator. However, for IOG evaporators, there were incidences where the temperature of the top of the tube was slightly higher than the rest of the circumference, which indicated the inability of the capillary effect to pull the liquid to the very top. Nonetheless, this was a rare occurrence and circumferential temperature uniformity was assumed throughout. This uniformity, and the constant evaporator temperature, allowed the use of a straight temperature-difference between tube and refrigerant for the calculation. The : overall heat transfer coefficient was calculated from Qe and the log mean temperature-difference (LMTD) between the evaporator temperature and that of the outside air. Both transfer coefficients were based on the outside surface-area of the tube. The performance data of the evaporator are presented in Figs. 3–11. The performance is shown against varying air velocity at three levels of air inlet (air-on) temperature; low, medium and high. Figs. 3–5 show that the evaporators’ capacity increases with air velocity. This is due to the improvement of heat transfer with increased velocity and mixing. The figures also reveal that the IPC outperformed both IOG evaporators, for all temperature levels, by quite a margin. The enhancement is attributed to factors relating to the IPC capillary structure, namely:
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Fig. 3. Comparison of evaporator’s capacities at low air temperatures (20–30 C).
Fig. 4. Comparison of evaporator’s capacities at medium air temperatures (30–40 C).
. The smaller pore size (0.02 mm) ensured a better capillary lift and surface coverage by liquid. The sheer number of pores helped decrease the resistance to liquid flow, which might have resulted from the small pore size. The net effect of the two factors was to enhance the evaporation power. . The pores also offered a large number of natural ebullition sites for boiling, thus enhancing boiling heat-transfer.
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Fig. 5. Comparison of evaporator’s capacities at high air temperatures (40–50 C).
Fig. 6. Comparison of boiling heat-transfer coefficients at low air temperatures (20–30 C).
The figures also show that the DG is slightly better than the SG. Again, this is attributed to the better coverage of liquid due to wider channels. It is not clear which of the two would have the larger mass flow rate, because, on the one hand, the larger re of the DG is expected to result in a relatively poorer lift which reduces the mass flow rate. On the other hand, its wider channels will decrease the resistance to fluid flow and increase the flow rate. In any case, it is reasonable to assume that there is an optimal combination of these two effects which results in maximum performance.
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The IPC evaporator achieved an average capacity of about 700, compared with 500 and 300 W for the DG and SG, respectively. Figs. 6–8 show the effects of air velocity on the boiling heat-transfer coefficient. Here, the figures, except Fig. 8, show little dependency of the boiling coefficient on outside (air) conditions. The figures reveal that the IPC again outperforms the two IOG evaporators. The picture becomes unclear when comparing the performances of DG and SG. The crossing of the SG and DG lines in Figs. 6 and 7 seems to support the argument, presented earlier, that the better evaporation rate of DG is due to better coverage (i.e. larger exchange area) rather than increased mass flow rate. This also suggests that the similarity of capillary structure of SG and DG results in a more or less similar boiling heat-transfer coefficient of an average value of about 1000 W m2 K1. In comparison, the IPC evaporator achieved an average boiling heat-transfer coefficient of about 2750 W m2 K1. The downward trend of IPC and SG lines in Fig. 8 is somewhat puzzling. However, a combination of less than a sufficient number of tests (SG) and under-performing absorber (due to air entrainment—a constant risk in systems operating at sub-atmospheric pressures) may explain this behaviour. It is also possible that the IPC and SG are more sensitive to high surface temperatures induced by high air temperatures. Figs. 9–11 present the overall heat-transfer coefficient as a function of air velocity. The figures show that the overall coefficient increases with increasing velocity. The trend is to be expected, since the overall coefficient is more akin to the forced convection heat-transfer coefficient on the air side, where the larger thermal resistance exists. The latter is of course a strong function of air velocity. The figures, again, reveal the relative superiority of the IPC’s overall coefficient. The DG also has a slightly better overall coefficient compared with SG because of its better evaporation capacity. The
Fig. 7. Comparison of boiling heat-transfer coefficients at medium air temperatures (30–40 C).
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figures also show the relative independence of the overall heat transfer coefficient from air temperature. The average values of the overall heat-transfer coefficients are estimated to be about 250, 150 and 130 W m2 K1 for the IPC, DG and SG, respectively. It must be emphasised that the average values of capacity, boiling and overall heat-transfer coefficients are rather approximate. Their real strength lies in their relative values, which favours the IPC evaporator over the IOG ones.
Fig. 8. Comparison of boiling heat-transfer coefficients at high air-temperatures (40–50 C).
Fig. 9. Comparison of overall heat-transfer coefficients at low air-temperatures (20–30 C).
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Fig. 10. Comparison of overall heat-transfer coefficients at medium air temperatures (30–40 C).
Fig. 11. Comparison of overall heat-transfer coefficients at high air temperatures (40–50 C).
8. Conclusion Capillary-assisted, direct exchange evaporators have been developed to be used in conjunction with vapour-absorption refrigeration systems. The new evaporators are expected to reduce capital costs and improve thermal performances of air conditioning vapour-absorption systems.
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The IPC evaporator offers a better performance compared with both IOG evaporators. The improved performance is due to its internal capillary structure, which improves mass flow rate and the boiling process. IPC evaporator achieved an average evaporation capacity of about 700 W resulting in a heat flux of 2.8 kW m2 at the tested air velocities and temperatures, while the DG and SG evaporators achieved average evaporation capacities of about 500 and 300 W resulting in heat fluxes of 1.6 and 1.2 kW m2, respectively, at the same test conditions. The IPC evaporator had respective average boiling and overall heat-transfer coefficients of about 2750 and 250 W m2 K1. Both IOG evaporators achieved average boiling heat-transfer coefficients of about 1000 W m2 K1. DG and SG evaporators had average values of overall heat-transfer coefficients of about 150 and 130 W m2 K1, respectively. Further improvements of performance can be attained by using finned tubes instead of the bare tubes used in this research work. Nevertheless, it has been shown that the evaporators achieved performance levels so encouraging their commercial exploitation.
Acknowledgements The authors acknowledge the support of the Engineering and Physical Sciences Research Council (EPSRC), UK, which funded this work.
References [1] Dunn PD, Reay DA. Heat pipes. UK: Pergammon, 1994. [2] Sabir HM, Eames, I.E. Water film evaporator for air conditioning of buildings, Proceedings of CIBSE’97 conference, vol. 1. London, 1997. pp. 292–8. [3] Eames, IW, Wu, S, Sabir, H. Direct-heat exchange water evaporator for vapour-absorption cycle refrigerators, Proceedings of the IIR/IIF conference, Oslo, 1998. [4] Faghri A. Heat pipe science and technology. Washington, DC: Taylor & Francis, 1995. [5] Madhusudana Rao S, Balakrishnan AR. Analysis of pool-boiling heat transfer over porous surfaces. Journal of Heat and Mass Transfer 1997;32:463–9.
Experimental study of capillary-assisted water evaporators for vapour-absorption systems H.M. Sabir*, A.C. Bwalya School of Engineering, Kingston University, London SW15 3DW, UK Received 23 July 2001; received in revised form 27 September 2001; accepted 29 September 2001
Abstract Direct exchange (DX) water evaporators with internal capillary structure have been developed and tested. The performances of three evaporators, two with internal open grooves (IOG) and one with a layer of sintered metal powder, are presented for various air velocities and temperatures. The internally powder coated (IPC) evaporator was found to offer the best performance, and achieved an average evaporation capacity of 700 W. The two IOG evaporators achieved average evaporation capacities of 500 and 300 W, for evaporators having deep and shallow grooves (SG), respectively. The IPC evaporator also had higher boiling and overall heat transfer coefficients compared with the two IOG evaporators. # 2002 Elsevier Science Ltd. All rights reserved.
1. Introduction Vapour absorption cycle refrigerators, powered by low-grade heat from domestic waste incinerators, combined heat-and-power systems or solar energy, potentially offer a more environmentally friendly alternative way, in terms of CO2 emissions, for cooling buildings, compared with vapour-compression systems powered by electricity from the mains supply. In addition, vapour-absorption systems are capable of using water as their refrigerant which also has a number of environmental advantages. The commercial application of absorption-cycle machines for building cooling is resisted because of their high capital cost when compared with electric vapour-compression machines. To increase the use of vapour-absorption cycle refrigeration in building Abbreviations: CA, capillary assisted; DG, deep grooved; DX, direct exchange; IOG, internally open grooved; IPC, internally powder coated; LMTD, log mean temperature difference * Corresponding author. Tel.: +44-20-8547-2000; fax: +44-20-8547-7992. E-mail address: [email protected] (H.M. Sabir). 0306-2619/02/$ - see front matter # 2002 Elsevier Science Ltd. All rights reserved. PII: S0306-2619(01)00042-3
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H.M. Sabir, A.C. Bwalya / Applied Energy 71 (2002) 45–57
Nomenclature a Cp d Di g hfg Ja K : m: e Qe q00 qmax re T t
Groove width (m) Specific heat capacity (kJ kg1 K1) Pore diameter (m) Internal tube diameter (m) Gravitational acceleration (m s2) Latent heat of vaporisation (kJ kg1) Jacob number Permeability of porous matrix (m2) Evaporation rate (kg s1) Evaporator capacity (W) Boiling heat flux (W m2) Maximum heat transfer rate (W) Effective radius (m) Temperature (K) Groove depth (m)
Greeks T "
Surface tension (N m1) Temperature difference (K) Porosity Dynamic viscosity (N s m2) Density (kg m3) Thickness of porous layer (m)
Subscripts g Vapour l Liquid sat Saturation condition
cooling systems it will be necessary to reduce their capital costs and, at the same time, improve their thermal efficiencies. In some applications, one way to do this would be to replace the conventional water evaporator with a direct heat-exchange (DX) unit. In the conventional water evaporator, used in vapour absorption cycles, refrigerant water is sprayed over the outside of a coil through which chilled water flows. This water is then circulated through fan-coil units, which, in turn, cool the air in a building. This requires two heat exchange processes, in two separate heat exchangers, and is consequently inefficient and expensive. Instead, it is proposed here to cool the air directly using a DX water evaporator. This requires only one heat-exchange process and will, therefore, be less expensive and more thermally efficient. Water has a high saturation-temperature compared with other refrigerants. An evaporator of, say 1 m in height, of the type used in conventional electric vapour-
H.M. Sabir, A.C. Bwalya / Applied Energy 71 (2002) 45–57
47
compression systems filled with refrigerant water would not be practicable for these reasons. The hydrostatic pressure created by the head of water would suppress boiling until its temperature reaches its local equivalent saturation temperature as determined by the static-head pressure. In this case, with 1 m of water head, the water contained in the bottom of the evaporator would not boil until it reached at least 46 C, which is clearly too high for air-conditioning purposes. An alternative way of producing water vapour within a DX coil is to cause the liquid to flow only over the inner surface of a tube in an evaporating film, leaving the centre of the tube open for the passage of vapour. The cooled air would flow across the outside of the evaporator coil exchanging heat directly with the refrigerant water. One way to achieve this effect is to utilise capillary action to pump a liquid film over the internal surface of the tube. The capillary action can be achieved by, for example, open channels or a layer of absorbing powder. These two capillary designs are discussed in this report.
2. Review of previous work The use of capillary effects is well established in the field of heat pipes [1]. However, the technology is just beginning to be applied to water film evaporators. In relatively recent reports [2,3], capillary effects were utilised through the use of a cloth wick, which was held against the inner surface of the evaporator tube by a metal wire mesh, rolled into a cylinder. The mesh kept the cloth pressed against the inner tube wall, while allowing the evolving vapour to escape to the centre of the tube. The researchers reported an overall heat-transfer coefficient of between 59 Wm2 K1 [2] and 54 Wm2 K1 [3] for bare evaporators tubes. Their evaporator proved the concept, but the design appeared to suffer from a number of mechanical and thermal deficiencies, including poor thermal-contact between wick and tube, non-optimised thermal properties of the wick and possible deterioration of the wick structure under long and constant exposure to moisture.
3. Evaporators with internal capillary structure This report presents the experimental performance data of three capillary-assisted (CA) evaporators, two with open grooves and one with a sintered layer of metal powder. The grooves and the powder layer utilise capillary effects to create a liquid film over the internal surface of the evaporator tube. The evaporators differ from other CA evaporators in that their capillary structure forms an integral part of the internal surface of their tubes. 3.1. Open grooved evaporators The two internally open-grooved (IOG) evaporators had open, circumferential, Vshaped grooves engraved in the internal surface of their tubes. Both were produced
48
H.M. Sabir, A.C. Bwalya / Applied Energy 71 (2002) 45–57
from copper tubes of external diameter of 28.6 mm (1 18 inch) using a specially designed tap. The first, deep grooved (DG) evaporator was made out of thick walled copper tube (of smaller internal diameter), which resulted in grooves 0.42 mm deep and 0.31 mm wide. The second, shallow grooved (SG) evaporator was made out of thinner walled tubes (of larger internal diameter), resulting in shallower and narrower grooves of 0.32 mm depth and 0.23 mm width. The two evaporators had the same number of grooves, i.e. 2000 grooves per metre (0.5 mm pitch). However, the wider channels of DG meant that more of its internal surface was covered by liquid compared with SG. The capillary action (maximum head, flow rate, etc.) depends strongly on the effective radius re, which for open channels is given by re ¼ 2
flow area wetted perimeter
ð1Þ
Applying Eq. (1) to the geometry of V-shaped channels, one can derive re as: re ¼
at qffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffiffi a þ 2 t2 þ ða=2Þ2
ð2Þ
where a and t are the groove’s width and depth, respectively. Calculated from Eq. (2), the effective radii of DG and SG are 0.107 and 0.082 mm, respectively. The maximum evaporative power is achieved when the evaporation rate equals the maximum flow rate of liquid that can be supported by the capillary action. This rate is given by: r4e hfg 2 qmax ¼ Di g ð3Þ 4Di re which is a form of the Hagen–Poiseulle equation [4]. The term in brackets is the difference between the capillary lift and the static liquid-head. 3.2. Powder coated evaporator The internally powder coated (IPC) evaporator had a 0.5 mm thick layer of porous copper powder sintered to the internal surface of its tube. The tube had the same external diameter as the IOG evaporators, i.e. 28.6 mm. The effective radius here is equal to the average pore size, which is a function of the particle size of the powder, its material and the sintering process. The average pore (= re) and particle sizes of the powder were 20 and 50 mm, respectively. The boiling heat flux for the powder layer can be calculated from the expression derived by Madhusudana Rao and Balakrishnan [5], which assumes a close-packed arrangement of spherical metal powder particles of uniform diameter
H.M. Sabir, A.C. Bwalya / Applied Energy 71 (2002) 45–57
00
q ¼ 2:4 10
4
l g 2 Kh2fg ð=dÞ0:8 Tsat Ja0:22 g Tsat ðl g Þ"1:23
where Ja is the Jacob number=
49
ð4Þ
l CpTsat g hfg
4. Effective radius significance The performance of CA evaporators is a strong function of re. The smaller the effective radius, the higher the capillary lift and the more the coverage of the surface. However, for a given tube diameter (i.e. the required capillary lift), the associated effective radius should not be decreased further than is necessary to affect such a lift, as this, apart from unnecessarily increasing the lift, will increase resistance to liquid flow and decrease the flow rate. The latter effect limits the cooling capacity of the evaporator.
5. Evaporator design Each of the three evaporators was made of nine tubes connected in series and in such a way to allow the vapour to escape from each tube to two risers (one on each side), while allowing the un-evaporated liquid to flow through the entire length, (Fig. 1). The effective length of each evaporator was 2.79 m, giving an external surface area of 0.25 m2.
6. Experimental rig The test evaporator was fitted to a single effect lithium bromide/water (LiBr/H2O) refrigeration system. The evaporator load was provided by air passing across the evaporator tubes, which were placed inside a wind tunnel, through which the air flowed. The temperature and velocity of the air were controlled by a heat-exchanger/ water-boiler combination. The combination provided heating and circulation of the load air in the following manner. Hot water, produced in the boiler, was circulated to the heat exchanger, where its heat is transferred to the air. The air temperature was controlled by varying the input heat to the water boiler and by controlling the circulation rate of the water. The velocity of the air was controlled through varying the setting of the air fan of the heat exchanger. The rig is shown schematically in Fig. 2. The refrigerant (liquid water) entered the evaporator through a throttle valve. Some of the water evaporated and the remainder passed to a holding tank before it : was circulated to the absorber. The evaporation rate, me , was obtained by measuring the difference in mass flow rates of those for the liquid water at inlet and outlet from the evaporator. The evaporator capacity was calculated from:
50
H.M. Sabir, A.C. Bwalya / Applied Energy 71 (2002) 45–57
Fig. 1. Schematic representation of the evaporator and wind tunnel.
: : Q e ¼ me hfg
ð5Þ
where hfg is the latent heat of vaporisation at the evaporator pressure. Eq. (5) implies that the liquid enters the evaporator at saturated conditions. This might not be entirely accurate. However, the experimental conditions were such that the difference between the condensing and evaporation temperatures was of the order of 10– 12 C. Under these conditions, the error resulting from using Eq. (5) was of the order of 2%.
7. Results and discussion : The performance is presented in terms of the evaporation capacity, Qe , boiling heat-transfer coefficient, and overall heat-transfer coefficient. The boiling heat-
H.M. Sabir, A.C. Bwalya / Applied Energy 71 (2002) 45–57
51
Fig. 2. Schematic diagram of the complete test rig.
: transfer coefficient is calculated from Qe and the difference between the evaporator temperature (taken to be that of the liquid refrigerant leaving the evaporator) and the temperature of the outside surface of the evaporator tube. The latter was uniform longitudinally and circumferentially for the IPC evaporator. However, for IOG evaporators, there were incidences where the temperature of the top of the tube was slightly higher than the rest of the circumference, which indicated the inability of the capillary effect to pull the liquid to the very top. Nonetheless, this was a rare occurrence and circumferential temperature uniformity was assumed throughout. This uniformity, and the constant evaporator temperature, allowed the use of a straight temperature-difference between tube and refrigerant for the calculation. The : overall heat transfer coefficient was calculated from Qe and the log mean temperature-difference (LMTD) between the evaporator temperature and that of the outside air. Both transfer coefficients were based on the outside surface-area of the tube. The performance data of the evaporator are presented in Figs. 3–11. The performance is shown against varying air velocity at three levels of air inlet (air-on) temperature; low, medium and high. Figs. 3–5 show that the evaporators’ capacity increases with air velocity. This is due to the improvement of heat transfer with increased velocity and mixing. The figures also reveal that the IPC outperformed both IOG evaporators, for all temperature levels, by quite a margin. The enhancement is attributed to factors relating to the IPC capillary structure, namely:
52
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Fig. 3. Comparison of evaporator’s capacities at low air temperatures (20–30 C).
Fig. 4. Comparison of evaporator’s capacities at medium air temperatures (30–40 C).
. The smaller pore size (0.02 mm) ensured a better capillary lift and surface coverage by liquid. The sheer number of pores helped decrease the resistance to liquid flow, which might have resulted from the small pore size. The net effect of the two factors was to enhance the evaporation power. . The pores also offered a large number of natural ebullition sites for boiling, thus enhancing boiling heat-transfer.
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Fig. 5. Comparison of evaporator’s capacities at high air temperatures (40–50 C).
Fig. 6. Comparison of boiling heat-transfer coefficients at low air temperatures (20–30 C).
The figures also show that the DG is slightly better than the SG. Again, this is attributed to the better coverage of liquid due to wider channels. It is not clear which of the two would have the larger mass flow rate, because, on the one hand, the larger re of the DG is expected to result in a relatively poorer lift which reduces the mass flow rate. On the other hand, its wider channels will decrease the resistance to fluid flow and increase the flow rate. In any case, it is reasonable to assume that there is an optimal combination of these two effects which results in maximum performance.
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The IPC evaporator achieved an average capacity of about 700, compared with 500 and 300 W for the DG and SG, respectively. Figs. 6–8 show the effects of air velocity on the boiling heat-transfer coefficient. Here, the figures, except Fig. 8, show little dependency of the boiling coefficient on outside (air) conditions. The figures reveal that the IPC again outperforms the two IOG evaporators. The picture becomes unclear when comparing the performances of DG and SG. The crossing of the SG and DG lines in Figs. 6 and 7 seems to support the argument, presented earlier, that the better evaporation rate of DG is due to better coverage (i.e. larger exchange area) rather than increased mass flow rate. This also suggests that the similarity of capillary structure of SG and DG results in a more or less similar boiling heat-transfer coefficient of an average value of about 1000 W m2 K1. In comparison, the IPC evaporator achieved an average boiling heat-transfer coefficient of about 2750 W m2 K1. The downward trend of IPC and SG lines in Fig. 8 is somewhat puzzling. However, a combination of less than a sufficient number of tests (SG) and under-performing absorber (due to air entrainment—a constant risk in systems operating at sub-atmospheric pressures) may explain this behaviour. It is also possible that the IPC and SG are more sensitive to high surface temperatures induced by high air temperatures. Figs. 9–11 present the overall heat-transfer coefficient as a function of air velocity. The figures show that the overall coefficient increases with increasing velocity. The trend is to be expected, since the overall coefficient is more akin to the forced convection heat-transfer coefficient on the air side, where the larger thermal resistance exists. The latter is of course a strong function of air velocity. The figures, again, reveal the relative superiority of the IPC’s overall coefficient. The DG also has a slightly better overall coefficient compared with SG because of its better evaporation capacity. The
Fig. 7. Comparison of boiling heat-transfer coefficients at medium air temperatures (30–40 C).
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figures also show the relative independence of the overall heat transfer coefficient from air temperature. The average values of the overall heat-transfer coefficients are estimated to be about 250, 150 and 130 W m2 K1 for the IPC, DG and SG, respectively. It must be emphasised that the average values of capacity, boiling and overall heat-transfer coefficients are rather approximate. Their real strength lies in their relative values, which favours the IPC evaporator over the IOG ones.
Fig. 8. Comparison of boiling heat-transfer coefficients at high air-temperatures (40–50 C).
Fig. 9. Comparison of overall heat-transfer coefficients at low air-temperatures (20–30 C).
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Fig. 10. Comparison of overall heat-transfer coefficients at medium air temperatures (30–40 C).
Fig. 11. Comparison of overall heat-transfer coefficients at high air temperatures (40–50 C).
8. Conclusion Capillary-assisted, direct exchange evaporators have been developed to be used in conjunction with vapour-absorption refrigeration systems. The new evaporators are expected to reduce capital costs and improve thermal performances of air conditioning vapour-absorption systems.
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The IPC evaporator offers a better performance compared with both IOG evaporators. The improved performance is due to its internal capillary structure, which improves mass flow rate and the boiling process. IPC evaporator achieved an average evaporation capacity of about 700 W resulting in a heat flux of 2.8 kW m2 at the tested air velocities and temperatures, while the DG and SG evaporators achieved average evaporation capacities of about 500 and 300 W resulting in heat fluxes of 1.6 and 1.2 kW m2, respectively, at the same test conditions. The IPC evaporator had respective average boiling and overall heat-transfer coefficients of about 2750 and 250 W m2 K1. Both IOG evaporators achieved average boiling heat-transfer coefficients of about 1000 W m2 K1. DG and SG evaporators had average values of overall heat-transfer coefficients of about 150 and 130 W m2 K1, respectively. Further improvements of performance can be attained by using finned tubes instead of the bare tubes used in this research work. Nevertheless, it has been shown that the evaporators achieved performance levels so encouraging their commercial exploitation.
Acknowledgements The authors acknowledge the support of the Engineering and Physical Sciences Research Council (EPSRC), UK, which funded this work.
References [1] Dunn PD, Reay DA. Heat pipes. UK: Pergammon, 1994. [2] Sabir HM, Eames, I.E. Water film evaporator for air conditioning of buildings, Proceedings of CIBSE’97 conference, vol. 1. London, 1997. pp. 292–8. [3] Eames, IW, Wu, S, Sabir, H. Direct-heat exchange water evaporator for vapour-absorption cycle refrigerators, Proceedings of the IIR/IIF conference, Oslo, 1998. [4] Faghri A. Heat pipe science and technology. Washington, DC: Taylor & Francis, 1995. [5] Madhusudana Rao S, Balakrishnan AR. Analysis of pool-boiling heat transfer over porous surfaces. Journal of Heat and Mass Transfer 1997;32:463–9.